Handbook of Air Conditioning and Refrigeration, Second Edition

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HANDBOOK OF AIR CONDITIONING AND REFRIGERATION

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Shan K. Wang

Second Edition

McGraw-Hill New York San Francisco Washington, D.C. Auckland Bogotá Caracas Lisbon London Madrid Mexico City Milan Montreal New Delhi San Juan Singapore Sydney Tokyo Toronto

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Library of Congress Cataloging-in-Publication Data Wang, Shan K. (Shan Kuo) Handbook of air conditioning and refrigeration / Shan K. Wang — 2nd ed. p. cm. Includes index. ISBN 0-07-068167-8 1. Air conditioning. 2. Refrigeration and refrigerating machinery. I. Title. TH7687.W27 697.93 — dc21

2000

00-060576

McGraw-Hi l l Copyright © 2001, 1993 by The McGraw-Hill Companies, Inc. All rights reserved. Printed in the United States of America. Except as permitted under the United States Copyright Act of 1976, no part of this publication may be reproduced or distributed in any form or by any means, or stored in a data base or retrieval system, without the prior written permission of the publisher. 1 2 3 4 5 6 7 8 9 0

DOC/DOC

0 6 5 4 3 2 1 0

ISBN 0-07-068167-8 The sponsoring editor for this book was Linda Ludewig, the editing supervisor was David E. Fogarty, and the production supervisor was Pamela A. Pelton. It was set in Times Roman by Progressive Information Technologies, Inc. Printed and bound by R. R. Donnelley & Sons Company. This book was printed on acid-free paper. McGraw-Hill books are available at special quantity discounts to use as premiums and sales promotions, or for use in corporate training programs. For more information, please write to the Director of Special Sales, Professional Publishing, McGraw-Hill, Two Penn Plaza, New York, NY 10121-2298. Or contact your local bookstore.

Information contained in this work has been obtained by The McGraw-Hill Companies, Inc. (“McGraw-Hill”) from sources believed to be reliable. However, neither McGraw-Hill nor its authors guarantee the accuracy or completeness of any information published herein, and neither McGraw-Hill nor its authors shall be responsible for any errors, omissions, or damages arising out of use of this information. This work is published with the understanding that McGraw-Hill and its authors are supplying information but are not attempting to render engineering or other professional services. If such services are required, the assistance of an appropriate professional should be sought.

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This book is dedicated to my dear wife Joyce for her encouragement, understanding, and contributions, and to my daughter Helen and my sons Roger and David.

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ABOUT THE AUTHOR

Shan K. Wang received his B.S. in mechanical engineering from Southwest Associated University in China in 1946. Two years later, he completed his M.S. de gree in mechanical engineering at Harvard Graduate School of Engineering. In 1949, he obtained his M.S. in te xtile technology from the Massachusetts Institute of Technology. From 1950 to 1974, Wang worked in the field of air conditioning and refrigeration in China. H was the first Technical Deputy Director of the Research Institute of Air Conditioning in Beijing from 1963 to 1966 and from 1973 to 1974. He helped to design space dif fusion for the air conditioning system in the Capital and Worker’s Indoor Stadium. He also designed man y HVAC&R systems for industrial and commercial b uildings. Wang published tw o air conditioning books and many papers in the 1950s and 1960s. He is one of the pioneers of air conditioning in China. Wang joined Hong K ong Polytechnic as senior lecturer in 1975. He established the air conditioning and refrigeration laboratories and established courses in air conditioning and refrigeration at Hong Kong Polytechnic. Since 1975, he has been a consultant to Associated Consultant Engineers and led the design of the HVAC&R systems for Queen Elizabeth Indoor Stadium, Aberdeen Market Complex, Koshan Road Recreation Center, and South Sea Textile Mills in Hong K ong. From 1983 to 1987, Wang Published Principles of Refrig eration Engineering and Air Conditioning as the teaching and learning package, and presented se veral papers at ASHRAE meetings. The First Edition of the Handbook of Air Conditioning and Refrigeration was published in 1993. Wang has been a member of ASHRAE since 1976. He has been a go vernor of the ASHRAE Hong Kong Chapter-At-Large since the Chapter w as established in 1984. Wang retired from Hong Kong Polytechnic in June 1987 and immigrated to the United States in October 1987. Since then, he has joined the ASHRAE Southern California Chapter and devoted most of his time to writing.

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PREFACE TO SECOND EDITION

Air conditioning, or HVAC&R, is an acti ve, rapidly developing technology. It is closely related to the living standard of the people and to the outdoor en vironment, such as through ozone depletion and global w arming. Currently, air conditioning consumes about one-sixth of the annual national energy use in the United States. At the be ginning of a ne w millennium, in addition to the publication of ASHRAE Standard 90.1-1999 and ASHRAE Standard 62-1999, often called the Ener gy standard and Indoor Air Quality standard, the second edition of Handbook of Air Conditioning and Refrig eration is intended to summarize the follo wing advances, developments, and valuable experience in HVAC&R technology as they pertain to the design and effective, energy-efficient operation of H AC&R systems: First, to solve the primary problems that e xist in HVAC&R, improve indoor air quality through minimum v entilation control by means of CO 2-based demand-controlled or mix ed plenum controlled ventilation, toxic gas adsorption and chemisorption, medium- and high-ef ficien y filtration and damp surf ace prevention along conditioned air passages. ANSI/ASHRAE Standard 52.2-1999 uses 16 minimum ef ficien y reporting v alues (MERVs) to select air filters based on particle-siz composite efficien y. Energy conservation is a k ey factor in mitigating the global w arming effect. Electric dere gulation and the use of real-time pricing instead of the time-of-use rate structure in the United States have a significant impact on the ene gy cost. ANSI/ASHRAE Standard 90.1-1999 has accumulated valuable HVAC&R ener gy-efficient xperiences since the publication of Standard 90.1-1989 and during the discussions of the two public reviews. For buildings of one or tw o stories when the outdoor wind speed is normal or less than normal, the space or building pressurization depends mainly on the air balance of the HVAC&R system and on the leakiness of the b uilding. A proper space pressurization helps to pro vide a desirable indoor environment. Second, there is a need for a well-designed and -maintained microprocessor -based energy management and control system for medium-size or lar ge projects with generic controls in graphical display, monitoring, trending, totalization, scheduling, alarming, and numerous specific functiona controls to perform HV AC&R operations in air , water, heating, and refrigeration systems. HVAC&R operations must be controlled because the load and outside weather vary. The sequence of operations comprises basic HV AC&R operations and controls. In the second edition, the sequence of operations of zone temperature control of a single-zone VAV system, a VAV reheat system, a dual-duct VAV system, a fan-powered VAV system, and a four -pipe fan-coil system is analyzed. Also the sequence of operations of a plant-b uilding loop water system control, the discharge air temperature control, and duct static pressure control in an air-handling unit are discussed. Third, new and updated advanced technology improvements include • Artificial intelligence such as fuzzy logic, artificial neural net orks, and e xpert systems, is widely used in microprocessor-based controllers. • BACnet is an open protocol in control that enables system components from different vendors to be connected to a single control system to maximize efficien y at lowest cost. • Computational fluid dynamics is becoming an important simulation technology in airf w, space diffusion, clean rooms, and heat-transfer developments.

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PREFACE

• Scroll compressors are gradually replacing reciprocating compressors in packaged units and chillers because of their higher efficien y and simple construction. • Ice storage systems with cold air distrib ution shift the electric po wer demand from on-peak hours to off-peak hours and thus significantly reduce the ene gy cost. • Desiccant-based air conditioning systems replace part of the refrigeration by using e vaporative cooling or other systems in supermarkets, medical operation suites, and ice rinks. • Fault detection and diagnostics determine the reason for defects and f ailures and recommend a means to solve the problem. It is a key device in HVAC&R operation and maintenance. Fourth, air conditioning is designed and operated as a system. In the second edition, HVAC&R systems are classified in three l vels. At the air conditioning system le vel, systems are classified a individual, evaporative, space, packaged, desiccant-based, thermal storage, clean-room, and central systems. At the subsystem level, systems are classified as ai , water, heating, refrigeration, and control systems. At the main component le vel, components such as f ans, coils, compressors, boilers, evaporators, and condensers are further di vided and studied. Each air conditioning system has its own system characteristics. Ho wever, each air conditioning system, subsystem, and main component can be clearly distinguished from the others, so one can thus easily , properly, and more precisely select a require system. Fifth, computer-aided design and drafting (CADD) links the engineering design through calculations and the graphics to drafting. CADD provides the ability to develop and compare the alternative design schemes quickly and the capability to redesign or to match the changes during construction promptly. A savings of 40 percent of design time has been claimed. Current CADD for HVAC&R can be divided into two categories: engineering design, including calculations, and graphical model drafting. Engineering design includes load calculations, energy use estimates, equipment selection, equipment schedules, and specifications. Compute -aided drafting includes softw are to develop duct and pipe work layouts and to produce details of refrigeration plant, heating plant, and fan room with accessories.

ACKNOWLEDGMENTS The author wishes to e xpress his sincere thanks to McGra w-Hill editors Linda R. Lude wig and David Fogarty, Professor Emeritus W. F. Stoecker, Steve Chen, and Professor Yongquan Zhang for their valuable guidance and kind assistance. Thanks also to ASHRAE, EIA, and many others for the use of their published materials. The author also wishes to thank Philip Yu and Dr. Sam C. M. Hui for their help in preparing the manuscript, especially to Philip for his assistance in calculating the cooling load of Example 6.2 by using load calculation software TRACE 600. Shan K. Wang

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Air conditioning, or more specificall , heating, ventilating, air ventilating, air conditioning, and refrigeration (HVAC&R), was first systematically d veloped by Dr . Willis H. Carrier in the early 1900s. Because it is closely connected with the comfort and health of the people, air conditioning became one of the most significant actors in national energy consumption. Most commercial buildings in the United States were air conditioned after World War II. In 1973, the ener gy crisis stimulated the de velopment of v ariable-air-volume systems, energy management, and other HV AC&R technology. In the 1980s, the introduction of microprocessor based direct-digital control systems raised the technology of air conditioning and refrigeration to a higher level. Today, the standards of a successful and cost-ef fective new or retrofit H AC&R projects include maintaining a healthy and comfortable indoor en vironment with adequate outdoor ventilation air and acceptable indoor air quality with an ener gy index lower than that required by the federal and local codes, often using off-air conditioning schemes to reduce energy costs. The purpose of this book is to provide a useful, practical, and updated technical reference for the design, selection, and operation of air conditioning and refrigeration systems. It is intended to summarize the v aluable experience, calculations, and design guidelines from current technical papers, engineering manuals, standards, ASHRAE handbooks, and other publications in air conditioning and refrigeration. It is also intended to emphasize a systemwide approach, especially system operating characteristics at design load and part load. It provides a technical background for the proper selection and operation of optimum systems, subsystems, and equipment. This handbook is a logical combination of practice and theory, system and control, and experience and updated new technologies. Of the 32 chapters in this handbook, the first 30 were written by the autho , and the last tw o were written by Walter P. Bishop, P. E., president of Walter P. Bishop, Consulting Engineer, P. C., who has been an HV AC&R consulting engineer since 1948. Walter also pro vided many insightful comments for the other 30 chapters. Another contributor, Herbert P. Becker, P. E., reviewed Chaps. 1 through 6.

ACKNOWLEDGMENTS The authors wishes to express his sincere thanks to McGra w-Hill Senior Editor Robert Hauserman, G. M. Eisensber g, Robert O. P armley, and Robert A. Parsons for their v aluable guidance and kind assistance. Thanks also to ASHRAE, EIA, SMACNA, The Trane Compan y, Carrier Corporation, Honeywell, Johnson Controls, and many others for the use of their published materials. The author also wishes to thank Leslie Kwok, Colin Chan, and Susanna Chang, who assisted in the preparation of the manuscript. Shan K. Wang

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CONTENTS

Preface to Second Edition xi Preface to First Edition xiii

Chapter 1. Introduction

1.1

Chapter 2. Psychrometrics

2.1

Chapter 3. Heat and Moisture Transfer through Building Envelope

3.1

Chapter 4. Indoor and Outdoor Design Conditions

4.1

Chapter 5. Energy Management and Control Systems

5.1

Chapter 6. Load Calculations

6.1

Chapter 7. Water Systems

7.1

Chapter 8. Heating Systems, Furnaces, and Boilers

8.1

Chapter 9. Refrigerants, Refrigeration Cycles, and Refrigeration Systems

9.1

Chapter 10. Refrigeration Systems: Components

10.1

Chapter 11. Refrigeration Systems: Reciprocating, Rotary, Scroll, and Screw

11.1

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Chapter 12. Heat Pumps, Heat Recovery, Gas Cooling, and Cogeneration Systems

12.1

Chapter 13. Refrigeration Systems: Centrifugal

13.1

Chapter 14. Refrigeration Systems: Absorption

14.1

Chapter 15. Air Systems: Components — Fans, Coils, Filters, and Humidifiers

15.1

Chapter 16. Air Systems: Equipment — Air-Handling Units and Packaged Units

16.1

Chapter 17. Air Systems: Air Duct Design

17.1

Chapter 18. Air Systems: Space Air Diffusion

18.1

Chapter 19. Sound Control

19.1

Chapter 20. Air Systems: Basics and Constant-Volume Systems

20.1

Chapter 21. Air Systems: Variable-Air-Volume Systems

21.1

Chapter 22. Air Systems: VAV Systems — Fan Combination, System Pressure, and Smoke Control

22.1

Chapter 23. Air Systems: Minimum Ventilation and VAV System Controls

23.1

Chapter 24. Improving Indoor Air Quality

24.1

Chapter 25. Energy Management and Global Warming

25.1

Chapter 26. Air Conditioning Systems: System Classification, Selection, and Individual Systems

26.1

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Chapter 27. Air Conditioning Systems: Evaporative Cooling Systems and Evaporative Coolers

27.1

Chapter 28. Air Conditioning Systems: Space Conditioning Systems

28.1

Chapter 29. Air Conditioning Systems: Packaged Systems and Desiccant-Based Systems

29.1

Chapter 30. Air Conditioning Systems: Central Systems and Clean-Room Systems

30.1

Chapter 31. Air Conditioning Systems: Thermal Storage Systems

31.1

Chapter 32. Commissioning and Maintenance

32.1

Appendix A. Nomenclature and Abbreviations

A.1

Appendix B. Psychrometric Chart, Tables of Properties, and I-P Units to SI Units Conversion

B.1

Index follows Appendix B

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CHAPTER 1

INTRODUCTION 1.1 AIR CONDITIONING 1.1 1.2 COMFORT AND PROCESSING AIR CONDITIONING SYSTEMS 1.2 Air Conditioning Systems 1.2 Comfort Air Conditioning Systems 1.2 Process Air Conditioning Systems 1.3 1.3 CLASSIFICATION OF AIR CONDITIONING SYSTEMS ACCORDING TO CONSTRUCTION AND OPERATING CHARACTERISTICS 1.3 Individual Room Air Conditioning Systems 1.4 Evaporative-Cooling Air Conditioning Systems 1.4 Desiccant-Based Air Conditioning Systems 1.4 Thermal Storage Air Conditioning Systems 1.5 Clean-Room Air Conditioning Systems 1.5 Space Conditioning Air Conditioning Systems 1.5 Unitary Packaged Air Conditioning Systems 1.6 1.4 CENTRAL HYDRONIC AIR CONDITIONING SYSTEMS 1.6 Air System 1.6 Water System 1.8 Central Plant 1.8 Control System 1.9 Air, Water, Refrigeration, and Heating Systems 1.10 1.5 DISTRIBUTION OF SYSTEMS USAGE 1.10 1.6 HISTORICAL DEVELOPMENT 1.11 Central Air Conditioning Systems 1.11 Unitary Packaged Systems 1.12 Refrigeration Systems 1.12 1.7 POTENTIALS AND CHALLENGES 1.13

Providing a Healthy and Comfortable Indoor Environment 1.13 The Cleanest, Quietest, and Most Precise and Humid Processing Environment 1.13 Energy Use and Energy Efficiency 1.13 Environmental Problems — CFCs and Global Warming 1.15 Air Conditioning or HVAC&R Industry 1.15

1.8 AIR CONDITIONING PROJECT DEVELOPMENT 1.16 Basic Steps in Development 1.16 Design-Bid and Design-Build 1.17 The Goal — An Environmentally Friendlier, Energy-Efficient, and Cost-Effective HVAC&R System 1.17 Major HVAC&R Problems 1.17 1.9 DESIGN FOR AIR CONDITIONING SYSTEM 1.18 Engineering Responsibilities 1.18 Coordination between Air Conditioning and Other Trades, Teamwork 1.19 Retrofit, Remodeling, and Replacement 1.19 Engineer’s Quality Control 1.20 Design of the Control System 1.20 Field Experience 1.21 New Design Technologies 1.21 1.10 DESIGN DOCUMENTS 1.21 Drawings 1.22 Specifications 1.22 1.11 CODES AND STANDARDS 1.23 1.12 COMPUTER-AIDED DESIGN AND DRAFTING (CADD) 1.25 Features of CADD 1.25 Computer-Aided Design 1.25 Computer-Aided Drafting (CAD) 1.26 Software Requirements 1.26 REFERENCES 1.26

1.1 AIR CONDITIONING Air conditioning is a combined process that performs man y functions simultaneously. It conditions the air, transports it, and introduces it to the conditioned space. It provides heating and cooling from its central plant or rooftop units. It also controls and maintains the temperature, humidity, air movement, air cleanliness, sound level, and pressure dif ferential in a space within predetermined 1.1

1.2

CHAPTER ONE

limits for the comfort and health of the occupants of the conditioned space or for the purpose of product processing. The term HVAC&R is an abbreviation of heating, ventilating, air conditioning, and refrigerating. The combination of processes in this commonly adopted term is equi valent to the current definitio of air conditioning. Because all these indi vidual component processes were de veloped prior to the more complete concept of air conditioning, the term HVAC&R is often used by the industry.

1.2 COMFORT AND PROCESSING AIR CONDITIONING SYSTEMS Air Conditioning Systems An air conditioning, or HVAC&R, system is composed of components and equipment arranged in sequence to condition the air, to transport it to the conditioned space, and to control the indoor environmental parameters of a specific space within required limits Most air conditioning systems perform the following functions: 1. Provide the cooling and heating energy required 2. Condition the supply air , that is, heat or cool, humidify or dehumidify , clean and purify , and attenuate any objectionable noise produced by the HVAC&R equipment 3. Distribute the conditioned air, containing sufficient outdoor ai , to the conditioned space 4. Control and maintain the indoor en vironmental parameters – such as temperature, humidity, cleanliness, air movement, sound level, and pressure differential between the conditioned space and surroundings — within predetermined limits Parameters such as the size and the occupancy of the conditioned space, the indoor environmental parameters to be controlled, the quality and the ef fectiveness of control, and the cost involved determine the various types and arrangements of components used to pro vide appropriate characteristics. Air conditioning systems can be classified according to their applications as (1) comfort ai conditioning systems and (2) process air conditioning systems. Comfort Air Conditioning Systems Comfort air conditioning systems pro vide occupants with a comfortable and healthy indoor en vironment in which to carry out their acti vities. The various sectors of the economy using comfort air conditioning systems are as follows: 1. The commercial sector includes of fice uildings, supermarkets, department stores, shopping centers, restaurants, and others. Man y high-rise of fice uildings, including such structures as the World Trade Center in Ne w York City and the Sears Tower in Chicago, use complicated air conditioning systems to satisfy multiple-tenant requirements. In light commercial b uildings, the air conditioning system serves the conditioned space of only a single-zone or comparati vely smaller area. For shopping malls and restaurants, air conditioning is necessary to attract customers. 2. The institutional sector includes such applications as schools, colleges, universities, libraries, museums, indoor stadiums, cinemas, theaters, concert halls, and recreation centers. F or e xample, one of the large indoor stadiums, the Superdome in New Orleans, Louisiana, can seat 78,000 people. 3. The residential and lodging sector consists of hotels, motels, apartment houses, and private homes. Man y systems serving the lodging industry and apartment houses are operated continuously, on a 24-hour, 7-day-a-week schedule, since they can be occupied at any time. 4. The health care sector encompasses hospitals, nursing homes, and convalescent care f acilities. Special air filters are generally used in hospitals to rem ve bacteria and particulates of submicrometer

INTRODUCTION

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size from areas such as operating rooms, nurseries, and intensive care units. The relative humidity in a general clinical area is often maintained at a minimum of 30 percent in winter. 5. The transportation sector includes aircraft, automobiles, railroad cars, buses, and cruising ships. P assengers increasingly demand ease and en vironmental comfort, especially for longdistance tra vel. Modern airplanes flying at high altitudes may require a pressure di ferential of about 5 psi between the cabin and the outside atmosphere. According to the Commercial Buildings Characteristics (1994), in 1992 in the United States, among 4,806,000 commercial b uildings having 67.876 billion ft 2 (6.31 billion m 2) of floor area 84.0 percent were cooled, and 91.3 percent were heated.

Process Air Conditioning Systems Process air conditioning systems pro vide needed indoor en vironmental control for manuf acturing, product storage, or other research and de velopment processes. The following areas are examples of process air conditioning systems: 1. In textile mills, natural fibers and manu actured fibers are hygroscopic. Proper control of hu midity increases the strength of the yarn and fabric during processing. For many textile manufacturing processes, too high a v alue for the space relati ve humidity can cause problems in the spinning process. On the other hand, a lower relative humidity may induce static electricity that is harmful for the production processes. 2. Many electronic products require clean rooms for manufacturing such things as integrated circuits, since their quality is adv ersely af fected by airborne particles. Relati ve-humidity control is also needed to pre vent corrosion and condensation and to eliminate static electricity . Temperature control maintains materials and instruments at stable condition and is also required for workers who wear dust-free garments. F or example, a class 100 clean room in an electronic f actory requires a temperature of 72  2°F (22.2  1.1°C), a relative humidity at 45  5 percent, and a count of dust particles of 0.5-m (1.97  105 in.) diameter or larger not to exceed 100 particles /ft3 (3531 particles/m3). 3. Precision manufacturers always need precise temperature control during production of precision instruments, tools, and equipment. Bausch and Lomb successfully constructed a constanttemperature control room of 68  0.1°F (20  0.56°C) to produce light grating products in the 1950s. 4. Pharmaceutical products require temperature, humidity, and air cleanliness control. F or instance, liver extracts require a temperature of 75°F (23.9°C) and a relati ve humidity of 35 percent. If the temperature exceeds 80°F (26.7°C), the extracts tend to deteriorate. High-efficien y air filter must be installed for most of the areas in pharmaceutical factories to prevent contamination. 5. Modern refrigerated w arehouses not only store commodities in coolers at temperatures of 27 to 32°F (  2.8 to 0°C) and frozen foods at  10 to  20°F ( 23 to  29°C), but also pro vide relative-humidity control for perishable foods between 90 and 100 percent. Refrigerated storage is used to pre vent deterioration. Temperature control can be performed by refrigeration systems only, but the simultaneous control of both temperature and relati ve humidity in the space can only be performed by process air conditioning systems.

1.3 CLASSIFICATION OF AIR CONDITIONING SYSTEMS ACCORDING TO CONSTRUCTION AND OPERATING CHARACTERISTICS Air conditioning systems can also be classified according to their construction and operatin characteristics as follows.

1.4

CHAPTER ONE

Individual Room Air Conditioning Systems Individual room, or simply indi vidual air conditioning systems emplo y a single, self-contained room air conditioner, a packaged terminal, a separated indoor-outdoor split unit, or a heat pump. A heat pump e xtracts heat from a heat source and rejects heat to air or w ater at a higher temperature for heating. Unlike other systems, these systems normally use a totally independent unit or units in each room. Individual air conditioning systems can be classified into t o categories: ●

●

Room air conditioner (window-mounted) Packaged terminal air conditioner (PTAC), installed in a sleeve through the outside wall

The major components in a f actory-assembled and ready-for -use room air conditioner include the following: An evaporator fan pressurizes and supplies the conditioned air to the space. In tubeand-fin coil the refrigerant e vaporates, expands directly inside the tubes, and absorbs the heat energy from the ambient air during the cooling season; it is called a direct e xpansion (DX) coil. When the hot refrigerant releases heat energy to the conditioned space during the heating season, it acts as a heat pump. An air filter rem ves airborne particulates. A compressor compresses the refrigerant from a lower evaporating pressure to a higher condensing pressure. A condenser liquefies refriger ant from hot gas to liquid and rejects heat through a coil and a condenser fan. A temperature control system senses the space air temperature (sensor) and starts or stops the compressor to control its cooling and heating capacity through a thermostat (refer to Chap. 26). The difference between a room air conditioner and a room heat pump, and a packaged terminal air conditioner and a packaged terminal heat pump, is that a four-way reversing valve is added to all room heat pumps. Sometimes room air conditioners are separated into tw o split units: an outdoor condensing unit with compressor and condenser , and an indoor air handler in order to ha ve the air handler in a more advantageous location and to reduce the compressor noise indoors. Individual air conditioning systems are characterized by the use of a DX coil for a single room. This is the simplest and most direct w ay of cooling the air . Most of the indi vidual systems do not employ connecting ductw ork. Outdoor air is introduced through an opening or through a small air damper. Individual systems are usually used only for the perimeter zone of the building.

Evaporative-Cooling Air Conditioning Systems Evaporative-cooling air conditioning systems use the cooling ef fect of the e vaporation of liquid water to cool an airstream directly or indirectly . It could be a f actory-assembled packaged unit or a field- uilt system. When an evaporative cooler provides only a portion of the cooling ef fect, then it becomes a component of a central hydronic or a packaged unit system. An evaporative-cooling system consists of an intake chamber, f lter(s), supply fan, direct-contact or indirect-contact heat e xchanger, exhaust fan, water sprays, recirculating water pump, and water sump. Evaporative-cooling systems are characterized by lo w energy use compared with refrigeration cooling. They produce cool and humid air and are widely used in southwest arid areas in the United States (refer to Chap. 27).

Desiccant-Based Air Conditioning Systems A desiccant-based air conditioning system is a system in which latent cooling is performed by desiccant dehumidification and sensible cooling by vaporative cooling or refrigeration. Thus, a considerable part of expensive vapor compression refrigeration is replaced by ine xpensive evaporative cooling. A desiccant-based air conditioning system is usually a hybrid system of dehumidifica tion, evaporative cooling, refrigeration, and regeneration of desiccant (refer to Chap. 29). There are two airstreams in a desiccant-based air conditioning system: a process airstream and a regenerative airstream. Process air can be all outdoor air or a mixture of outdoor and recirculating

INTRODUCTION

1.5

air. Process air is also conditioned air supplied directly to the conditioned space or enclosed manufacturing process, or to the air -handling unit (AHU), packaged unit (PU), or terminal for further treatment. Regenerative airstream is a high-temperature airstream used to reacti vate the desiccant. A desiccant-based air conditioned system consists of the follo wing components: rotary desiccant dehumidifiers heat pipe heat e xchangers, direct or indirect e vaporative coolers, DX coils and v apor compression unit or water cooling coils and chillers, fans, pumps, filters controls, ducts, and piping.

Thermal Storage Air Conditioning Systems In a thermal storage air conditioning system or simply thermal storage system, the electricity-driven refrigeration compressors are operated during of f-peak hours. Stored chilled w ater or stored ice in tanks is used to pro vide cooling in b uildings during peak hours when high electric demand char ges and electric ener gy rates are in ef fect. A thermal storage system reduces high electric demand for HVAC&R and partially or fully shifts the high electric energy rates from peak hours to off-peak hours. A thermal storage air conditioning system is al ways a central air conditioning system using chilled water as the cooling medium. In addition to the air , water, and refrigeration control systems, there are chilled-w ater tanks or ice storage tanks, storage circulating pumps, and controls (refer to Chap. 31).

Clean-Room Air Conditioning Systems Clean-room or clean-space air conditioning systems serv e spaces where there is a need for critical control of particulates, temperature, relative humidity, ventilation, noise, vibration, and space pressurization. In a clean-space air conditioning system, the quality of indoor en vironmental control directly affects the quality of the products produced in the clean space. A clean-space air conditioning system consists of a recirculating air unit and a mak eup air unit — both include dampers, prefilters coils, fans, high-efficien y particulate air (HEP A) filters ductwork, piping work, pumps, refrigeration systems, and related controls except for a humidifier i the makeup unit (refer to Chap. 30).

Space Conditioning Air Conditioning Systems Space conditioning air conditioning systems are also called space air conditioning systems . They have cooling, dehumidification heating, and filtration performed predominately by an coils, watersource heat pumps, or other de vices within or abo ve the conditioned space, or very near it. A fan coil consists of a small f an and a coil. A water-source heat pump usually consists of a f an, a finne coil to condition the air , and a water coil to reject heat to a w ater loop during cooling, or to extract heat from the same w ater loop during heating. Single or multiple f an coils are always used to serve a single conditioned room. Usually , a small console w ater-source heat pump is used for each control zone in the perimeter zone of a b uilding, and a large water-source heat pump may serve several rooms with ducts in the core of the building (interior zone, refer to Chap. 28). Space air conditioning systems normally ha ve only short supply ducts within the conditioned space, and there are no return ducts e xcept the lar ge core w ater-source heat pumps. The pressure drop required for the recirculation of conditioned space air is often equal to or less than 0.6 in. w ater column (WC) (150 P a). Most of the ener gy needed to transport return and recirculating air is sa ved in a space air conditioning system, compared to a unitary packaged or a central hydronic air conditioning system. Space air conditioning systems are usually emplo yed with a dedicated (separate) outdoor v entilation air system to pro vide outdoor air for the occupants in the conditioned space. Space air conditioning systems often ha ve comparati vely higher noise le vel and need more periodic maintenance inside the conditioned space.

1.6

CHAPTER ONE

Unitary Packaged Air Conditioning Systems Unitary packaged air conditioning systems can be called, in brief, packaged air conditioning systems or packaged systems. These systems emplo y either a single, self-contained packaged unit or two split units. A single packaged unit contains f ans, filters DX coils, compressors, condensers, and other accessories. In the split system, the indoor air handler comprises controls and the air system, containing mainly fans, filters and DX coils; and the outdoor condensing unit is the refrigeration system, composed of compressors and condensers. Rooftop packaged systems are most widely used (refer to Chap. 29). Packaged air conditioning systems can be used to serv e either a single room or multiple rooms. A supply duct is often installed for the distrib ution of conditioned air, and a DX coil is used to cool it. Other components can be added to these systems for operation of a heat pump system; i.e., a centralized system is used to reject heat during the cooling season and to condense heat for heating during the heating season. Sometimes perimeter baseboard heaters or unit heaters are added as a part of a unitary packaged system to provide heating required in the perimeter zone. Packaged air conditioning systems that emplo y large unitary packaged units are central systems by nature because of the centralized air distrib uting ductwork or centralized heat rejection systems. Packaged air conditioning systems are characterized by the use of inte grated, factory-assembled, and ready-to-use packaged units as the primary equipment as well as DX coils for cooling, compared to chilled water in central hydronic air conditioning systems. Modern lar ge rooftop packaged units have many complicated components and controls which can perform similar functions to the central hydronic systems in many applications.

1.4 CENTRAL HYDRONIC AIR CONDITIONING SYSTEMS Central hydronic air conditioning systems are also called central air conditioning systems. In a central hydronic air conditioning system, air is cooled or heated by coils filled with chilled or hot ater distributed from a central cooling or heating plant. It is mostly applied to lar ge-area buildings with many zones of conditioned space or to separate buildings. Water has a far greater heat capacity than air. The following is a comparison of these tw o media for carrying heat energy at 68°F (20°C): Air Specific heat Btu/lb °F Density, at 68°F, lb/ft3 Heat capacity of fluid at 68° , Btu/ft3  °F

Water

0.243 1.0 0.075 62.4 0.018 62.4

The heat capacity per cubic foot (meter) of w ater is 3466 times greater than that of air . Transporting heating and cooling energy from a central plant to remote air -handling units in fan rooms is far more efficient using ater than conditioned air in a lar ge air conditioning project. Ho wever, an additional water system lowers the evaporating temperature of the refrigerating system and makes a small- or medium-size project more complicated and expensive. A central hydronic system of a high-rise of fice uilding, the NBC Tower in Chicago, is illustrated in Fig. 1.1. A central hydronic air conditioning system consists of an air system, a w ater system, a central heating/cooling plant, and a control system. Air System An air system is sometimes called the air-handling system . The function of an air system is to condition, to transport, to distribute the conditioned, recirculating, outdoor, and exhaust air, and to control the indoor environment according to requirements. The major components of an air system

FIGURE 1.1 Schematic diagram of the central hydronic air conditioning system in NBC Tower.

1.7

1.8

CHAPTER ONE

are the air-handling units, supply/return ductwork, fan-powered boxes, space diffusion devices, and exhaust systems. An air-handling unit (AHU) usually consists of supply f an(s), filter(s) a cooling coil, a heating coil, a mixing box, and other accessories. It is the primary equipment of the air system. An AHU conditions the outdoor /recirculating air, supplies the conditioned air to the conditioned space, and extracts the returned air from the space through ductwork and space diffusion devices. A fan-powered variable-air-volume (VAV) box, often abbreviated as f an-powered box, employs a small fan with or without a heating coil. It dra ws the return air from the ceiling plenum, mixes it with the conditioned air from the air -handling unit, and supplies the mixture to the conditioned space. Space diffusion devices include slot diffusers mounted in the suspended ceiling; their purpose is to distribute the conditioned air e venly over the entire space according to requirements. The return air enters the ceiling plenum through many scattered return slots. Exhaust systems have exhaust fan(s) and ductw ork to e xhaust air from the la vatories, mechanical rooms, and electrical rooms. The NBC Tower in Chicago is a 37-story high-rise of fice compl x constructed in the late 1980s. It has a total air conditioned area of about 900,000 ft 2 (83,600 m2). Of this, 256,840 ft2 (23,870 m2) is used by NBC studios and other departments, and 626,670 ft 2 (58,240 m2) is rental offices locate on upper floors. Special air conditioning systems are empl yed for NBC studios and departments at the lower level. For the rental of fice floor four air -handling units are located on the 21st floo . Outdoor air either is mix ed with the recirculating air or enters directly into the air -handling unit as sho wn in Fig. 1.2. The mixture is filtrated at the filter and is then cooled and dehumidified at the cooling c during cooling season. After that, the conditioned air is supplied to the typical floor through th supply fan, the riser, and the supply duct; and to the conditioned space through the fan-powered box and slot diffusers.

Water System The water system includes chilled and hot w ater systems, chilled and hot w ater pumps, condenser water system, and condenser w ater pumps. The purpose of the w ater system is (1) to transport chilled water and hot w ater from the central plant to the air -handling units, fan-coil units, and fanpowered boxes and (2) to transport the condenser water from the cooling tower, well water, or other sources to the condenser inside the central plant. In Figs. 1.1 and 1.2, the chilled w ater is cooled in three centrifugal chillers and then is distributed to the cooling coils of v arious air-handling units located on the 21st floo . The temperature of the chilled water leaving the coil increases after absorbing heat from the airstream fl wing over the coil. Chilled water is then returned to the centrifugal chillers for recooling through the chilled w ater pumps. After the condenser water has been cooled in the cooling tower, it fl ws back to the condenser of the centrifugal chillers on lo wer level 3. The temperature of the condenser w ater again rises o wing to the absorption of the condensing heat from the refrigerant in the condenser . After that, the condenser water is pumped to the cooling towers by the condenser water pumps.

Central Plant The refrigeration system in a central plant is usually in the form of a chiller package. Chiller packages cool the chilled water and act as a cold source in the central hydronic system. The boiler plant, consisting of boilers and accessories, is the heat source of the heating system. Either hot w ater is heated or steam is generated in the boilers. In the NBC Tower, the refrigeration system has three centrifugal chillers located in lower level 3 of the basement. Three cooling towers are on the roof of the b uilding. Chilled water cools from 58 to 42°F (14.4 to 5.6°C) in the e vaporator when the refrigerant is e vaporated. The refrigerant is then

INTRODUCTION

1.9

FIGURE 1.2 Schematic drawing of air system for a typical floor of o fices in the NBC Tower.

compressed to the condensing pressure in the centrifugal compressor and is condensed in liquid form in the condenser, ready for evaporation in the evaporator. There is no boiler in the central plant of the NBC Tower. To compensate heat loss in the perimeter zone, heat energy is pro vided by the w arm plenum air and the electric heating coils in the f anpowered boxes. Control System Modern air conditioning control systems for the air and w ater systems and for the central plant consist of electronic sensors, microprocessor-operated and -controlled modules that can analyze and perform calculations from both digital and analog input signals, i.e., in the form of a continuous variable. Control systems using digital signals compatible with the microprocessor are called direct digital control (DDC) systems. Outputs from the control modules often actuate dampers, valves, and relays by means of pneumatic actuators in lar ge b uildings and by means of electric actuators for small projects. In the NBC Tower, the HVAC&R system is monitored and controlled by a microprocessor-based DDC system. The DDC controllers re gulate the air-handling units and the terminals. Both communicate with the central operating station through interf ace modules. In case of emer gency, the

1.10

CHAPTER ONE

fir protection system detects alarm conditions. The central operating station gi ves emer gency directions to the occupants, operates the HVAC&R system in a smok e control mode, and actuates the sprinkler water system.

Air, Water, Refrigeration, and Heating Systems Air, water, refrigeration, heating, and control systems are the subsystems of an air conditioning or HVAC&R system. Air systems are often called secondary systems. Heating and refrigeration systems are sometimes called primary systems. Central hydronic and space conditioning air conditioning systems both have air, water, refrigeration, heating, and control systems. The w ater system in a space conditioning system may be a chilled/hot water system. It also could be a centralized w ater system to absorb heat from the condenser during cooling, or provide heat for the evaporator during heating. For a unitary packaged system, it consists of mainly air , refrigeration, and control systems. The heating system is usually one of the components in the air system. Sometimes a separate baseboard hot water heating system is employed in the perimeter zone. A evaporative-cooling system always has an air system, a water system, and a control system. A separate heating system is often employed for winter heating. In an indi vidual room air conditioning system, air and refrigeration systems are installed in indoor and outdoor compartments with their own control systems. The heating system is often a component of the supply air chamber in the room air conditioner . It can be also a centralized hot w ater heating system in a PTAC system. Air conditioning or HVAC&R systems are therefore often first described and analyzed throug their subsystems and main components: such as air, water, heating, cooling/refrigeration, and control systems. Air conditioning system classification system operating characteristics, and system selection must take into account the whole system. Among the air , water, and refrigeration systems, the air system conditions the air , controls and maintains the required indoor environment, and has direct contact with the occupants and the manufacturing processes. These are the reasons why the operating characteristics of an air conditioning system are esssentially represented by its air system.

1.5 DISTRIBUTION OF SYSTEMS USAGE According to surv eys conducted in 1995 by the Department of Ener gy/Energy Information Administration (DOE /EIA) of the United States, for a total floor space of 58,772 million f 2 (5462 million m 2) in commercial b uildings in 1995 and for a total of 96.6 million homes in 1993 (among these, 74.1 million homes were air conditioned), the floor space in million square feet, and the number of homes using various types of air conditioning systems are as follows:

Percent of floor spac

Million homes

Percent of homes

12,494 2,451

22 4 (8)

33.1

45

26,628

48

41.0

55

13,586 949

24 2

Million ft2 Individual room systems Evaporative-cooling systems Space conditioning systems (estimated) Unitary packaged systems (including air-source heat pump as well as desiccant-based systems) Central hydronic systems (including thermal storage and clean-room systems) Others

INTRODUCTION

1.11

Much of the floor space may be included in more than one air conditioning system. G ven the possibility that the floor space may be counted repeatedl , the original data listed in the DOE /EIA publication were modified. The 8 percent of the space system includes part in central hydronic systems and part in unitary packaged systems. Among the air conditioned homes in 1993, the unitary packaged system is the predominate air conditioning system in U.S. homes.

1.6 HISTORICAL DEVELOPMENT The historical development of air conditioning can be summarized briefl . Central Air Conditioning Systems As part of a heating system using f ans and coils, the first rudimentary ice system in the Unite States, designed by McKin, Mead, and White, was installed in Ne w York City’s Madison Square Garden in 1880. The system delivered air at openings under the seats. In the 1890s, a leading consulting engineer in Ne w York City, Alfred R. Wolf, used ice at the outside air intak e of the heating and ventilating system in Carne gie Hall. Another central ice system in the 1890s w as installed in the Auditorium Hotel in Chicago by Buf falo Forge Company of Buf falo, New York. Early central heating and v entilating systems used steam-engine-dri ven fans. The mixture of outdoor air and return air w as discharged into a chamber . In the top part of the chamber , pipe coils heat the mixture with steam. In the bottom part is a bypass passage with damper to mix conditioned air and bypass air according to the requirements. Air conditioning w as first systematically d veloped by Willis H. Carrier , who is recognized as the father of air conditioning. In 1902, Carrier discovered the relationship between temperature and humidity and how to control them. In 1904, he developed the air w asher, a chamber installed with several banks of w ater sprays for air humidification and cleaning. His method of temperature an humidity regulation, achieved by controlling the dew point of supply air, is still used in many industrial applications, such as lithographic printing plants and textile mills. Perhaps the first ai -conditioned of fice as the Larkin Administration Building, designed by Frank L. Wright and completed in 1906. Ducts handled air that w as drawn in and exhausted at roof level. Wright specified a refrigeration plant which distri uted 10°C cooling w ater to air -cooling coils in air-handling systems. The U.S. Capitol w as air -conditioned by 1929. Conditioned air w as supplied from o verhead diffusers to maintain a temperature of 75°F (23.9°C) and a relati ve humidity of 40 percent during summer, and 80°F (26.7°C) and 50 percent during winter . The volume of supply air w as controlled by a pressure regulator to prevent cold drafts in the occupied zone. Perhaps the first fully air conditioned o fice uilding w as the Milan Building in San Antonio, Texas, which was designed by George Willis in 1928. This air conditioning system consisted of one centralized plant to serve the lower floors and ma y small units to serve the top office floor In 1937, Carrier developed the conduit induction system for multiroom b uildings, in which recirculation of space air is induced through a heating /cooling coil by a high-v elocity dischar ging airstream. This system supplies only a limited amount of outdoor air for the occupants. The variable-air-volume (VAV) systems reduce the volume fl w rate of supply air at reduced loads instead of varying the supply air temperature as in constant-v olume systems. These systems were introduced in the early 1950s and gained wide acceptance after the ener gy crisis of 1973 as a result of their lower energy consumption in comparison with constant-v olume systems. With many variations, VAV systems are in common use for new high-rise office uildings in the United States today. Because of the rapid development of space technology after the 1960s, air conditioning systems for clean rooms were de veloped into sophisticated arrangements with e xtremely ef fective air filters Central air conditioning systems al ways will pro vide a more precisely controlled, healthy, and safe indoor en vironment for high-rise b uildings, large commercial comple xes, and precisionmanufacturing areas.

1.12

CHAPTER ONE

Unitary Packaged Systems The first room cooler d veloped by Frigidaire was installed about in 1928 or 1929, and the “Atmospheric Cabinet” developed by the Carrier Engineering Compan y w as first installed in May 1931 The first self-contained room air conditioner as developed by General Electric in 1930. It w as a console-type unit with a hermetically sealed motor-compressor (an arrangement in which the motor and compressor are encased together to reduce the leakage of refrigerant) and w ater-cooled condenser, using sulfur dioxide as the refrigerant. Thirty of this type of room air conditioner, were built and sold in 1931. Early room air conditioners were rather b ulky and heavy. They also required a drainage connection for the municipal w ater used for condensing. During the postw ar period the air -cooled model was developed. It used outdoor air to absorb condensing heat, and the size and weight were greatly reduced. Annual sales of room air conditioners have exceeded 100,000 units since 1950. Self-contained unitary packages for commercial applications, initially called store coolers, were introduced by the Airtemp Division of Chrysler Corporation in 1936. The early models had a refrigeration capacity of 3 to 5 tons and used a w ater-cooled condenser . Air-cooled unitary packages gained wide acceptance in the 1950s, and many were split systems incorporating an indoor air handler and an outdoor condensing unit. Packaged units ha ve been de veloped since the 1950s, from indoor to rooftops, from constantvolume to v ariable-air-volume, and from fe w to man y functions. Currently , packaged units enjo y better performance and ef ficien y with better control of capacity to match the space load. Computerized direct digital control, one of the important reasons for this improvement, places unitary packaged systems in a better position to compete with central hydronic systems.

Refrigeration Systems In 1844, Dr. John Gorrie designed the first commercial reciprocating refrigerating machine in th United States. The hermetically sealed motor -compressor was first d veloped by General Electric Company for domestic refrigerators and sold in 1924. Carrier invented the first open-type gea -driven factory-assembled, packaged centrifugal chiller in 1922 in which the compressor w as manuf actured in German y; and the hermetic centrifugal chiller, with a hermetically sealed motor -compressor assembly, in 1934. The direct-driven hermetic centrifugal chiller was introduced in 1938 by The Trane Company. Up to 1937, the capacity of centrifugal chillers had increased to 700 tons. During the 1930s, one of the outstanding de velopments in refrigeration w as the disco very by Midgely and Hene of the nontoxic, nonflammable fluorinated hydrocarbon refrigerant amily called Freon in 1931. Refrigerant-11 and refrigerant-12, the chlorofluorocarbons (CFCs) became widely adopted commercial products in reciprocating and centrifugal compressors. No w, new refrigerants have been de veloped by chemical manuf acturers such as DuPont to replace CFCs, so as to prevent the depletion of the ozone layer. The first aqueous-ammonia absorption refrigeration system as invented in 1815 in Europe. In 1940, Servel introduced a unit using water as refrigerant and lithium bromide as the absorbing solution. The capacities of these units ranged from 15 to 35 tons (52 to 123 kW). Not until 1945 did Carrier introduce the first la ge commercial lithium bromide absorption chillers. These units were developed with 100 to 700 tons (352 to 2460 kW) of capacity , using low-pressure steam as the heat source. Positive-displacement scre w compressors ha ve been de veloped in the United States since the 1950s and scroll compressors since the 1970s because of their higher ef ficien y and smoother rotary motion than reciprocating compressors. No w, the scroll compressors gradually replace the reciprocating compressors in small and medium-size refrigeration systems. Another trend is the development of more ener gy-efficient centrifugal and absorption chillers for ene gy conserv ation. The ener gy consumption per ton of refrigeration of a ne w centrifugal chiller dropped from 0.80 kW/ton (4.4 COP ref) in the late 1970s to 0.50 kW/ton (7.0 COP ref) in the 1990s. A series of rotary motion refrigeration

INTRODUCTION

1.13

compressors with small, medium, to large capacity and using scroll, screw, or centrifugal compressors will be manufactured from now on.

1.7 POTENTIALS AND CHALLENGES Air conditioning or HV AC&R is an industry of man following.

y potentials and challenges,

including the

Providing a Healthy and Comfortable Indoor Environment Nowadays, people in the United States spend most of their time indoors. A healthy and comfortable indoor environment provided by air conditioning is a necessity for people staying indoors, no matter ho w hot or cold and dry or humid the outside climate might be. According to the American Housing Survey conducted by the U.S. Census Bureau in 1991, of 92.3 million homes, 66 million, or 71 percent of the total, were air conditioned, and 81.9 million, or nearly 89 percent of the total, were heated. According to the Ener gy Information Administration (EIA), in 1992, for a total floo area of 67.8 billion ft 2 (6.3 billion m2) of commercial buildings, 84 percent were cooled and 91 percent were heated.

The Cleanest, Quietest, and Most Precise and Humid Processing Environment A class 1 clean room to manuf acture inte grated circuits in a semiconductor f actory may be the cleanest processing en vironment that is pro vided in the semiconductor industry in the 1990s. The dust particle count does not e xceed 1 particle /ft3 (35 particles /m3) of a size of 0.5 m and lar ger, with no particle e xceeding 5 m. A constant-temperature room of 68°F  0.1°F (20  0.56°C) is always surrounded by other constant-temperature rooms or by a b uffer area of lo wer tolerance to maintain the fluctuation of its sensed temperature withi  0.05°F (0.028°C) during w orking hours. A recording studio in a tele vision broadcasting station often needs a noise criteria (NC) curv e (refer to Chap. 4) of NC 15 to 20, in which a sound le vel less than that of a b uzzing insect can be heard. A refrigerated warehouse that stores v egetables such as cabbage, carrots, and celery needs a temperature of 32°F (0°C) and a relati ve humidity of 98 to 100 percent to pre vent deterioration and loss of water. Air conditioning or HVAC&R systems will provide not only the cleanest, precisest, quietest, and most humid environment with fluctuations of the controlled ariable within required limits, but also at optimum energy use and first cost

Energy Use and Energy Efficiency Based on the data published in the Annual Energy Review in 1993, the total energy use in 1992 in the United States w as 82.14 quad Btu or 10 15 Btu (86.66 EJ, or 10 18 J). The United States alone consumed about one-fourth of the w orld’s total production. Of the total ener gy use of 82.14 quad Btu in the United States, the residential /commercial sector consumed about 36 percent of the total, the industrial sector consumed another 36 percent, and transportation consumed the rest — 28 percent. Petroleum, natural gas, and coal were the three main sources, providing more than 85 percent of the energy supply in 1992 in the United States. According to DOE /EIA energy consumption surv ey 0321 for residential b uildings (1993) and survey 0318 for commercial b uildings (1995), the average annual energy use of HVAC&R systems

CHAPTER ONE

in the residential / commercial sector was about 45 percent of the total building energy consumption. Also assuming that the annual ener gy use of HVAC&R systems was about 1 percent of the total in both industrial and transportation sectors, then the estimate of annual ener gy use of the HV AC&R systems in 1992 in the United States w as about 17 percent of the total national ener gy use, or onesixth of the total national energy use. The world energy resources of petroleum, natural gas, and coal are limited. The population of the United States in 1992 w as only about one-twentieth of the w orld’s total population; ho wever, we consumed nearly one-fifth of the orld’s total energy produced. Energy use must be reduced. After the ener gy crisis in 1973, the U.S. Congress enacted the Ener gy Policy and Conservation Act of 1975 and the National Ener gy Policy Act of 1992. The enactment of energy efficien y legislation by federal and state go vernments and the establishment of the Department of Ener gy (DOE) in 1977 had a definite impact on the implementation of ene gy efficien y in United States. In 1975, the American Society of Heating, Refrigerating and Air Conditioning Engineers (ASHRAE) published Standard 90-75, Energy Conservation in New Building Design. This standard was re vised and cosponsored by the Illuminating Engineering Society of North America as ASHRAE/IES Standard 90.1-1989, Energy Ef ficient Design of N w Buildings Except Ne w LowRise Residential Buildings, in 1989; and it w as revised again as current ASHRAE/IESNA Standard 90.1-1999, Energy Standard for Buildings Except Low-Rise Residential Buildings , in 1999. Man y other or ganizations also of fered v aluable contrib utions for ener gy conserv ation. All these e vents started a ne w era in which ener gy efficien y has become one of the important goals of HV AC&R system design and operation. Because of all these efforts, the increase in net annual energy use for residential and commercial buildings, i.e., the total energy use minus the electrical system ener gy loss in the po wer plant, from 1972 to 1992 in the United States w as only 1.3 percent, as shown in Fig. 1.3. Ho wever, during the same period, the increase in floor area of commercial uildings in the United States w as about 83

16.09

15.89

80

15 70

10

60

Energy use

50 40

Commercial building floor space

30

5 20

U. S. commercial buildings, billion ft2

Residential and commercial buildings energy use, 1015 Btu

1.14

10 0 1950

1960

1970

1980

1990

0

FIGURE 1.3 Residential and commercial b uilding energy use and commercial b uilding f oor space increase between 1992 and 1950 in the United States.

INTRODUCTION

1.15

percent. Energy eff ciency will be a challenge to everyone in the HVAC&R industry in this generation as well as in many, many generations to come.

Environmental Problems — CFCs and Global Warming The surface of the earth is surrounded by a layer of air , called the atmosphere. The lower atmosphere is called the homosphere, and the upper atmosphere is called the stratosphere. In the mid-1980s, chlorof uorocarbons (CFCs) were widely used as refrigerants in mechanical refrigeration systems, to produce thermal insulation foam, and to produce aerosol propellants for many household consumer products. CFC-11 (CCl 3F) and CFC-12 (CCl 2F2) are commonly used CFCs. They are very stable. Halons are also halogenated hydrocarbons. If CFCs and halons leak or are discharged from a refrigeration system during operation or repair to the lo wer atmosphere, they will migrate to the upper stratosphere and decompose under the action of ultra violet rays throughout their decades or centuries of life. The free chlorine atoms will react with oxygen atoms of the ozone layer in the upper stratosphere and will cause a depletion of this layer . The theory of the depletion of the ozone layer w as proposed in the early and middle-1970s. The ozone layer f lters out harmful ultra violet rays, which may cause skin cancer and are a serious threat to human beings. Furthermore, changes in the ozone layer may signi f cantly inf uence weather patterns. Since 1996, actions have been taken to ban the production of CFCs and halons, before it is too late. A cloudless homosphere is mainly transparent to short-w ave solar radiation but is quite opaque to long-w ave infrared rays emitted from the surf ace of the earth. Carbon dioxide (CO 2) has the greatest blocking effect of all; water vapor and synthetic CFCs also play important roles in blocking the direct escape of infrared energy. The phenomenon of transparency to incoming solar radiation and blank eting of outgoing infrared rays is called the greenhouse effect. The increase of the CO2, water vapor, CFCs, and other gases, often called greenhouse gases (GHGs), eventually will result in a rise in air temperature near the earth ’s surf ace. This is kno wn as the global w arming effect. Over the past 100 years, global warming has caused an increase of 1 °F. For the same period, there w as a 25 percent increase of CO 2. During the 1980s, the release of CO 2 to the atmosphere was responsible for about 50 percent of the increase in global w arming that w as attrib utable to human activity, and the release of CFCs had a 20 percent share. Some scientists ha ve predicted an accelerated global w arming in the coming 50 years because of the increase in the w orld’s annual energy use. Further increases in global temperatures may lead to lo wer rainfalls, drop in soil moisture, more extensive forest f res, more f ood, etc. CFC production in developed countries has been banned since January 1, 1996. Carbon dioxide is the product of man y combustion processes. Alternative refrigerants to replace CFCs must also have a low global warming potential. Designers and operators of the HVAC&R systems can reduce the production of CO 2 through energy eff ciency and the replacement of coal, petroleum, and natural gas po wer by hydroelectric, solar, and nuclear ener gy, etc. More studies and research are needed to clarify the theory and actual effect of global warming.

Air Conditioning or HVAC&R Industry The air conditioning or HVAC&R industry in the United States is an e xpanding and progressing industry. In 1995, the installed v alue of nonresidential air conditioning hit $20 billion. According to American Refrigeration Institute (ARI) and Heating/Piping/Air Conditioning data, from 1985 to 1995, the annual rate of increase of installed v alue of air conditioning systems is 8.7 percent. Because of the replacement of the old chillers using CFCs, in 1994, the installed v alue of retro f t, remodeling, and replacement accounted for up to tw o-thirds of all HV AC&R e xpenditures. This trend may continue in the beginning of the new century. Based on data from ARI, in 1995, shipment of packaged units reached a record f gure of 5 million products, of which heat pumps comprised a one- f fth share. Centrifugal and scre w chiller

1.16

CHAPTER ONE

shipments were 7500 units, absorption chillers 502 units, signif cantly smaller capacity, were 14,000 units.

and reciprocating chillers,

at a

1.8 AIR CONDITIONING PROJECT DEVELOPMENT Basic Steps in Development The basic steps in the de velopment and use of a lar ge air conditioning system are the design, construction, commissioning, operation, energy eff ciency upgrading, and maintenance. Figure 1.4 is a diagram which outlines the relationship between these steps and the parties in volved. The owner sets the criteria and the requirements. Design professionals in mechanical engineering consulting

FIGURE 1.4 Steps in the development and use of air conditioning systems in buildings.

INTRODUCTION

1.17

f rms design the air conditioning system and prepare the design documents. Manuf acturers supply the equipment, instruments, and materials. Contractors install and construct the air conditioning system. After construction, the air conditioning system is commissioned by a team, and then it is handed over to the operation and maintenance group of the property management for daily operation. F ollowing a certain period of operation, an ener gy service compan y (ESCO) may often be required to upgrade the energy eff ciency of the HVAC&R system (energy retrof t).

Design-Bid and Design-Build There are tw o types of project de velopment: design-bid and design-b uild. A design-bid project separates the design and installation responsibilities, whereas in a design-build project, engineering is done by the installing contractor. Some reasons for a design-build are that the project is too small to retain an engineering consultant, or that there is insuf f cient time to go through the normal design-bid procedures. According to Bengard (1999), the main adv antages of design-b uild include established f rm price early, single-source responsibility, accelerated project deli very, and performance guarantees. The market has experienced nearly 300 percent domestic growth since 1986.

The Goal — An Environmentally Friendlier, Energy-Efficient, and Cost-Effective HVAC&R System The goal is to pro vide an HVAC&R system which is en vironmentally friendlier, energy-eff cient, and cost-effective as follows: ●

●

●

●

●

●

Effectively control indoor en vironmental parameters, usually to k eep temperature and humidity within required limits. Provide an adequate amount of outdoor ventilation air and an acceptable indoor air quality. Use energy-eff cient equipment and HVAC&R systems. Minimize ozone depletion and the global warming effect. Select cost-effective components and systems. Ensure proper maintenance, easy after-hour access, and necessary f re protection and smok e control systems.

Major HVAC&R Problems In Coad’s paper (1985), a study by Wagner-Hohn-Ingles Inc. re vealed that man y large new buildings constructed in the early 1980s in the United States suf fer from complaints and major defects. High on the list of these problems is the HVAC&R system. The major problems are these: 1. Poor indoor air quality (IA Q) — sick b uilding syndrome. Poor indoor air quality causes the sick building syndrome. The National Institute of Occupational Safety and Health (NIOSH) in 1990 reported that between 1971 and 1988, 529 f eld investigations found that lack of outdoor air , improper use, and poor operation and maintenance of HV AC&R systems were responsible for more than one-half of sick b uilding syndrome incidents. Field in vestigations found that 20 to 30 percent of the b uildings had poor -air-quality problems. Sick b uilding syndrome is co vered in detail in Chap. 4. 2. Updated technology. In recent years, there has been a rapid change in the technology of air conditioning. Various types of VAV systems, air and water economizer, heat recovery, thermal storage, desiccant dehumidif cation, variable-speed drives, and DDC devices have become more ef fective and more adv anced for ener gy ef f ciency. Man y HVAC&R designers and operators are not

1.18

CHAPTER ONE

properly equipped to apply and use these systems. Unfortunately , these sophisticated systems are managed, constructed, and operated under the same b udget and schedule constraints as the less sophisticated systems. After years of operation, most HVAC&R equipment and controls need to be upgraded for energy eff ciency. 3. Insufficient communication between design p ofessionals, construction groups, and operators. Effective operation requires a knowledgeable operator to make adjustments if necessary. The operator will operate the system at his or her le vel of understanding. If adequate operating and maintenance documents are not pro vided by the designer and the contractor for the operator , the HVAC&R system may not be operated according to the designer’s intentions. 4. Overlooked commissioning . Commissioning means testing and balancing all systems, functional testing and adjusting of components and the inte grated system, and adjusting and tuning the direct digital controls. An air conditioning system is different from the manufacturing products having models and prototypes. All the defects and errors of the prototypes can be check ed and corrected during their individual tests, but the more complicated HVAC&R system, as constructed and installed, is the end product. Therefore, proper commissioning, which permits the system to perform as specif ed in the design documents, is extremely important. Unfortunately, the specif cations seldom clearly designate competent technicians for the responsibility of commissioning the entire integrated system. Commissioning is covered in detail in Chap. 32. 5. Reluctant to try inno vative approaches. In addition to Coad ’s paper, a survey was conducted by the American Consulting Engineers Council (ACEC) in 1995. For the 985 engineering f rms that responded out of about 5500 f rms total, there were 522 le gal claims f led against them, 49 percent of suits resolv ed without payment. Among the 522 le gal claims, 9 percent were f led against mechanical engineering f rms. Fifty-tw o percent of 1995 le gal claims were f led by o wners, and 13 percent were f led by contractors or subcontractors. Because of the le gal claims, insurance companies discourage inno vation, and engineering consulting f rms are reluctant to try inno vative techniques. Twenty-one percent of these f rms said they were v ery much reluctant, 61 percent said the y were some what affected, and 25 percent said they were a little affected. Support and encouragement must be gi ven to engineering f rms for carefully developed innovative approaches to projects. In addition to the ASHRAE Technology Awards, more HVAC&R inno vative a wards should be established in lar ge cities in the United States to promote innovative approaches.

1.9 DESIGN FOR AIR CONDITIONING SYSTEM System design determines the basic characteristics. After an air conditioning system is constructed according to the design, it is diff cult and expensive to change the design concept. Engineering Responsibilities The normal procedure in a design-bid project includes the following steps and requirements: 1. 2. 3. 4.

Initiation of a construction project by owner or developer Selection of design team Setting of the design criteria and indoor environmental parameters Selection of conceptual alternati ves for systems and subsystems; preparation of schematic layouts of HVAC&R 5. Preparation of contract documents, working drawings, specif cations, materials and construction methods, commissioning guidelines 6. Competitive bidding by contractors 7. Evaluation of bids; negotiations and modif cation of contract documents

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1.19

8. 9. 10. 11. 12.

Advice on awarding of contract Review of shop drawings and commissioning schedule, operating and maintenance manuals Monitoring, supervision, and inspection of construction Supervision of commissioning: testing and balancing; functional performance tests Modif cation of dra wings to the as-b uilt condition and the f nalization of the operation and maintenance manual 13. Acceptance Construction work starts at contract award following the bidding and negotiating and ends at the acceptance of the project after commissioning. It is necessary for the designer to select among the a vailable alternatives for optimum comfort, economics, energy conservation, noise, safety, f exibility, reliability, convenience, and maintainability. Experience, education, and judgment all enter into the selection process. If both a complicated system and a simple system yield the same performance, the simple system is preferred for its reliability, operator convenience, and lower cost.

Coordination between Air Conditioning and Other Trades, Teamwork Air conditioning, plumbing, and f re protection systems are mechanical systems in a b uilding. Both mechanical and electrical systems pro vide building services for the occupants and goods inside the building. Coordination between the mechanical engineer for HVAC&R and the architect, as well as between mechanical and structural or electrical engineers, or teamwork, becomes important. Factors requiring input from both the architect and the mechanical engineer include the follo wing: 1. Shape and the orientation of the building 2. Thermal characteristics of the b uilding envelope, especially the type and size of the windo ws and the construction of external walls and roofs 3. Location of the ductw ork and piping to a void interference with each other , or with other trades 4. Layout of the diffusers and supply and return grilles 5. Minimum clearance provided between the structural members and the suspended ceiling for the installation of ductwork and piping 6. Location and size of the rooms for central plant, fan rooms, duct and pipe shafts If the architect mak es a decision that is thermally unsound, the HVAC&R engineer must of fset the additional loads by increasing the HVAC&R system capacity. Lack of such coordination results in greater energy consumption. HVAC&R and other building services must coordinate the following: 1. 2. 3. 4.

Utilization of daylight and the type of artif cial light to be installed The layout of diffusers, grilles, return inlets, and light troffers in the suspended ceiling Integration with f re alarm and smoke control systems Electric power and plumbing requirements for the HV AC&R equipment and lighting for equipment rooms 5. Coordination of the layout of the ductwork, piping, electric cables, etc. Retrofit, Remodeling, and Replacement Retrof t of an HVAC&R system must be tailored to the e xisting building and integrated with existing systems. Each retrof t project has a motivation which may be related to environment, safety and health, indoor environment, energy conservation, change of use, etc. Because of ozone depletion,

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CHAPTER ONE

the production of CFCs ceased in the United States on January 1, 1996. All refrigeration systems using CFC refrigerants will be replaced by or con verted to non-CFC refrigerant systems. F or the sake of safety , smoke control systems and stairwell pressurization are items to be considered. F or the sake of the occupants ’ health, an increase in the amount of outdoor air may be the right choice. If energy conservation is the f rst priority, the following items are among the many that may be considered: eff ciency of lighting; energy consumption of the f ans, pumps, chillers, and boilers; insulation of the building envelope; and the remodeling or replacement of fenestration. F or improving the thermal comfort of the occupant, an increase of the supply v olume f ow rate, an increase of cooling and heating capacity, and the installation of appropriate controls may be considered.

Engineer’s Quality Control In 1988, the American Consulting Engineers Council chose “People, Professionalism, Prof ts: A Focus on Quality ” as the theme of its annual con vention in New York City. Quality does not mean perfection. In HVAC&R system design, quality means functionally ef fective, for health, comfort, safety, energy conservation, and cost. Quality also implies the best job, using accepted professional practices and talent. Better quality al ways means fewer complaints and less litigation. Of course, it also requires additional time and higher design cost. This fact must be recognized by the owner and developer. The use of safety f actors allows for the unpredictable in design, installation, and operation. For example, a safety factor of 1.1 for the calculated cooling load and of 1.1 to 1.15 for the calculated total pressure drop in ductw ork is often used to tak e into account une xpected inferiority in f abrication and installation. An HVAC&R system should not be overdesigned by using a greater safety factor than is actually required. The initial cost of an o verdesigned HVAC&R system is always higher, and it is energy-ineff cient. When an HVAC&R system design is under pressure to reduce the initial cost, some avenues to be considered are as follows: 1. 2. 3. 4.

Select an optimum safety factor. Minimize redundancy, such as standby units. Conduct a detailed economic analysis for the selection of a better alternative. Calculate the space load, the capacity of the system, and the equipment requirements more precisely. 5. Adopt optimum di versity factors based on actual e xperience data observ ed from similar b uildings.

A diversity factor is def ned as the ratio of the simultaneous maximum load of a system to the sum of the indi vidual maximum loads of the subdi visions of a system. It is also called the simultaneous-use factor. For example, in load calculations for a coil, block load is used rather than the sum of zone maxima for sizing coils for AHUs installed with modulating control for the chilled w ater f ow rate. The need for a higher -quality design requires that engineers ha ve a better understanding of the basic principles, practical aspects, and updated technology of HV AC&R systems in order to a void overdesign or underdesign and to produce a satisfactory product.

Design of the Control System Controlling and maintaining the indoor en vironmental parameters within predetermined limits depends mainly on adequate equipment capacity and the quality of the control system. Ener gy can be saved when the systems are operated at part load with the equipment ’s capacity follo wing the system load accurately by means of capacity control.

INTRODUCTION

1.21

Because of the recent rapid change of HV AC&R controls from con ventional systems to ener gy management systems, to DDC with microprocessor intelligence, and then to open-protocol BACnet, many designers ha ve not k ept pace. In 1982, Haines did a surv ey on HV AC&R control system design and found that man y designers preferred to prepare a conceptual design and a sequence of operations and then to ask the representati ve of the control manuf acturer to design the control system. Only one-third of the designers designed the control system themselv es and ask ed the representative of the control manufacturer to comment on it. HVAC&R system control is a decisi ve factor in system performance. Man y of the troubles with HVAC&R arise from inadequate controls and/or their improper use. The designer should keep pace with the development of new control technology. He or she should be able to prepare the sequence of operations and select the best- f t control sequences for the controllers from a v ariety of the manufacturers that of fer equipment in the HV AC&R f eld. The designer may not be a specialist in the details of construction or of wiring diagrams of controllers or DDC modules, but he or she should be quite clear about the function and sequence of the desired operation, as well as the criteria for the sensors, controllers, DDC modules, and controlled devices. If the HVAC&R system designer does not perform these duties personally , preparation of a systems operation and maintenance manual with clear instructions w ould be diff cult. It would also be diff cult for the operator to understand the designer ’s intention and to operate the HVAC&R system satisfactorily.

Field Experience It is helpful for the designer to visit similar projects that ha ve been operated for more than 2 years and talk with the operator before initiating the design process. Such practice has man y advantages. 1. The designer can in vestigate the actual performance and ef fectiveness of the air conditioning and control systems that he or she intends to design. 2. According to the actual operating records, the designer can judge whether the system is o verdesigned, underdesigned, or the exact right choice. 3. Any complaints or problems that can be corrected may be identif ed. 4. The results of ener gy conservation measures can be e valuated from actual performance instead of theoretical calculations. 5. The designer can accumulate v aluable practical e xperience from the visit, even from the de f ciencies.

New Design Technologies Computer-aided design and drafting (CADD) and knowledge-based system (KBS) or expert system assisted design are prospecti ve new design technologies that ha ve been used in HV AC&R design. CADD is covered in detail in Sec. 1.12, and KBS is covered in Chap. 5.

1.10 DESIGN DOCUMENTS In the United States, the construction (installation) of air conditioning or HV AC&R systems is usually performed according to a contract between the o wner or de veloper and the installer , the contractor. The owner specif es the work or job that is committed to be accomplished within a time period. The contractor shall furnish and install equipment, ductwork, piping, instruments, and the related material of the HVAC&R system for a gi ven compensation. The construction of an air conditioning or HVAC&R system is usually a part of the construction of a building.

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Both dra wings and speci f cations are le gal documents. They are le gal because the designer conveys the requirements of the o wner or de veloper to the contractor through these documents. Drawings and specif cations def ne the work to be done by the contractor. They should be clear and should precisely and completely sho w the w ork to be accomplished. Dra wings and speci f cations are complements to each other. Things to be def ned should be shown or specif ed only once, either in drawing or in specif cation. Repetition often causes ambiguity and error.

Drawings The layout of an HVAC&R system and the locations and dimensions of its equipment, instruments, ducts, pipes, etc., are best shown and illustrated by drawings. HVAC&R drawings consist of mainly the following: ●

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Floor plans . System layout including plant room, fan rooms, mechanical room, ductwork, and pipelines is al ways illustrated on f oor plans. Each f oor has at least one f oor plan. HV AC&R f oor plans are always drawn over the same f oor plan of the architectural drawing. Detail drawings. These drawings show the details of a certain section of an HV AC&R system, or the detail of the installation of certain equipment, or the connection between equipment and ductwork or pipeline. Standard details are often used to save time. Sections and ele vations. Sectional dra wings are helpful to sho w the inner part of a section of a system, a piece of equipment, or a device. They are especially useful for places such as the plant room, fan room, and mechanical room where lots of equipment, ductwork, and pipelines are found. Elevations often show clearly the relationship between the HV AC&R components and the building structure. Piping dia gram. This diagram sho ws the piping layout of the w ater system(s) and the f ow o f water from the central plant to the HVAC&R equipment on each f oor. Air duct dia gram. This diagram illustrates the air duct layout as well as the air f ow from the air handling unit or packaged unit to the conditioned spaces on each f oor through space dif fusion devices. Control diagrams. These diagrams sho w the zone le vel control systems, each type of functional control system for air -handling units or packaged units, water systems, outdoor air v entilation systems, sequencing of compressors, network communication, etc. Equipment sc hedule. This schedule pro vides the quatity and performance characteristics of the equipment or device in tabulated form as drawings. Drawings are always available to the installer at any time b ut the speci f cations are not. This is why the equipment schedule should appear on drawings instead of inside the specif cation. Legends. Symbols and abbre viations are often de f ned in a le gend. ASHRAE has proposed a set of symbols in the Fundamentals handbook.

Sometimes, three-dimensional isometric drawings are necessay for piping and air duct diagrams. For a f oor plan, a scale of 1/8 in.  1 ft (1:100) is often used. The size of the dra wings should be selected according to the size of the project. Dra wing sheet sizes of 24  36 in., 30  42 in., and 36  48 in. (610  915 mm, 762  1067 mm, and 915  1219 mm) are widely adopted for lar ge projects. After completion, every drawing should be carefully checked for errors and omissions.

Specifications Detailed descriptions of equipment, instruments, ductwork, and pipelines, as well as performances, operating characteristics, and control sequences are better de f ned in speci f cations. Specif cations usually consist of the le gal contract between the o wner and the contractor , installer, or vendor, and

INTRODUCTION

1.23

the technical speci f cations that specify in detail the equipment and material to be used and ho w they are installed. The Construction Speci f cations Institute (CSI) has de veloped a format, called the Masterformat, for specif cations. This Masterformat is widely adopted by most HVAC&R construction projects. Masterformat promotes standardization and thereby f acilitates the retrie val of information and improves construction communication. The 1988 edition of Masterformat consists of 16 divisions: 01000 General Requirements, 02000 Site Work, 03000 Concrete, . . ., 14000 Conveying Systems, 15000 Mechanical, and 16000 Electrical. In mechanical, it is subdivided into the follo wing major sections: 15050 15250 15300 15400 15500 15550 15650 15750 15850 15880 15950 15990

Basic Mechanical Materials and Methods Mechanical Insulation Fire Protection Plumbing Heating, Ventilating, and Air Conditioning Heat Generation Refrigeration Heat Transfer Air Handling Air Distribution Controls Testing, Adjusting, and Balancing

Each of the abo ve-mentioned sections may contain: general considerations, equipment and material, and f eld installation. According to whether the w anted vendor is specif ed or not, specif cations can be classif ed into two categories: (1) performance specif cation, which depends only on the performance criteria, and (2) or-equal specif cation, which specif es the wanted vendor. Here are some recommendations for writing an HVAC&R specif cation: ●

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The required indoor en vironmental parameters to be maintained in the conditioned space during summer and winter outdoor design conditions, such as temperature, humidity, outdoor ventilation air rates, air cleanliness, sound level, and space pressure, shall be clearly speci f ed in the general consideration of section 15500. Use simple, direct, and clear language without repetition. All the terms must be well def ned and written in a consistent manner. Don’t write specif cations or refer to other w orks without having personal knowledge of the content or even understanding its meaning. All the specif cations must be tailored to f t the designed HVAC&R system. Never list an item that is not listed in the specif ed HVAC&R system, such as the return ducts in a fan-coil system.

1.11 CODES AND STANDARDS Codes are generally mandatory state or city laws or regulations that force the designer to create the design without violating human safety and welf are. State and city codes concerning structural inte grity, electrical safety, f re protection, and prevention of explosion of pressure vessels must be followed. Standards describe consistent methods of testing, specify conf rmed design guidelines, and recommend standard practices. Conformance to standards is usually v oluntary. Ho wever, if design

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CHAPTER ONE

criteria or system performance is not co vered in local codes, ASHRAE Standards become the vital document to assess the indoor air quality in a lawsuit against designer or contractor. Since the energy crisis of 1973, the U. S. federal go vernment and Congress ha ve published legislation which profoundly af fects the design and operation of HV AC&R systems. As mentioned before, in 1975, the Energy Policy and Conservation Act and in 1978 the National Ener gy Conservation Policy Act were enacted for ener gy conservation. Also in 1992, the Clean Air Act Amendments phase out the use of CFCs and rec ycling of refrigerants, to protect the ozone layer as well as to provide an acceptable indoor air quality. Among the HV AC&R related standards published by v arious institutions, the follo wing by ASHRAE and some other institutions have greatly inf uenced the design and operation of air conditioning systems: ●

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ASHRAE/IES Standard 90.1-1999, Energy Standard for Buildings Except Low-Rise Residential Buildings, which is often called the Ener gy Standard. This standard is under continuous maintenance by a standing standard committee (SSPC) for re gular publication of addenda or re visions. ASHRAE Standard 62-1999, Ventilation for Acceptable Indoor Air Quality, which is often called the Indoor Air Quality Standard. This standard is under continuous maintenance. ANSI/ASHRAE Standard 55-1992, Thermal Environmental Conditions for Human Occupancy ASHRAE Standard 15-1992, Safety Code for Mechanical Refrigeration. ANSI/ASHRAE Standard 34-1997, Number Designation and Safety Classification of Refri erants. This standard is under continuous maintenance. ANSI/ASHRAE Standard 135-1995, BACnet: A Data Communication Pr otocol for Building Automation and Control Networks. Air Conditioning and Refrigeration Institute, Standards for Unitary Equipment. National Fire Protection Association NFPA 90A, Standard for the Installation of Air Conditioning and Ventilating Systems. Sheet Metal and Air Conditioning Contractors ’ National Association, HVAC Duct Construction Standards — Metal and Flexible.

These standards are covered in detail in later chapters. Recently, many HVAC&R consulting f rms seek compliance with the International Or ganization for Standardization (ISO) 9000 for quality control. ISO is a specialized international agenc y for standardization, at present comprising the national standards institutions of about 90 countries. The American National Standards Institue (ANSI) is a member that represents the United States. The ISO 9000 series has now become the quality mangement standard worldwide. It set the guidelines for management participation, design control, review of specif cations, supplier monitoring, and services provided. The goal of ISO 9000 is to guarantee that products or the services pro vided by a manufacturer or an engineering consulting f rm are appropriate to the speci f cations and are within the tolerances agreed upon. The ISO 9000 series consists of f ve quality system standards: ISO 9000 Guidelines of the Selection and the Application of Quality Management and Quality Assurance Standards ISO 9001 Modeling of Quality Assurance in Design Development, Manufacturing, Installation and Servicing ISO 9002 Modeling of Quality Assurance in Production and Installation ISO 9003 Modeling of Quality Assurance in Final Inspection and Testing ISO 9004 Guidelines for the Implementation of Quality Mana gement and to Pr ovide Quality Systems There are no legal requirements to be registered to ISO 9000 standards. Meeting ISO 9000 standards does not mean that the f rm provides better service or products than those of nonre gistered companies.

INTRODUCTION

1.25

Many companies are assembling quality assurance programs which show that they are accepting compliance to ISO 9000 standards instead of re gistration. Such an arrangement will identify the ISO 9000 standard that can be best applied to your operation and also support your customer ’s ISO 9000 quality audit.

1.12 COMPUTER-AIDED DESIGN AND DRAFTING (CADD) Because personal computers (PCs) pro vide a lo w-cost tool for computations and graphics and owing to the a vailability of lots of design computer softw are during the 1980s, computer-aided design and drafting (CADD) for air conditioning systems has gro wn quickly in recent years. Today, computer software is often used to develop design documents, drawings, and specif cations in common use in engineering consulting f rms. According to the 1994 CADD Application and User Survey of design f rms in Engineering Systems (June 1994), “ . . . f rms with high producti vity reported that they perform 95 percent of projects on CADD.” In addition to the CADD, there is softw are available for computer -aided facilities management (CAFM) for operation and maintenance.

Features of CADD Compared with con ventional design and drafting, CADD of air conditioning systems has the following features: ●

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Different trade designers such as those in architecture, HVAC&R, plumbing, f re protection, and electrical engineers can mer ge their w ork in plot output f les to produce composite dra wings in different layers. Mechanical system design can integrate with equipment selection programs. Graphical model construction of air conditioning systems and subsystems in tw o-dimensional (2D) or three-dimensional (3D) presentation uses architectural and structural models as backgrounds. CADD links the engineering design models based on calculation and the graphical model based on drafting. It provides the ability to develop and compare the alternate design schemes quickly as well as the capability to redesign or to match the changes during construction promptly. CADD establishes database libraries for design and graphical models. It develops design documents including as-built drawings and equipment schedules.

A saving of design time up to 40 percent has been claimed by some engineering consulting

f rms.

Computer-Aided Design Current CADD for HVAC&R systems can be classi f ed into two categories: (1) engineering design and simulation, and (2) drafting graphical model construction. The software for engineering design can again be subdivided into the following groups: ●

Load calculations, energy simulation, and psychrometric analysis. Software for space load calculations could be a part of the ener gy simulation softw are. However, load calculations are often mainly used to determine the peak or block load at design conditions. These loads are the inputs for equipment selection. However, the software for energy simulation is used mainly to determine the optimum selection from dif ferent alternatives. Psychrometric analysis is sometimes a useful tool for load calculation and energy simulation.

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CHAPTER ONE ●

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Equipment selection. All large HVAC&R equipment manufacturers offer software, often called an electronic catalog, to help customers select their product according to the capacity , size, conf guration, and performance. Equipment schedule and specification . Software generates equipment schedules or CSI master format for specif cations. Price list. Software reports the price of base unit, accessories, and materials.

Computer-Aided Drafting (CAD) ●

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Software to reproduce architectural dra wings is the foundation of CADD. Graphical models of items such as ducts, pipes, equipment, and system components in HV AC&R systems are de veloped against backgrounds of architecture and structural models. Automated computer-aided drafting (AutoCAD) is the leading PC-based drafting softw are used by most design f rms to assist the designer during design with automated drawings, databases, and layering schemes. Software applications to develop duct and piping work layouts are the two primary CADs used in air conditioning system design. They can interf ace with architectural, electrical, and plumbing drawings through AutoCAD. CAD for duct and piping work can also have hydraulic modeling capacities. Tags and an HVAC&R equipment schedule including components and accessories can be produced as well. CAD for ducts and piping w ork is covered in detail in corresponding chapters. Software is also available to produce details of the refrigerating plant, heating plant, and fan room with accessories, duct, and piping layout. Man y manuf acturers also supply libraries of f les for AutoCAD users to input the details of their products into CAD drawings. HVAC&R CAD for ducts and piping w ork layout and details often uses graphical model construction. The elements of the graphical model are link ed to its associated information and stored in the databases. Model construction starts by locating an end de vice such as a dif fuser or terminal. Add the supply and return pipes from different layers. Software for graphical model construction has the features of cop ying, repeating, moving, rotating, and mirror-imaging which enable the designer to construct the graphical models quickly. After supply and return main ducts, equipment, and components are added, this completes the supply and return duct systems of this f oor. Duct and piping layouts can also be accomplished by dra wing a single-line layout on the CAD, and the software will convert to two-dimensional or three-dimensional drawings.

Software Requirements ●

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A CADD software tailored to a specif c HVAC&R design or drafting purpose is needed. Built-in error checking and error f nding are necessary features. Transferability of dra wings and data among dif ferent CADD systems is important. ASHRAE Handbook 1995, HVAC Applications, recommended that software interface to the most prominent formats: the Initial Graphics Exchange Speci f cation (IGES), the Standard Exchange F ormat (SEF), and the Data Exchange Format (DXF). While using the man y a vailable computer softw are programs for loads, energy, and hydraulic property calculations, one should verify and calibrate the results against f eld-measured values in order to impro ve the ef fectiveness of the computer softw are when the conditions and af fecting factors may be varied.

REFERENCES Amistadi, H., Design and Drawing: Software Review, Engineering Systems, June 1993, pp. 18 – 29. Amistadi, H., HVAC Product Software, Engineering Systems, January 1994, pp. 56 – 62.

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Amistadi, H., Software Review: HVACR Product Software, Engineering Systems, January 1995, pp. 60 – 69. Arnold, D., The Evolution of Modern Off ce Buildings and Air Conditioning, ASHRAE Journal, no. 6, 1999, pp. 40 – 54. ASHRAE, ASHRAE Handbook 1995, HVAC Applications, ASHRAE Inc., Atlanta, GA, 1995. Bengard, M., The Future of Design-Build, Engineered Systems, no. 12, 1999, pp. 60 – 64. Census Bureau, Snapshot Portrays Prof le of Heating and Cooling Units 1991, AC, Heating & Refrigeration News, Aug. 28, 1995, pp. 23 – 24. Coad, W. J., Courses and Cures for Building System Defects, Heating/Piping/Air Conditioning, February 1985, pp. 98 – 100. Coad, W. J., Safety Factors in HVAC Design, Heating/Piping/Air Conditioning, January 1985, pp. 199 – 203. Crandell, M. S., NIOSH Indoor Air Quality Investigations 1971 through 1987, Proceedings of the Air Pollution Control Association Annual Meeting, Dallas, 1987. Denny, R. J., The CFC Footprint, ASHRAE Journal, November 1987, pp. 24 – 28. DiIorio, E., and Jennett, Jr., E. J., High Rise Opts for High Tech, Heating/Piping/Air Conditioning, January 1989, pp. 83 – 87. Department of Energy/Energy Information Administration (DOE/EIA), Household Energy Consumption and Expenditures 1993, Part I: National Data, DOE/EIA-0321 Washington, DC, 1995. Department of Energy/Energy Information Administration (DOE/EIA), Nonresidential Building Energy Consumption Survey: Characteristics of Commercial Buildings, 1995, DOE/EIA-0246(95), Washington, DC, 1997. Department of Energy/Energy Information Administration (DOE/EIA), Nonresidential Building Energy Consumption Survey: Commercial Buildings Consumption and Expenditures 1995, DOE/EIA-0318(95), Washington, DC, 1998. Energy Information Administration, Commercial Buildings Characteristics 1992 Commercial Building Energy Consumption Survey, Washington, D.C., 1994. Faust, F. H., The Early Development of Self-Contained and Packaged Air Conditioner, ASHRAE Transactions, vol. 92, 1986, Part II B, pp. 353 – 360. Grant, W. A., From ’36 to ’56 — Air Conditioning Comes of Age, ASHVE Transactions, vol. 63, 1957, pp. 69 – 110. Guedes, P., Encyclopedia of Architectural Technology, section 4, 1st ed., McGraw-Hill, New York, 1979. Haines, R.W., How Are Control Systems Designed? Heating/Piping/Air Conditioning, February 1982, p. 94. Haines, R.W., Operating HVAC Systems, Heating/Piping/Air Conditioning, July 1984, p. 106. Haines, R.W., and Wilson, C. L., HVAC Systems Design Handbook, McGraw-Hill, New York, 1994. Houghten, F. C., Blackshaw, J. L., Pugh, E. M., and McDermott, P., Heat Transmission as Inf uenced by Heat Capacity and Solar Radiation, ASHVE Transactions, vol. 38, 1932, pp. 231 – 284. Kohloss, F. H., The Engineer’s Liability in Avoiding Air Conditioning System Overdesign, ASHRAE Transactions, vol. 87, 1983, Part I B, pp. 155 – 162. Korte, B., Existing Buildings: Vast HVAC Resource, Heating/ Piping/ Air Conditioning, March 1989, pp. 57 – 63. Korte, B., The Health of the Industry, Heating/Piping/and Air Conditioning, January 1996, pp. 69 – 70. Lewis, L. L., and Stacey, Jr., A. E., Air Conditioning the Hall of Congress, ASHVE Transactions, vol. 36, 1930, pp. 333 – 346. MacCraken, C. D., The Greenhouse Effect on ASHRAE, ASHRAE Journal, June 1989, pp. 52 – 54. Maynich, P., and Bettano, M., CADD Aids Fast-Track Hospital Expansion, Engineering Systems, June 1994, pp. 38 – 40. McClive, J. R., Early Development in Air Conditioning and Heat Transfer, ASHRAE Transactions, vol. 92, 1986, Part II B, pp. 361 – 365. Miller, A., Thompson, J. C., Peterson, R. E., and Haragan, D.R., Elements of Meteorology, 4th ed., Bell & Howell Co., Columbus, OH, 1983. Nagengast, B., The First Centrifugal Chiller: The German Connection, ASHRAE Journal, no. 1, 1998, pp. 18 – 19. Penny, T., and Althof, J., Trends in Commercial Buildings, Heating/Piping/Air Conditioning, September 1992, pp. 59 – 66.

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Rowland, F. S., The CFC Controversy: Issues and Answers, ASHRAE Journal, December 1992, pp. 20 – 27. Simens, J., A Case for Unitary Systems, Heating/Piping/Air Conditioning, May 1982, pp. 60, 68 – 73. Thomas, V. C., Mechanical Engineering Design Computer Programs for Buildings, ASHRAE Transactions, 1991, Part II, pp. 701 – 710. Turner, F., Industrial News: Study Shows 49% of Suits Resolved without Payment, ASHRAE Journal, April 1996, pp. 10 – 11. Whalen, J. M., An Organized Approach to Energy Management, Heating/Piping/Air Conditioning, September 1985, pp. 95 – 102. Wilson, L., A Case for Central Systems, Heating/Piping/Air Conditioning, May 1982, pp. 61 – 67.

CHAPTER 2

PSYCHROMETRICS 2.1 PSYCHROMETRICS 2.1 Moist Air 2.1 Equation of State of an Ideal Gas 2.2 Equation of State of a Real Gas 2.2 Calculation of the Properties of Moist Air 2.3 2.2 DALTON’S LAW AND THE GIBBSDALTON LAW 2.3 2.3 AIR TEMPERATURE 2.4 Temperature and Temperature Scales 2.4 Thermodynamic Temperature Scale 2.5 Temperature Measurements 2.6 2.4 HUMIDITY 2.7 Humidity Ratio 2.7 Relative Humidity 2.7 Degree of Saturation 2.8 2.5 PROPERTIES OF MOIST AIR 2.8 Enthalpy 2.8 Moist Volume 2.9 Density 2.10 Sensible Heat and Latent Heat 2.10 Specific Heat of Moist Air at Constant Pressure 2.10 Dew-Point Temperature 2.11 2.6 THERMODYNAMIC WET-BULB

TEMPERATURE AND THE WET-BULB TEMPERATURE 2.11 Ideal Adiabatic Saturation Process 2.11 Thermodynamic Wet-Bulb Temperature 2.12 Heat Balance of an Ideal Adiabatic Saturation Process 2.12 Psychrometer 2.12 Wet-Bulb Temperature 2.12 Relationship between Wet-Bulb Temperature and Thermodynamic Wet-Bulb Temperature 2.14 2.7 SLING AND ASPIRATION PSYCHROMETERS 2.14 2.8 HUMIDITY MEASUREMENTS 2.16 Mechanical Hygrometers 2.16 Electronic Hygrometers 2.16 Comparison of Various Methods 2.19 2.9 PSYCHROMETRIC CHARTS 2.19 2.10 DETERMINATION OF THERMODYNAMIC PROPERTIES ON PSYCHROMETRIC CHARTS 2.22 Computer-Aided Psychrometric Calculation and Analysis 2.25 REFERENCES 2.25

2.1 PSYCHROMETRICS Psychrometrics is the study of the thermodynamic properties of moist air . It is used e xtensively to illustrate and analyze the characteristics of various air conditioning processes and cycles.

Moist Air The surface of the earth is surrounded by a layer of air called the atmosphere, or atmospheric air . From the point of vie w of psychrometrics, the lower atmosphere, or homosphere, is a mixture of dry air (including various contaminants) and water vapor, often known as moist air. The composition of dry air is comparati vely stable. It v aries slightly according to geographic location and from time to time. The approximate composition of dry air by v olume percent is the following:

2.1

2.2

CHAPTER TWO

Nitrogen Oxygen Argon Carbon dioxide Other gases such as neon, sulfur dioxide, etc.

78.08 20.95 0.93 0.03 0.01

The amount of w ater v apor present in moist air at a temperature range of 0 to 100°F ( 17.8 to 37.8°C) varies from 0.05 to 3 percent by mass. It has a significant influence on the characteristics moist air. Water vapor is lighter than air . A cloud in the sk y is composed of microscopic beads of liquid water that are surrounded by a thin layer of w ater vapor. These layers gi ve the cloud the needed buoyancy to float in the ai .

Equation of State of an Ideal Gas The equation of state of an ideal gas indicates the relationship between its thermodynamic properties, or pv  RTR

(2.1)

where p  pressure of gas, psf (Pa) v  specific olume of gas, ft3 /lb (m3 /kg) R  gas constant, ftlbf /lbm °R (J/kgK) TR  absolute temperature of gas, °R (K) Since v  V/m, then Eq. (2.1) becomes pv  mRTR

(2.2)

where V  total volume of gas, ft3 (m3) m  mass of gas, lb (kg) Using the relationship m  nM, and R  Ro /M, we can write Eq. (2.2) as pv  nRoTR

(2.3)

where n  number of moles, mol M  molecular weight Ro  universal gas constant, ftlbf /lbm °R (J/molK) Equation of State of a Real Gas A modified form of the equation of state for a real gas can be xpressed as pv  1  Ap  Bp 2  Cp 3      Z RTR

(2.4)

where A, B, C,     virial coefficients and Z  compressibility factor. The compressibility factor Z illustrates the degree of deviation of the behavior of the real gas, moist air, from the ideal gas due to the following: 1. Effect of air dissolved in water

PSYCHROMETRICS

2.3

2. Variation of the properties of water vapor attributable to the effect of pressure 3. Effect of intermolecular forces on the properties of water vapor itself For an ideal gas, Z  1. According to the information published by the former National Bureau of Standards of the United States, for dry air at standard atmospheric pressure (29.92 in. Hg, or 760 mm Hg) and a temperature of 32 to 100 °F (0 to 37.8 °C) the maximum de viation is about 0.12 percent. F or w ater vapor in moist air under saturated conditions at a temperature of 32 to 100 °F (0 to 37.8°C), the maximum deviation is about 0.5 percent. Calculation of the Properties of Moist Air The most e xact calculation of the thermodynamic properties of moist air is based on the formulations developed by Hyland and Wexler of the U.S. National Bureau of Standards. The psychrometric chart and tables of ASHRAE are constructed and calculated from these formulations. Calculations based on the ideal gas equations are the simplest and can be easily formulated. According to the analysis of Nelson and Pate, at a temperature between 0 and 100°F (17.8 and 37.8°C), calculations of enthalpy and speci f c volume using ideal gas equations sho w a maximum de viation of 0.5 percent from the e xact calculations by Hyland and Wexler. Therefore, ideal gas equations will be used in this text for the formulation and calculation of the thermodynamic properties of moist air. Although air contaminants may seriously af fect the health of occupants of the air conditioned space, they have little effect on the thermodynamic properties of moist air since their mass concentration is low. For simplicity, moist air is always considered as a binary mixture of dry air and water vapor during the analysis and calculation of its properties.

2.2 DALTON’S LAW AND THE GIBBS-DALTON LAW Dalton’s law shows that for a mixture of gases occup ying a given volume at a certain temperature, the total pressure of the mixture is equal to the sum of the partial pressures of the constituents of the mixture, i.e., pm  p1  p2    

(2.5)

where pm  total pressure of mixture, psia (Pa) p1, p2, . . .  partial pressure of constituents 1, 2, . . . , psia (Pa) The partial pressure e xerted by each constituent in the mixture is independent of the e xistence of other gases in the mixture. Figure 2.1 sho ws the variation of mass and pressure of dry air and w ater vapor, at an atmospheric pressure of 14.697 psia (101,325 P a) and a temperature of 75 °F (23.9°C). The principle of conservation of mass for nonnuclear processes gives the following relationship: mm  ma  mw

(2.6)

where mm  mass of moist air, lb (kg) ma  mass of dry air, lb (kg) mw  mass of water vapor, lb (kg) Applying Dalton’s law for moist air, we have pat  pa  pw where pat  atmospheric pressure or pressure of the outdoor moist air, psia (Pa) pa  partial pressure of dry air, psia (Pa) pw  partial pressure of water vapor, psia (Pa)

(2.7)

2.4

CHAPTER TWO

FIGURE 2.1 Mass and pressure of dry air, water vapor, and moist air.

Dalton’s la w is based on e xperimental results. It is more accurate for gases at lo w pressures. Dalton’s law can be further e xtended to state the relationship of the internal ener gy, enthalpy, and entropy of the gases in a mixture as the Gibbs-Dalton law: mm um  m1u1  m2 u2     mm hm  m1h1  m2 h2    

(2.8)

mm sm  m1s1  m2 s2     where mm  mass of gaseous mixture, lb (kg) m1, m2, . . .  mass of the constituents, lb (kg) um  specif c internal energy of gaseous mixture, Btu/lb (kJ/kg) u1, u2, . . .  specif c internal energy of constituents, Btu/lb (kJ/kg) hm  specif c enthalpy of gaseous mixture, Btu/lb (kJ/kg) h1, h2, . . .  specif c enthalpy of constituents, Btu/lb (kJ/kg) sm  specif c entropy of gaseous mixture, Btu/lb°R (kJ/kgK) s1, s2, . . .  specif c entropy of constituents, Btu/lb °R (kJ/kg K)

2.3 AIR TEMPERATURE Temperature and Temperature Scales The temperature of a substance is a measure of ho w hot or cold it is. Two systems are said to ha ve equal temperatures only if there is no change in an y of their observ able thermal characteristics when they are brought into contact with each other . Various temperature scales commonly used to measure the temperature of various substances are illustrated in Fig. 2.2. In con ventional inch-pound (I-P) units, at a standard atmospheric pressure of 14.697 psia (101,325 Pa), the Fahrenheit scale has a freezing point of 32 °F (0°C) at the ice point, and a boiling point of 212°F (100°C). For the triple point with a pressure of 0.08864 psia (611.2 P a), the magnitude on the Fahrenheit scale is 32.018°F (0.01°C). There are 180 divisions, or degrees, between the boiling and freezing points in the F ahrenheit scale. In the International System of Units (SI units), the Celsius or Centigrade scale has a freezing point of 0 °C and a boiling point of 100 °C. There are

PSYCHROMETRICS

2.5

FIGURE 2.2 Commonly used temperature scales.

100 divisions between these points. The triple point is at 0.01°C. The conversion from Celsius scale to Fahrenheit scale is as follows: °F  1.8(°C )  32

(2.9)

For an ideal gas, at TR  0, the gas would have a vanishing specif c volume. Actually, a real gas has a negligible molecular volume when TR approaches absolute zero. A temperature scale that includes absolute zero is called an absolute temperature scale. The Kelvin absolute scale has the same boiling-freezing point division as the Celsius scale. At the freezing point, the Kelvin scale is 273.15 K. Absolute zero on the Celsius scale is 273.15°C. The Rankine absolute scale di vision is equal to that of the F ahrenheit scale. The freezing point is 491.67 °R. Similarly, absolute zero is  459.67°F on the Fahrenheit scale. Conversions between Rankine and Fahrenheit and between Kelvin and Celsius systems are R  459.67  °F

(2.10)

K  273.15  °C

(2.11)

Thermodynamic Temperature Scale On the basis of the second la w of thermodynamics, one can establish a temperature scale that is independent of the w orking substance and that pro vides an absolute zero of temperature; this is called a thermodynamic temper ature scale . The thermodynamic temperature T must satisfy the following relationship: TR Q  TRo Qo

(2.12)

where Q  heat absorbed by reversible engine, Btu/h (kW) Qo  heat rejected by reversible engine, Btu/h (kW) TR  temperature of heat source of reversible engine, °R (K) TRo  temperature of heat sink of reversible engine, °R (K) Two of the ASHRAE basic tables, “ Thermodynamic Properties of Moist Air” and “Thermodynamic Properties of Water at Saturation,” in ASHRAE Handbook 1993, Fundamentals, are based on the thermodynamic temperature scale.

2.6

CHAPTER TWO

Temperature Measurements During the measurement of air temperatures, it is important to recognize the meaning of the terms accuracy, precision, and sensitivity. 1. Accuracy is the ability of an instrument to indicate or to record the true v alue of the measured quantity. The error indicates the degree of accuracy. 2. Precision is the ability of an instrument to give the same reading repeatedly under the same conditions. 3. Sensitivity is the ability of an instrument to indicate change of the measured quantity. Liquid-in-glass instruments, such as mercury or alcohol thermometers, were commonly used in the early days for air temperature measurements. In recent years, many liquid-in-glass thermometers have been replaced by remote temperature monitoring and indication systems, made possible by sophisticated control systems. A typical air temperature indication system includes sensors, amplif ers, and an indicator. Sensors. Air temperature sensors needing higher accurac y are usually made from resistance temperature detectors (R TDs) made of platinum, palladium, nickel, or copper wires. The electrical resistance of these resistance thermometers characteristically increases when the sensed ambient air temperature is raised; i.e., they have a positi ve temperature coef f cient . In man y engineering applications, the relationship between the resistance and temperature can be given by R  R 32(1   T )



R 212  R 32 180 R 32

(2.13)

where R  electric resistance,  R32, R212  electric resistance, at 32 and 212°F (0 and 100°C), respectively,  T  temperature, °F (°C) The mean temperature coef f cient  for several types of metal wires often used as R TDs is sho wn below:

Platinum Palladium Nickel Copper

Measuring range,°F

,  / °F

400 to 1350 400 to 1100 150 to 570 150 to 400

0.00218 0.00209 0.0038 0.0038

Many air temperature sensors are made from thermistors of sintered metallic oxides, i.e., semiconductors. They are available in a lar ge variety of types: beads, disks, washers, rods, etc. Thermistors have a ne gative temperature coef f cient. Their resistance decreases when the sensed air temperature increases. The resistance of a thermistor may drop from approximately 3800 to 3250  when the sensed air temperature increases from 68 to 77 °F (20 to 25 °C). Recently de veloped high-quality thermistors are accurate, stable, and reliable. Within their operating range, commercially a vailable thermistors will match a resistance-temperature curv e within approximately 0.1 °F (0.056 °C). Some manufacturers of thermistors can supply them with a stability of 0.05 °F (0.028°C) per year. For direct digital control (DDC) systems, the same sensor is used for both temperature indication, or monitoring, and temperature control. In DDC systems, RTDs with positi ve temperature coef f cient are widely used.

PSYCHROMETRICS

2.7

Amplifier(s) The measured electric signal from the temperature sensor is ampli f ed at the solid state amplif er to produce an output for indication. The number of ampli f ers is matched with the number of the sensors used in the temperature indication system. Indicator. An analog-type indicator , one based on directly measurable quantities, is usually a moving coil instrument. For a digital-type indicator, the signal from the ampli f er is compared with an internal reference voltage and converted for indication through an analog-digital transducer.

2.4 HUMIDITY Humidity Ratio The humidity ratio of moist air w is the ratio of the mass of water vapor mw to the mass of dry air ma contained in the mixture of the moist air, in lb/lb (kg/kg). The humidity ratio can be calculated as w

mw ma

(2.14)

Since dry air and water vapor can occupy the same volume at the same temperature, we can apply the ideal gas equation and Dalton’s law for dry air and water vapor. Equation (2.14) can be rewritten as w 

mw pwVR aTR pw Ra   ma PaVR wTR R w pat  pw pw pw 53.352  0.62198 85.778 pat  pw pat  pw

(2.15)

where Ra, Rw  gas constant for dry air and w ater vapor, respectively, ftlbf /lbm°R(J/kgK). Equation (2.15) is e xpressed in the form of the ratio of pressures; therefore, pw and pat must have the same units, either psia or psf (Pa). For moist air at saturation, Eq. (2.15) becomes ws  0.62198

pws pat  pws

(2.16)

where pws  pressure of water vapor of moist air at saturation, psia or psf (Pa). Relative Humidity The relative humidity  of moist air, or RH, is def ned as the ratio of the mole fraction of w ater vapor xw in a moist air sample to the mole fraction of the w ater vapor in a saturated moist air sample xws at the same temperature and pressure. This relationship can be expressed as



xw  x ws T,p

(2.17)

And, by def nition, the following expressions may be written: xw 

nw na  nw

(2.18)

x ws 

n ws n a  n ws

(2.19)

2.8

CHAPTER TWO

where na  number of moles of dry air, mol nw  number of moles of water vapor in moist air sample, mol nws  number of moles of water vapor in saturated moist air sample, mol Moist air is a binary mixture of dry air and w ater vapor; therefore, we f nd that the sum of the mole fractions of dry air xa and water vapor xw is equal to 1, that is, xa  xw  1

(2.20)

If we apply ideal gas equations pwV  nwRoTR and paV  naRoTR, by substituting them into Eq. (2.19), then the relative humidity can also be expressed as



pw  pws T,p

(2.21)

The water vapor pressure of saturated moist air pws is a function of temperature T and pressure p, which is slightly dif ferent from the saturation pressure of w ater vapor ps. Here ps is a function of temperature T only. Since the difference between pws and ps is small, it is usually ignored. Degree of Saturation The degree of saturation is def ned as the ratio of the humidity ratio of moist air w to the humidity ratio of the saturated moist air ws at the same temperature and pressure. This relationship can be expressed as



w  ws T,p

(2.22)

Since from Eqs. (2.15), (2.20), and (2.21) w  0.62198 xw /xa and ws  0.62198 xws /xa, Eqs. (2.20), (2.21), and (2.22) can be combined, so that



1  (1  )x ws



1  (1  )( pws /pat )

(2.23)

In Eq. (2.23), pws

pat ; therefore, the difference between  and is small. Usually, the maximum difference is less than 2 percent.

2.5 PROPERTIES OF MOIST AIR Enthalpy The difference in specif c enthalpy h for an ideal gas, in Btu /lb (kJ /kg), at a constant pressure can be def ned as h  cp (T2  T1)

(2.24)

where cp  specif c heat at constant pressure, Btu / lb °F (kJ/kgK) T1, T2  temperature of ideal gas at points 1 and 2, °F (°C) As moist air is approximately a binary mixture of dry air and w ater vapor, the enthalpy of the moist air can be evaluated as h  ha  Hw

(2.25)

PSYCHROMETRICS

2.9

where ha and Hw are, respectively, enthalpy of dry air and total enthalp y of w ater vapor, in Btu /lb (kJ/kg). The following assumptions are made for the enthalpy calculations of moist air: 1. 2. 3. 4. 5.

The ideal gas equation and the Gibbs-Dalton law are valid. The enthalpy of dry air is equal to zero at 0°F (17.8°C). All water vapor contained in the moist air is vaporized at 0°F (17.8°C). The enthalpy of saturated water vapor at 0°F (17.8°C) is 1061 Btu/lb (2468 kJ/kg). For convenience in calculation, the enthalpy of moist air is taken to be equal to the enthalpy of a mixture of dry air and water vapor in which the amount of dry air is exactly equal to 1 lb (0.454 kg). Based on the preceeding assumptions, the enthalpy h of moist air can be calculated as h  ha  whw

(2.26)

where hw  specif c enthalpy of water vapor, Btu/lb (kJ /kg). In a temperature range of 0 to 100 °F (17.8 to 37.8°C), the mean value for the speci f c heat of dry air can be tak en as 0.240 Btu /lb °F (1.005 kJ / kg K). Then the specif c enthalpy of dry air ha is given by ha  cpd T  0.240 T

(2.27)

where cpd  specif c heat of dry air at constant pressure, Btu/lb °F (kJ/kgK) T  temperature of dry air, °F (°C) The specif c enthalpy of water vapor hw at constant pressure can be approximated as hw  hg0  cpsT

(2.28)

where hg0  specif c enthalpy of saturated water vapor at 0°F (17.8°C) — its value can be taken as 1061 Btu /lb (2468 kJ/kg) cps  specif c heat of water vapor at constant pressure, Btu/lb °F (kJ/kgK) In a temperature range of 0 to 100°F (17.8 to 37.8°C), its value can be taken as 0.444 Btu/lb °F (1.859 kJ / kg K). Then the enthalpy of moist air can be evaluated as h  cpd T  w(hg0  cps T)  0.240 T  w(1061  0.444 T)

(2.29)

Here, the unit of h is Btu /lb of dry air (kJ /kg of dry air). F or simplicity, it is often e xpressed as Btu/lb (kJ/kg).

Moist Volume The moist volume of moist air v, ft3 /lb (m 3 /kg), is def ned as the v olume of the mixture of the dry air and water vapor when the mass of the dry air is exactly equal to 1 lb (1 kg), that is, v

V ma

(2.30)

where V  total volume of mixture, ft3 (m3) ma  mass of dry air, lb (kg) In a moist air sample, the dry air, water vapor, and moist air occup y the same v olume. If we apply the ideal gas equation, then v

R aTR V  ma pat  pw

(2.31)

2.10

CHAPTER TWO

where pat and pw are both in psf (P a). From Eq. (2.15), pw  patw/(w  0.62198). Substituting this expression into Eq. (2.31) gives v

R aTR(1  1.6078 w) Pat

(2.32)

According to Eq. (2.32), the volume of 1 lb (1 kg) of dry air is al ways smaller than the v olume of the moist air when both are at the same temperature and the same atmospheric pressure.

Density Since the enthalp y and humidity ratio are al ways related to a unit mass of dry air , for the sak e of consistency, air density a, in lb /ft3 (kg/m3), should be def ned as the ratio of the mass of dry air to the total volume of the mixture, i.e., the reciprocal of moist volume, or

a 

ma 1  V v

(2.33)

Sensible Heat and Latent Heat Sensible heat is that heat ener gy associated with the change of air temperature between tw o state points. In Eq. (2.29), the enthalpy of moist air calculated at a datum state 0 °F ( 17.8°C) can be divided into two parts: h  (cpd  wcps)T  whg0

(2.34)

The f rst term on the right-hand side of Eq. (2.34) indicates the sensible heat. It depends on the temperature T above the datum 0°F (17.8°C). Latent heat hfg (sometimes called hig) is the heat ener gy associated with the change of the state of water vapor. The latent heat of v aporization denotes the latent heat required to v aporize liquid water into water vapor. Also, the latent heat of condensation indicates the latent heat to be remo ved in the condensation of w ater vapor into liquid water. When moisture is added to or remo ved from a process or a space, a corresponding amount of latent heat is al ways involved, to vaporize the water or to condense it. In Eq. (2.34), the second term on the right-hand side, whg0 , denotes latent heat. Both sensible and latent heat are expressed in Btu/lb (kJ/kg) of dry air.

Specific Heat of Moist Air at Constant Pressure The specif c heat of moist air at constant pressure cpa is def ned as the heat required to raise its temperature 1°F (0.56°C) at constant pressure. In (inch-pound) I-P units, it is e xpressed as Btu /lb°F (in SI units, as J/kgK). In Eq. (2.34), the sensible heat of moist air qsen, Btu/h (W), is represented by qsen  m˙a(cpd  wcps)T  m˙acpaT where mass m˙a 

(2.35)

f ow rate of moist air, lb/h (kg/s). Apparently cpa  cpd  wcps

(2.36)

Since cpd and cps are both a function of temperature, cpa is also a function of temperature and, in addition, a function of the humidity ratio. For a temperature range of 0 to 100 °F (17.8 to 37.8 °C), cpd can be tak en as 0.240 Btu /lb °F (1005 J / kgK) and cps as 0.444 Btu /lb°F (1859 J /kgK). Most of the calculations of cpa(T2  T1)

PSYCHROMETRICS

2.11

have a range of w between 0.005 and 0.010 lb /lb (kg /kg). Taking a mean v alue of w  0.0075 lb/lb (kg/kg), we f nd that cpa  0.240  0.0075 0.444  0.243 Btu/lb °F (1020 J/kgK) Dew-Point Temperature The dew-point temperature Tdew is the temperature of saturated moist air of the same moist air sample, having the same humidity ratio, and at the same atmospheric pressure of the mixture pat. Two moist air samples at the same Tdew will have the same humidity ratio w and the same partial pressure of water vapor pw. The dew-point temperature is related to the humidity ratio by ws ( pat, Tdew)  w

(2.37)

where ws  humidity ratio of saturated moist air, lb/lb (kg /kg). At a specif c atmospheric pressure, the dew-point temperature determines the humidity ratio w and the w ater vapor pressure pw of the moist air.

2.6 THERMODYNAMIC WET-BULB TEMPERATURE AND THE WET-BULB TEMPERATURE Ideal Adiabatic Saturation Process If moist air at an initial temperature T1, humidity ratio w1, enthalpy h1, and pressure p f ows over a water surface of in f nite length in a well-insulated chamber , as shown in Fig. 2.3, liquid water will evaporate into water vapor and will disperse in the air. The humidity ratio of the moist air will gradually increase until the air can absorb no more moisture. As there is no heat transfer between this insulated chamber and the surroundings, the latent heat required for the e vaporation of w ater will come from the sensible heat released by the moist air . This process results in a drop in temperature of the moist air. At the end of this evaporation process, the moist air is al ways saturated. Such a process is called an ideal adiabatic satur ation process, where an adiabatic process is def ned as a process without heat transfer to or from the process.

FIGURE 2.3 Ideal adiabatic saturation process.

2.12

CHAPTER TWO

Thermodynamic Wet-Bulb Temperature For an y state of moist air , there e xists a thermodynamic wet-b ulb temperature T* that e xactly equals the saturated temperature of the moist air at the end of the ideal adiabatic saturation process at constant pressure. Applying a steady f ow energy equation, we have h 1  (w* s  w1)h* w  h* s

(2.38)

where h 1, h* s  enthalpy of moist air at initial state and enthalpy of saturated air at end of ideal adiabatic saturation process, Btu/lb (kJ/kg) w1,w* s  humidity ratio of moist air at initial state and humidity ratio of saturated air at end of ideal adiabatic saturation process, lb/lb (kg/kg) h*w  enthalpy of water as it is added to chamber at a temperature T*, Btu/lb (kJ/kg) The thermodynamic wet-b ulb temperature T*, °F ( °C), is a unique property of a gi ven moist air sample that depends only on the initial properties of the moist air — w1, h1 and p. It is also a f ctitious property that only hypothetically exists at the end of an ideal adiabatic saturation process. Heat Balance of an Ideal Adiabatic Saturation Process When water is supplied to the insulation chamber at a temperature T* in an ideal adiabatic saturation process, then the decrease in sensible heat due to the drop in temperature of the moist air is just equal to the latent heat required for the e vaporation of w ater added to the moist air . This relationship is given by cpd(T1  T*)  cpsw1(T1  T*)  (w* s  w1)h* fg

(2.39)

where T1  temperature of moist air at initial state of ideal adiabatic saturation process, °F (°C) h*fg  latent heat of vaporization at thermodynamic wet-bulb temperature, Btu/lb (J/kg) Since cpa  cpd  w1cps, we f nd, by rearranging the terms in Eq. (2.39), cpa w*s  w1  T1  T* h*fg

(2.40)

Also T*  T1 

(w* s  w1)h* fg cpa

(2.41)

Psychrometer A psychrometer is an instrument that permits one to determine the relati ve humidity of a moist air sample by measuring its dry-b ulb and wet-b ulb temperatures. Figure 2.4 sho ws a psychrometer , which consists of tw o thermometers. The sensing b ulb of one of the thermometers is al ways kept dry. The temperature reading of the dry bulb is called the dry-bulb temperature. The sensing bulb of the other thermometer is wrapped with a piece of cotton wick, one end of which dips into a cup of distilled w ater. The surf ace of this b ulb is al ways wet; therefore, the temperature that this b ulb measures is called the wet-bulb temperature. The dry bulb is separated from the wet b ulb by a radiation-shielding plate. Both dry and wet bulbs are cylindrical. Wet-Bulb Temperature When unsaturated moist air f ows over the wet bulb of the psychrometer, liquid water on the surface of the cotton wick e vaporates, and as a result, the temperature of the cotton wick and the wet b ulb

PSYCHROMETRICS

2.13

FIGURE 2.4 A psychrometer.

drops. This depressed wet-bulb reading is called the wet-bulb temperature T, and the difference between the dry-bulb and wet-bulb temperatures is called the wet-bulb depression. Let us ne glect the conduction along the thermometer stems to the dry and wet b ulbs and also assume that the temperature of the w ater on the cotton wick is equal to the wet-b ulb temperature of the moist air . Since the heat transfer from the moist air to the cotton wick is e xactly equal to the latent heat required for vaporization, then, at steady state, the heat and mass transfer per unit area of the wet-bulb surface can be calculated as h c(T  T)  h r(Tra  T)  h d (ws  w1)hfg

(2.42)

where hc, hr  mean convective and radiative heat transfer coeff cients, Btu/h  ft2 °F (W/m2 K) hd  mean convective mass-transfer coeff cient, lb/hft2 (kg/s m2) T  temperature of undisturbed moist air at a distance from wet bulb, °F (°C) T  wet-bulb temperature, °F (°C) Tra  mean radiant temperature, °F (°C) w1, ws  humidity ratio of moist air and saturated air f lm at surface of cotton wick, lb/lb (kg/kg) hfg  latent heat of vaporization at wet-bulb temperature, Btu/lb (J/kg) Based on the correlation of cross- f ow forced convective heat transfer for a c ylinder, NuD  C Ren Pr0.333, and based on the analogy between con vective heat transfer and convective mass transfer, the following relationship holds: hd  hc /(cpaLe0.6667 ). Here, Nu is the Nusselt number , Re the Reynolds number, and Le the Lewis number. Also C is a constant. Substituting this relationship into Eq. (2.42), we have w1  ws  K(T  T)

(2.43)

In Eq. (2.43), K represents the wet-bulb constant. It can be calculated as K 

cpa Le 0.6667 hfg

1 

h r(Tra  T) h c(T  T)



(2.44)

2.14

CHAPTER TWO

The term T  T  indicates the wet-bulb depression. Combining Eqs. (2.43) and (2.44) then gives cpaLe 0.667 ws  w1  T  T hfg

1 

h r(Tra  T ) h c(T  T)



(2.45)

Relationship between Wet-Bulb Temperature and Thermodynamic Wet-Bulb Temperature Wet-bulb temperature is a function not only of the initial state of moist air , but also of the rate of heat and mass transfer at the wet b ulb. Comparing Eq. (2.40) with Eq. (2.45), we f nd that the wetbulb temperature measured by using a psychrometer is equal to thermodynamic wet-b ulb temperature only when the following relationship holds:



Le 0.6667 1 

h r(Tra  T) h c(T  T)

1

(2.46)

2.7 SLING AND ASPIRATION PSYCHROMETERS Sling and aspiration psychrometers determine the relati ve humidity through the measuring of the dry- and wet-bulb temperatures. A sling psychrometer with two bulbs, one dry and the other wet, is shown in Fig. 2.5a. Both dry and wet bulbs can be rotated around a spindle to produce an airf ow over the surfaces of the dry and wet bulbs at an air v elocity of 400 to 600 fpm (2 to 3 m /s). Also a shield plate separates the dry and wet bulbs and partly protects the wet bulb against surrounding radiation. An aspiration psychrometer that uses a small motor -driven fan to produce an air current f owing over the dry and wet b ulbs is illustrated in Fig. 2.5 b. The air velocity over the bulbs is usually k ept at 400 to 800 fpm (2 to 4 m /s). The dry and wet bulbs are located in separate compartments and are entirely shielded from the surrounding radiation. When the space dry-b ulb temperature is within a range of 75 to 80 °F (24 to 27 °C) and the space wet-b ulb temperature is between 65 and 70 °F (18 and 21°C), the following wet-bulb constants K can be used to calculate the humidity ratio of the moist air: Aspiration psychrometer K  0.000206

1/°F

K  0.000218

1/°F

Sling psychrometer

After the psychrometer has measured the dry- and wet-b ulb temperatures of the moist air , the humidity ratio w can be calculated by Eq. (2.43). Since the saturated w ater vapor pressure can be found from the psychrometric table, the relati ve humidity of moist air can be e valuated through Eq. (2.15). According to the analysis of Threlkeld (1970), for a wet-bulb diameter of 0.1 in. (2.5 mm) and an air v elocity f owing over the wet b ulb of 400 fpm (2 m /s), if the dry-b ulb temperature is 90 °F (32.2°C) and the wet-b ulb temperature is 70 °F (21.1 °C ), then ( T  T*)/(T  T) is about 2.5 percent. Under the same conditions, if a sling psychrometer is used, then the de viation may be reduced to about 1 percent. If the air v elocity f owing over the wet b ulb exceeds 400 fpm (2 m /s), there is no signif cant reduction in the deviation. Distilled water must be used to wet the cotton wick for both sling and aspiration psychrometers. Because dusts contaminate them, cotton wicks should be replaced re gularly to provide a clean surface for evaporation.

PSYCHROMETRICS

FIGURE 2.5 Sling and aspiration psychrometers (a) Sling psychrometer; (b) aspiration psychrometer.

2.15

2.16

CHAPTER TWO

2.8 HUMIDITY MEASUREMENTS Humidity sensors used in HV AC&R for direct humidity indication or operating controls are separated into the following categories: mechanical hygrometers and electronic hygrometers.

Mechanical Hygrometers Mechanical hygrometers operate on the principle that hygroscopic materials e xpand when the y absorb w ater v apor or moisture from the ambient air . They contract when the y release moisture to the surrounding air . Such hygroscopic materials include human and animal hairs, plastic polymers lik e n ylon ribbon, natural f bers, wood, etc. When these materials are link ed to mechanical linkages or electric transducers that sense the change in size and con vert it into electric signals, the results in these de vices can be calibrated to yield direct relati ve-humidity measurements of the ambient air.

Electronic Hygrometers There are three types of electronic hygrometers: Dunmore resistance hygrometers, ion-exchange resistance hygrometers, and capacitance hygrometers. Dunmore Resistance Hygrometer. In 1938, Dunmore of the National Bureau of Standards de veloped the f rst lithium chloride resistance electric hygrometer in the United States. This instrument depends on the change in resistance between tw o electrodes mounted on a hygroscopic material. Figure. 2.6a shows a Dunmore resistance sensor . The electrodes could be, e.g., a double-threaded winding of noble-metal wire mounted on a plastic c ylinder coated with hygroscopic material. The wires can also be in a grid-type arrangement with a thin f lm of hygroscopic material bridging the gap between the electrodes. At a speci f c temperature, electric resistance decreases with increasing humidity . Because the response is signi f cantly in f uenced by temperature, the results are often indicated by a series of isothermal curv es. Relati ve humidity is generally used as the humidity parameter , for it must be controlled in the indoor en vironment. Also the electrical response is more nearly a function of relative humidity than of the humidity ratio. The time response to accomplish a 50 percent change in relative humidity varies directly according to the air v elocity f owing o ver the sensor and also is in versely proportional to the saturated vapor pressure. If a sensor has a response time of 10 s at 70°F (21°C), it might need a response time of 100 s at 10°F (12°C). Because of the steep variation of resistance corresponding to a change in relati ve humidity, each of the Dunmore sensors only co vers a certain range of relati ve-humidity measurements. A set of several Dunmore sensors is usually needed to measure relative humidity between 1 percent and 100 percent. A curve for output, in direct-current (dc) v olts, versus relative humidity is sho wn in Fig. 2.6 b for a typical Dunmore sensor . It co vers a measuring range of 10 to 80 percent, which is usually suff cient for the indication of relati ve humidity for a comfort air conditioned space. This typical Dunmore sensor has an accurac y of 3 percent when the relati ve humidity varies between 10 and 60 percent at a temperature of 70 °F (21°C, see Fig. 2.6b). Its accuracy reduces to 4 percent when the relative humidity is in a range between 60 and 80 percent at the same temperature. In addition to lithium chloride, lithium bromide is sometimes used as the sensor. Ion-Exchange Resistance Hygrometer. The sensor of a ion-exchange resistance electric hygrometer is composed of electrodes mounted on a baseplate and a high-polymer resin f lm, used as a humidity-sensing material, cross-linking the electrodes as sho wn in Fig. 2.7 a and b. Humidity is

PSYCHROMETRICS

2.17

(a)

measured by the change in resistance between the electrodes. When the salt contained in the humidity-sensitive material bridging the electrodes becomes ion-conductive because of the presence of water vapor in the ambient air, mobile ions in the polymer film are formed. The higher the relative humidity of the ambient air, the greater the ionization becomes, and therefore, the greater the concentration of mobile ions. On the other hand, lower relative humidity reduces the ionization and results in a lower concentration of mobile ions. The resistance of the humidity-sensing material reflects the change of the relative humidity of the ambient air. In Fig. 2.7b, the characteristic curves of an ion-exchange resistance electric hygrometer show that there is a nonlinear relationship between resistance R and relative humidity I/>.These sensors cover a wider range than Dunmore sensors, from 20 to 90 percent relative humidity. Capacitance Hygrometer. The commonly used capacitance sensor consists of a thin-film plastic foil. A very thin gold coating covers both sides of the film as electrodes, and the film is mounted

2.18

CHAPTER TWO

FIGURE 2.7 Ion-exchange resistance-type electric hygrometer . (a) Front and side vie w of sensor; (b) characteristic curv e of R v ersus . (Reprinted by permission of General Eastern Instruments.)

inside a capsule. The golden electrodes and the di viding plastic foil form a capacitor . Water vapor penetrates the gold layer , which is affected by the v apor pressure of the ambient air and, therefore, the ambient relati ve humidity. The number of w ater molecules absorbed on the plastic foil determines the capacitance and the resistance between the electrodes.

PSYCHROMETRICS

2.19

Comparison of Various Methods The following table summarizes the sensor characteristics of v arious methods to be used within a temperature range of 32 to 120°F (0 to 50°C) and a range of 10 to 95 percent RH:

Psychrometer Mechanical Dunmore Ion-exchange Capacitance

Operating method

Accuracy, % RH

Wet-bulb depression Dimensional change Electric resistance Electric resistance Electric capacitance and resistance

3 3 to 5 1.5 2 to 5 3 to 5

Psychrometers are simple and comparatively low in cost. They suffer no irreversible damage at 100 percent RH, as do the sensors of electric hygrometers. Unfortunately, complete wet-bulb depression readings of psychrometers become diff cult when relative humidity drops below 20 percent or when the temperature is belo w the freezing point. F or remote monitoring, it is diff cult to keep suff cient water in the water reservoir. Therefore, psychrometers are sometimes used to check the temperature and relative humidity in the air conditioned space manually. Mechanical hygrometers directly indicate the relati ve humidity of the moist air . They are also simple and relati vely inexpensive. Their main dra wbacks are their lack of precision o ver an e xtensive period and their lack of accurac y at e xtreme high and lo w relati ve humidities. Electronic hygrometers, especially the polymer f lm resistance and the capacitance types, are commonly used for remote monitoring and for controls in man y air conditioning systems. Both the electronic and mechanical hygrometers need re gular calibration. Initial calibrations are usually performed either with precision humidity generators using tw o-pressure, two-temperature, and divided-f ow systems or with secondary standards during manuf acturing (refer to ASHRAE Standard 41.6-1982, Standard Method for Measur ement of Moist Air Properties). Re gular calibrations can be done with a precision aspiration psychrometer or with chilled mirror dew-point devices. Air contamination has signi f cant inf uence on the performance of the sensor of electronic and mechanical hygrometers. This is one of the reasons why they need regular calibration.

2.9 PSYCHROMETRIC CHARTS Psychrometric charts pro vide a graphical representation of the thermodynamic properties of moist air, various air conditioning processes, and air conditioning cycles. The charts are very helpful during the calculation, analysis, and solution of the complicated problems encountered in air conditioning processes and cycles. Basic Coordinates. ●

●

The currently used psychrometric charts have two types of coordinates:

h-w chart. Enthalpy h and humidity ratio w are basic coordinates. The psychrometric charts published by ASHRAE and the Chartered Institution of Building Services Engineering (CIBSE) are h-w charts. T-w chart. Temperature T and humidity ratio w are basic coordinates. Most of the psychrometric charts published by the large manufacturers in the United States are T-w charts.

For an atmospheric pressure of 29.92 in. Hg (760 mm Hg), an air temperature of 84°F (28.9°C), and a relative humidity of 100 percent, the humidity ratios and enthalpies found from the psychrometric charts published by ASHRAE and Carrier International Corporation are shown below:

2.20

CHAPTER TWO

ASHRAE’s chart Carrier’s chart

Humidity ratio, lb/lb

Enthalpy, Btu/lb

0.02560 0.02545

48.23 48.20

The last digit for humidity ratios and for enthalpies read from ASHRAE’s chart is an approximation. Nevertheless, the differences between the two charts are less than 1 percent, and these are considered negligible. In this handbook, for manual psychrometric calculations and analyses, ASHRAE’s chart will be used. Temperature Range and Barometric Pressure. ASHRAE’s psychrometric charts are constructed for various temperature ranges and altitudes. In Appendix B only the one for normal temperature, that is, 32 to 120 °F (0 to 50 °C), and a standard barometric pressure at sea le vel, 29.92 in. Hg (760 mm Hg), is shown. The skeleton of ASHRAE’s chart is shown in Fig. 2.8. Enthalpy Lines. For ASHRAE ’s chart, the molar enthalp y of moist air is calculated from the formulation recommended by Hyland and Wexler (1983) in their paper “Formulations for the Thermodynamic Properties of Dry Air from 173.15 K to 473.15 K, and of Saturated Moist Air from 173.15 K to 473.15 K, at Pressures to 5 MPa.” For ASHRAE’s chart, the enthalpy h lines incline at an angle of 25° to the horizontal lines. The scale factor for the enthalpy lines Ch, Btu/lbft (kJ /kgm), is Ch 

h2  h1 20   48.45 Btu /lbft (370 kJ / kg m) Lh 0.4128

where Lh  shortest distance between enthalpy lines h2 and h1, ft (m).

FIGURE 2.8 Skeleton of ASHRAE’s psychrometric chart.

(2.47)

PSYCHROMETRICS

2.21

Humidity Ratio Lines. In ASHRAE’s chart, the humidity ratio w lines are horizontal. They form the ordinate of the psychrometric chart. The scale f actor Cw, lb/lbft (kg /kgm), for w lines in ASHRAE’s chart is Cw 

w2  w1 0.020   0.040 lb / lbft (0.131 kg / kg m) Lw 0.5

(2.48)

where Lw  vertical distance between w2 and w1, ft (m). For ASHRAE’s chart, the humidity ratio w can be calculated by Eq. (2.15). Constant-Temperature Lines. For ASHRAE ’s chart, since enthalp y is one of the coordinates, only the 120 °F constant-temperature T line is a true v ertical. All the other constant-temperature lines incline slightly to the left at the top. From Eq. (2.27), T  ha /cpd; therefore, one end of the T line in ASHRAE’S chart can be determined from the enthalp y scale at w  0. The other end can be determined by locating the saturated humidity ratio ws on the saturation curve. Saturation Curve. A saturation curv e is a locus representing a series of state points of saturated moist air. For ASHRAE’s chart, the enthalpy of saturation v apor over liquid w ater or o ver ice at a certain temperature is calculated by the formula recommended in the Hyland and Wexler paper. The humidity ratios of the saturated moist air ws between 0 and 100 °F (17.8 and 37.8 °C) in the psychrometric chart can also be calculated by the following simpler polynomial: ws  a 1  a 2Ts  a 3Ts2  a 4Ts3  a 5Ts4

(2.49)

where Ts  saturated temperature of moist air, °F (°C) a1  0.00080264 a2  0.000024525 a3  2.5420 10  6 a4   2.5855 108 a5  4.038 1010 If we use Eq. (2.49) to calculate ws, the error is most probably less than 0.000043 lb /lb (kg /kg). It is far smaller than the value that can be identif ed on the psychrometric chart, and therefore the calculated ws is acceptable. Relative-Humidity and Moist Volume Lines. For ASHRAE ’s chart, relative-humidity  lines, thermodynamic wet-b ulb T* lines, and moist v olume v lines all are calculated and determined based on the formulations in Hyland and Wexler’s paper. Thermodynamic Wet-Bulb Lines. For ASHRAE’s chart, only thermodynamic wet-b ulb T* lines are shown. Since the dry-b ulb and the thermodynamic wet-b ulb temperatures coincide with each other on the saturation curve, one end of the T* line is determined. The other end of the T* line can be plotted on the w line where w  0. Let the state point of the other end of the T* line be represented by 1. Then, from Eq. (2.39), at w1  0 cpd (T1  T *)  w* s h* fg Solving for T *, we have T* 

cpdT1  w* s h* fg cpd

(2.50)

2.22

CHAPTER TWO

Cooling and Dehumidifying Curv es. The tw o cooling and dehumidifying curv es plotted on ASHRAE’s chart are based on data on coil performance published in the catalogs of U.S. manufacturers. These curves are very helpful in describing the actual locus of a cooling and dehumidifying process as well as determining the state points of air leaving the cooling coil.

2.10 DETERMINATION OF THERMODYNAMIC PROPERTIES ON PSYCHROMETRIC CHARTS There are seven thermodynamic properties or property groups of moist air sho wn on a psychrometric chart: 1. 2. 3. 4. 5. 6. 7.

Enthalpy h Relative humidity  Thermodynamic wet-bulb temperature T* Barometric or atmospheric pressure pat Temperature T and saturation water vapor pressure pws Density and moist volume v Humidity ratio w, water vapor pressure pw , and dew-point temperature Tdew

The f fth, sixth, and seventh are thermodynamic property groups. These properties or properties groups are independent of each other e xcept that the dif ference in slope between the enthalp y h line and thermodynamic wet-b ulb temperature T* line is small, and it is hard to determine their intersection. Usually, atmospheric pressure pat is a known value based on the altitude of the location. Then, in the f fth property group, pws is a function of temperature T only. In the sixth property group, according to Eq. (2.33), a  1/v; that is, air density and moist v olume are dependent on each other . In the seventh property group, for a given value of pat, properties w, pw , and Tdew are all dependent on each other. When pat is a known value, and if the moist air is not saturated, then any two known independent thermodynamic properties can determine the magnitudes of the remaining unkno wn properties. If the moist air is saturated, then any independent property will determine the remaining magnitudes. Example 2.1. The design indoor air temperature and relative humidity of an air conditioned space at sea level are 75°F (23.9°C) and 50 percent. Find the humidity ratio, the enthalpy, and the density of the indoor moist air 1. By using the ASHRAE chart 2. By calculation Determine also the de w-point and thermodynamic wet-b ulb temperatures of the moist air . The following information is required for the calculations: Atmospheric pressure at sea level Specif c heat of dry air Specif c heat of water vapor Enthalpy of saturated vapor at 0°F Gas constant of dry air

14.697 psi (101,325 Pa) 0.240 Btu/lb °F (1.005 kJ/kgK) 0.444 Btu/lb °F (1.859 kJ/kgK) 1061 Btu/lb (2468 kJ/kg) 53.352 ftlbf /lbm °R (0.287 kJ/kg K)

Solution 1. Plot the space point r on ASHRAE’s chart by f rst f nding the space temperature Tr  75°F on

PSYCHROMETRICS

2.23

FIGURE 2.9 Thermodynamic properties determined from ASHRAE’s psychrometric chart.

the abscissa and then follo wing along the 75 °F constant-temperature line up to a relati ve humidity   50 percent, as shown in Fig. 2.9. Draw a horizontal line from the space point r. This line meets the ordinate, humidity ratio w, at a value of wr  0.00927 lb /lb (kg /kg). This is the humidity ratio of the indoor space air . Draw a line parallel to the enthalp y line from the space point r. This line meets the enthalp y scale line at a value of hr  28.1 Btu /lb. This is the enthalp y of the indoor space air . Dra w a horizontal line from the space point r to the left. This line meets the saturation curv e at a de w-point temperature of 55°F (12.8 °C). Draw a line parallel to the thermodynamic wet-b ulb temperature lines through the space point r. The perpendicular scale to this line shows a thermodynamic wet-bulb temperature T* r  62.5°F (16.9 °C). Dra w a line parallel to the moist v olume lines through the space point r. The perpendicular scale to this line shows a moist volume vr  13.68 ft3/lb (0.853 m3 /kg). 2. The calculations of the humidity ratio, enthalpy, and moist volume are as follows: From Eq. (2.49), the humidity ratio of the saturated air at the dry-bulb temperature is ws  0.00080264  0.000024525T  2.542e-06T 2  2.5855e-08T 3  4.038e-10T 4  0.00080264  0.0018394  0.014299  0.010908  0.012776  0.018809 lb /lb According to Eq. (2.16), the saturated water vapor pressure of the indoor air is

2.24

CHAPTER TWO

pws 

ws pat 0.018809 14.697   0.4314 psia ws  0.62198 0.018809  0.62198

From Eq. (2.21), the water vapor pressure of indoor air is pw  pws  0.5 0.4314  0.2157 psia Then, from Eq. (2.15), the humidity ratio of the indoor air is wr 

0.62198pw 0.62198 0.2157   0.009264 lb / lb (0.009264 kg / kg) pat  pw 14.697  0.2157

From Eq. (2.29), the enthalpy of the indoor moist air is hr  cpd TF  w (hg  cps T) hr  0.240 75  0.009264(1061  0.444 75)  28.14 Btu/lb From Eq. (2.32), the moist volume of the indoor air is vr 

R aT 53.352 535  13.688 ft 3/ lb  pat  pw (14.697  0.2157) (144)

Then, from Eq. (2.33), the density of the indoor moist air is

r 

1 1   0.07306 lb / ft 3 (1.17 kg / m3) vr 13.688

Comparison of the thermodynamic properties read directly from ASHRAE’s chart and the calculated values is as follows:

Humidity ratio, lb/lb Enthalpy, Btu/lb Moist volume, ft3 / lb

ASHRAE’s chart

Calculated value

0.00927 28.1 13.68

0.009264 28.14 13.69

Apparently, the differences between the readings from ASHRAE’s chart and the calculated v alues are very small. Example 2.2. An HVAC&R operator measured the dry- and wet-b ulb temperatures in an air conditioned space as 75 °F (23.9 °C) and 63 °F (17.2 °C), respectively. Find the relati ve humidity of this air conditioned space by using ASHRAE’s chart and by calculation. The humidity ratios of the saturated air at temperatures of 75 and 63 °F are 0.018809 and 0.012355 lb /lb (kg /kg), respectively. Solution. It is assumed that the dif ference between the wet-b ulb temperature as measured by sling or aspiration psychrometer and the thermodynamic wet-b ulb temperature is ne gligible. At a measured dry-bulb temperature of 75 °F (23.9°C) and a wet-bulb temperature of 63 °F (17.2°C), the relative humidity read directly from ASHRAE’s chart is about 51.8 percent. From Sec. 2.6, the wetbulb constant K for a sling psychrometer is 0.000218 1/ °F. Then, from Eq. (2.43), the humidity ratio of the space air can be calculated as

PSYCHROMETRICS

2.25

w  ws  0.000218(T  T)  0.012355  0.000218(75  63)  0.009739 lb/lb From Eq. (2.15), the vapor pressure of the space air is wpat w  0.62198 0.009739 14.697   0.2266 psia 0.009739  0.62198

pw 

And from Eq. (2.16), the saturated vapor pressure at a space temperature of 75°F is pws  

ws pat ws  0.62198 0.018809 14.697  0.4314 psia 0.018809  0.62198

Hence, from Eq. (2.21), the calculated relative humidity of the space air is



pw 0.2266   0.5253 or 52.53% pws 0.4314

The dif ference between the v alue read directly from ASHRAE’s chart and the calculated one is 52.53 percent  51.8 percent  0.7 percent. Computer-Aided Psychrometric Calculation and Analysis There are tw o kinds of psychrometric computer -aided software available on the mark et: psychrometric calculations and psychrometric graphics. Most of the psychrometric softw are is Windowsbased computer programs. Software for psychrometric calculations can determine any one of the thermodynamic properties of the moist air if tw o of the independent properties are kno wn. Psychrometric calculation software usually also f nds the thermodynamic property of the mixture of airstreams and pro vides altitude effect adjustments. Software programs for psychrometric graphics are f ar more powerful tools than psychrometric calculations. On the computer screen, the following is shown: a psychrometic chart, any number of labeled air conditioned state points, the corresponding air conditioned processes and the air conditioning c ycle, and the spreadsheet that lists the thermodynamic properties, airf ow i n cubic feet per minute, and the heat transfer during the air conditioning processes. The thermodynamic properties of an y of the state points and therefore the characteristics of the air conditioning process and cycle can be varied. As a result, the operation of the air system at either full load or part load, under cooling or heating modes, can be investigated and analyzed.

REFERENCES ASHRAE, ASHRAE Handbook 1993, Fundamentals, Atlanta, GA, 1993. ASHRAE, ANSI/ASHRAE Standard 41.6-1982, Standard Method for Measurement of Moist Air Properties, Atlanta, GA, 1982. ASHRAE, ASHRAE Standard 41.1-1986, Standard Measurements Guide: Section on Temperature Measurements, Atlanta, GA, 1986.

2.26

CHAPTER TWO

Aslam, S., Charmchi, M., and Gaggioli, R. A., Psychrometric Analysis for Arbitrary Dry-Gas Mixtures and Pressures Using Microcomputers, ASHRAE Transactions, 1986, Part I B, pp. 448 – 460. Hedlin, C. P., Humidity Measurement with Dunmore Type Sensors, Symposium at ASHRAE Semiannual Meeting, ASHRAE Inc., New York, February 1968. Hyland, R. W., and Wexler, A., Formulations for the Thermodynamic Properties of Dry Air from 173.15 K to 473.15 K, and of Saturated Moist Air from 173.15 K to 372.15 K, at Pressures to 5 MPa, ASHRAE Transactions, 1983, Part II A, pp. 520 – 535. Hyland, R. W., and Wexler, A., Formulations for the Thermodynamic Properties of the Saturated Phases of H2O from 173.15 K to 473.15 K, ASHRAE Transactions, 1983, Part II A, pp. 500 – 519. Kamm, V., New Psychrometric Software Offer Free on the Internet, The Air Conditioning, Heating and Refrigeraion News , July 22, 1996, pp. 18 – 19. McGee, T. D., Principles and Methods of Temperature Measurements, 1st ed., Wiley, New York, 1988. Nelson, R. M., and Pate, M., A Comparison of Three Moist Air Property Formulations for Computer Applications, ASHRAE Transactions, 1986, Part I B, pp. 435 – 447. Stewart, R. B., Jacobsen, R. T., and Becker, J. H., Formulations for Thermodynamic Properties of Moist Air at Low Pressures as Used for Construction of New ASHRAE SI Unit Psychrometric Charts, ASHRAE Transactions, 1983, Part II A, pp. 536 – 548. Threlkeld, J. L., Thermal Environmental Engineering, 2d ed., Prentice-Hall, Englewood Cliffs, NJ, 1970. The Trane Company, Psychrometry, La Crosse, WI, 1979. Wang, S. K., Air Conditioning, vol. 1, 1st ed., Hong Kong Polytechnic, Hong Kong, 1987.

CHAPTER 3

HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE 3.1 BUILDING ENVELOPE 3.2 3.2 HEAT-TRANSFER FUNDAMENTALS 3.2 Conductive Heat Transfer 3.3 Convective Heat Transfer 3.4 Radiant Heat Transfer 3.5 Overall Heat Transfer 3.6 Heat Capacity 3.8 3.3 HEAT-TRANSFER COEFFICIENTS 3.8 Coefficients for Radiant Heat Transfer 3.8 Coefficients for Forced Convection 3.9 Coefficients for Natural Convection 3.10 Surface Heat-Transfer Coefficients 3.10 3.4 MOISTURE TRANSFER 3.11 Sorption Isotherm 3.11 Moisture-Solid Relationship 3.12 Moisture Migration in Building Materials 3.13 Moisture Transfer from the Surface of the Building Envelope 3.14 Convective Mass-Transfer Coefficients 3.15 Moisture Transfer in Building Envelopes 3.16 3.5 CONDENSATION IN BUILDINGS 3.17 Visible Surface Condensation 3.17 Concealed Condensation within the Building Envelope 3.18 3.6 THERMAL INSULATION 3.18 Basic Materials and Thermal Properties 3.19 Moisture Content of Insulation Material 3.19 Economic Thickness 3.21 Thermal Resistance of Airspaces 3.21 3.7 SOLAR ANGLES 3.22 Basic Solar Angles 3.22 Hour Angle and Apparent Solar Time 3.24 Solar Angle Relationships 3.24 Angle of Incidence and Solar Intensity 3.24

3.8 SOLAR RADIATION 3.25 Solar Radiation for a Clear Sky 3.26 Solar Radiation for a Cloudy Sky 3.28 3.9 FENESTRATION 3.29 Types of Window Glass (Glazing) 3.29 Optical Properties of Sunlit Glazing 3.30 3.10 HEAT ADMITTED THROUGH WINDOWS 3.32 Heat Gain for Single Glazing 3.32 Heat Gain for Double Glazing 3.34 Shading Coefficients 3.36 Solar Heat Gain Factors and Total Shortwave Irradiance 3.37 Selection of Glazing 3.39 3.11 SHADING OF GLASS 3.40 Indoor Shading Devices 3.40 External Shading Devices 3.42 Shading from Adjacent Buildings 3.43 3.12 HEAT EXCHANGE BETWEEN THE OUTER BUILDING SURFACE AND ITS SURROUNDINGS 3.46 Sol-Air Temperature 3.47 3.13 COMPLIANCE WITH ASHRAE/IESNA STANDARD 90.1-1999 FOR BUILDING ENVELOPE 3.48 General Requirements 3.48 Mandatory Provisions 3.49 Prescriptive Building Envelope Option 3.49 Building Envelope Trade-Off Option 3.50 3.14 ENERGY-EFFICIENT AND COSTEFFECTIVE MEASURES FOR BUILDING ENVELOPE 3.50 Exterior Walls 3.50 Windows 3.50 Infiltration 3.51 Energy-Efficient Measures for Commercial Buildings in the United States 3.51 REFERENCES 3.51

3.1

3.2

CHAPTER THREE

3.1 BUILDING ENVELOPE Building en velope consists of the b uilding components that enclose conditioned spaces. Heat, moisture, and contaminants may be transferred to or from the outdoors or unconditioned spaces and therefore affect the indoor environment of the conditioned space. Building envelope used for air conditioned space in b uildings consists of mainly w alls, roofs, windows, ceilings, and floors. There are tw o types of partitions: exterior partitions and demising partitions. An exterior partition is an opaque, translucent, or transparent solid barrier that separates conditioned space from outdoors or space which is not enclosed. A demising partition is a solid barrier that separates conditioned space from enclosed unconditioned space. ●

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●

●

An exterior wall is a solid e xterior partition which separates conditioned space from the outdoors. Exterior w alls are usually made from composite layers including an y of the follo wing: stucco, bricks, concrete, concrete blocks, wood, thermal insulation, vapor barrier, airspace, and interior fin ish. The gross exterior wall area is the sum of the windo w area, door area, and exterior wall area. A partition wall is an interior solid barrier which separates a conditioned space from others. Partition walls are usually made from interior surf ace finishes on t o sides, wooden studs and boards, concrete blocks, concrete, bricks, and thermal insulation. A demising partition w all separates a conditioned space from an enclosed unconditioned space. Wall below gr ade is a solid barrier belo w ground le vel which often separates a basement or a crawl space from soil. A roof is an exterior partition that has a slope less than 60° from horizontal and has a conditioned space below directly or through a ceiling indirectly. Roofs are usually made from clay tile, waterproof membrane, concrete and lightweight concrete, wood, and thermal insulation. A ceiling is an interior partition that separates the conditioned space from a ceiling plenum. The ceiling plenum may or may not be air conditioned. Ceilings are usually made from acoustic tile or boards, thermal insulation, and interior surface finishes An exterior floo is a horizontal e xterior partition under conditioned space. A floor placed ver a ventilated basement or a parking space is an e xterior floo . Exterior floors are usually made fro wood, concrete, thermal insulation, and face tiles. Slab on gr ade is a concrete floor slab on the ground. There is usually a v apor barrier, thermal insulation, and gravel and sand fill between the concrete slab and the ground. A window is glazing of an y transparent or translucent material plus sash, frame, mullions, and dividers in the b uilding envelope. Glazing is usually made from glass and transparent plastics. Frames are often made from w ood, aluminium, and steel. The windo w area is the area of the surface of glazing plus the area of frame, sash, and mullions. A fenestration is an y area on the exterior b uilding en velope which admits light indoors. Fenestrations include windo ws, glass doors, and skylights. A skylight is glazing ha ving a slope of less than 60° from the horizontal. There is often a conditioned space below skylight(s).

3.2 HEAT-TRANSFER FUNDAMENTALS Heat transfer between tw o bodies, two materials, or tw o re gions is the result of temperature difference. The science of heat transfer has pro vided calculations and analyses to predict rates of heat transfer . The design of an air conditioning system must include estimates of heat transfer between the conditioned space, its contents, and its surroundings, to determine cooling and heating loads. Heat-transfer analysis can be described in three modes: conduction, convection, and radiation.

HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE

3.3

Conductive Heat Transfer Conduction is the mechanism of heat transfer in opaque solid media, such as through w alls and roofs. For one-dimensional steady-state heat conduction qk, Btu/h (W), Fourier’s law gives the following relationship: qk  kA

dT dx

(3.1)

where k  thermal conductivity, Btu/h ft °F (W/m°C) A  cross-sectional area normal to heat fl w, ft2 (m2) T  temperature, °F (°C) x  coordinate dimension along heat fl w, ft (m) Equation (3.1) shows that the rate of heat transfer is directly proportional to the temperature gradient dT/dx, the thermal conductivity k, and the cross-sectional area A. The minus sign indicates that the heat must fl w in the direction of decreasing temperature; i.e., if the temperature decreases as x increases, the gradient dT/dx is negative, so that heat conduction is a positive quantity. For steady-state heat conduction through a plane composite w all with perfect thermal contact between each layer , as sho wn in Fig. 3.1, the rate of heat transfer through each section of the composite wall must be the same. From Fourier’s law of conduction, qk 

kC A kA A kB A (T1  T2)  (T2  T3)  (T3  T4) LA LB LC

where LA, LB, LC  thickness of layers A, B, and C, respectively, of composite wall, ft (m) T1, T2, T3, T4  temperature at surfaces 1, 2, 3, and 4, respectively, °F (°C)

FIGURE 3.1 Steady-state one-dimensional heat conduction through a composite wall.

(3.2)

3.4

CHAPTER THREE

kA, kB, kC  thermal conductivity of layers A, B, and C, respectively, of composite wall, Btu/h ft °F (W/m °C) Eliminating T2 and T3, we have qk 

T1  T4 L A /(k A A)  L B /(k B A)  L C /(k C A)

(3.3)

For a multilayer composite w all of n layers in perfect thermal contact, the rate of conduction heat transfer is given as qk 

T1  Tn1 L 1 /(k 1 A)  L 2 /(k 2 A)      L n /(k n A)

(3.4)

Subscript n indicates the nth layer of the composite wall. In Eq. (3.2), conduction heat transfer of any of the layers can be written as kA T T  L R* L R*  kA

qk 

(3.5)

where R*  thermal resistance, h°F/Btu (°C/W). In Eq. (3.5), an analogy can be seen between heat f ow and Ohm’s law for an electric circuit. Here the temperature difference T  T1  T2 indicates thermal potential, analogous to electric potential. Thermal resistance R* is analogous to electric resistance, and heat f ow qk is analogous to electric current. The total conducti ve thermal resistance of a composite w all of n layers RT, h°F/Btu (°C/W), can be calculated as R* T  R* 1  R* 2      R* n

(3.6)

where R*1 , R* 2 ,    R* n  thermal resistances of layers 1, 2,   , n layer of the composite w all, h °F/Btu ( °C/W). The thermal circuit of a composite w all of three layers is sho wn in the lo wer part of Fig. 3.1.

Convective Heat Transfer Convective heat transfer occurs when a f uid comes in contact with a surf ace at a different temperature, such as the heat transfer taking place between the airstream in a duct and the duct wall. Convective heat transfer can be di vided into tw o types: forced convection and natural or free convection. When a f uid is forced to mo ve along the surf ace by an outside moti ve force, heat is transferred by forced con vection. When the motion of the f uid is caused by the density dif ference of the tw o streams in the f uid as a product of contacting a surf ace at a dif ferent temperature, the result is called natural or free convection. The rate of con vective heat transfer qc, Btu/h (W), can be e xpressed in the form of Ne wton’s law of cooling as qc  hc A(Ts  T )

(3.7)

where hc  average convective heat-transfer coeff cient, Btu/h  ft2 °F (W/m2 °C) Ts  surface temperature, °F (°C) T  temperature of f uid away from surface, °F (°C) In Eq. (3.7), the con vective heat-transfer coef f cient hc is usually determined empirically . It is related to a dimensionless group of f uid properties, such as the correlation of f at-plate forced

HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE

3.5

convection, as shown in the following equation: Nu L 

hc L  C Re nL Pr m k

(3.8)

where C  constant and k  thermal conductivity, Btu/hft °F (W/m°C). In Eq. (3.8), NuL is the Nusselt number, which is based on a characteristic length L. Characteristic length can be the length of the plate, the diameter of the tube, or the distance between tw o plates, in feet (meters). The Reynolds number ReL is represented by the following dimensionless group: Re L 

vL



vL

(3.9)

where   density of f uid, lb/ft3 (kg/m3 ) v  velocity of f uid, ft/s (m/s)  absolute viscosity of the f uid, lb/fts (kg/ms)

 kinematic viscosity of f uid, ft2/s (m2/s) The Prandtl number Pr is represented by Pr 

3600 cp

(3.10)

k

where cp  specif c heat at constant pressure of the f uid, Btu/lb°F (J /kg°C). Fluid properties used to calculate dimensionless groups are usually related to the f uid temperature Tf . That is, Tf 

Ts  T 2

(3.11)

Convective heat transfer can also be considered analogous to an electric circuit. From Eq. (3.7), the convective thermal resistance R*, c h °F/Btu (°C/W), is given as R c* 

1 hc A

(3.12)

Radiant Heat Transfer In radiant heat transfer , heat is transported in the form of electromagnetic w aves traveling at the speed of light. The net rate of radiant transfer qr, Btu/h (W), between a gray body at absolute temperature TR1 and a black surrounding enclosure at absolute temperature TR2 (for e xample, the approximate radiation exchange between occupant and surroundings in a conditioned space) can be calculated as qr  A1 1(T 4R1  T 4R2)

(3.13)

where   Stefan-Boltzmann constant  0.1714 108 Btu/h ft2 °R4 (W/m2 K4 ) A1  area of gray body, ft2 (m2 )

1  emissivity of surface of gray body TR1, TR2  absolute temperature of surfaces 1 and 2, °R (K) TR1  TR2)/(TR1  TR2), and let the thermal

If we multiply the right-hand side of Eq. (3.13) by ( resistance R* r 

TR1  TR2

A1 1(T 4R1



T 4R2 )



1 h r A1

(3.14)

3.6

CHAPTER THREE

where hr  radiant heat-transfer coef f cient, Btu/h  ft2 °F (W/ m2 °C), we f nd that as a consequence, qr  h r A1(T1  T2)

(3.15)

where T1, T2  temperature of surfaces 1 and 2, °F (°C). If either of the two surfaces is a black surface or can be approximated as a black surf ace, as was previously assumed for the conditioned space surroundings, the net rate of radiant heat transfer between surfaces 1 and 2 can be evaluated as qr  A1F1– 2(T 4R1  T 4R2 )

(3.16)

where F1–2  shape factor for a diffuse emitting area A1 and a receiving area A2. The thermal resistance can be similarly calculated.

Overall Heat Transfer In actual practice, many calculations of heat-transfer rates are combinations of conduction, convection, and radiation. Consider the composite w all sho wn in Fig. 3.1; in addition to the conduction through the wall, convection and radiation occur at inside and outside surf aces 1 and 4 of the composite w all. At the inside surface of the composite w all, the rate of heat transfer qi, Btu/h (W), consists of convec-tive heat transfer between f uid, the air, and solid surface qc and the radiant heat transfer qr , as follows: qi  qc  qr  h c A1(Ti  T1)  h r A1(Ti  T1)  h i A1(Ti  T1)

(3.17)

where Ti  indoor temperature, °F (°C). From Eq. (3.17), the inside surface heat-transfer coeff cient hi at the liquid-to-solid interface, Btu/hft2 °F (W/m2 °C) is hi  hc  hr and the thermal resistance (°C/W), is

R*i of the inner surf ace due to con vection and radiation, h °F/Btu R i* 

1 h i A1

(3.18)

Similarly, at the outside surf ace of the composite w all, the rate of heat-transfer qo, Btu/h ft2 °F (W/m2 °C), is qo  h o A4(T4  To)

(3.19)

where ho  outside surface heat-transfer coeff cient at f uid-to-solid interface, Btu/h ft2 °F (W/m2 °C) A4  area of surface 4, ft2 (m2 ) T4  temperature of surface 4, °F (°C) To  outdoor temperature, °F (°C) The outer thermal resistance Ro*, h °F/Btu (°C/W), is R* o 

1 h o A4

For one-dimensional steady-state heat transfer , the o verall heat-transfer rate of the composite

HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE

3.7

wall q, Btu/h (W), can be calculated as q  qi  qk  qo  UA(Ti  To) 

Ti  To R* T

(3.20)

where U  overall heat-transfer coeff cient, often called the U value, Btu/h ft2 °F (W/m2 °C) A  surface area perpendicular to heat f ow, ft2 (m2 ) RT*  overall thermal resistance of composite wall, h°F/Btu (°C/W) and R*T  R* i  R* A  R* B  R* C  R* o 

1 UA

(3.21)

Also, the thermal resistances can be written as R* A 

LA kA A

R* B 

LB kB B

R* C 

LC kC C

Therefore, the overall heat transfer coeff cient U is given as 1 1/h i  L A / k A  L B / k B  L C / k C  1/h o

U 

(3.22)

For plane surf aces, area A  A1  A2  A3  A4. F or c ylindrical surf aces, because the inside and outside surf ace areas are dif ferent, and because UA  1/RT*, it must be clari f ed whether the area is based on the inside surface area Ai, outside surface area Ao, or any chosen surface area. For convenient HVAC&R heat-transfer calculations, the reciprocal of the o verall heat-transfer coeff cient, often called the overall R value R T, hft2 °F/Btu (m 2 °C/W), is used. So RT can be expressed as RT 

1  Ri  R1  R2      Rn  Ro U

(3.23a)

Ri, Ro  R values of inside and outside surfaces of composite wall, h ft2 °F/Btu (m2 °C/W) R1, R2,   , Rn  R values of components 1, 2, . . ., n, h ft2 °F/Btu (m2 °C/W)

where

and Ri 

1 hi

R1 

L1 k1

R2 

L2 k2

  

Rn 

Ln kn

Ro 

1 ho

(3.23b)

Sometimes, for convenience, the unit of R is often omitted; for e xample, R-10 means the R value equals 10 h ft2 °F/Btu (m2 °C/W). A building envelope assembly, or a building shell assembly, includes the exterior wall assembly (i.e., walls, windows, and doors), the roof and ceiling assembly , and the f oor assembly. The areaweighted a verage o verall heat-transfer coef f cient of an en velope assembly Uav, Btu / hft2 °F (W/m2 °C), can be calculated as Uav 

U1 A1  U2 A2      Un An Ao

(3.24)

where A1, A2 . . ., An  area of individual elements 1, 2, . . ., n of envelope assembly, ft2 (m2 ) Ao  gross area of envelope assembly, ft2 (m2) U1, U2, . . ., Un  overall heat-transfer coeff cient of individual paths 1, 2, . . ., n of the envelope assembly, such as paths through windows, paths through walls, and paths through roof, Btu/h ft2 °F (W/m2 °C)

3.8

CHAPTER THREE

Heat Capacity The heat capacity (HC) per square foot (meter) of an element or component of a b uilding envelope or other structure depends on its mass and speci f c heat. Heat capacity HC, Btu/ft2 °F (kJ /m2 °C), can be calculated as HC  Lc A

(3.25)

where m  mass of building material, lb (kg) c  specif c heat of building material, Btu/lb°F (kJ/kg °C) A  area of building material, ft2 (m2)   density of building material, lb/ft3 (kg/m3 ) L  thickness or height of building material, ft (m)

3.3 HEAT-TRANSFER COEFFICIENTS Determination of heat-transfer coef f cients to be used for load calculations or year estimates is complicated by the following types of variables: ●

●

●

●

●

-round ener gy

Building envelopes, exterior wall, roof, glass, partition wall, ceiling, or f oor Fluid f ow, turbulent f ow, or laminar f ow, forced or free convection Heat f ow, horizontal heat f ow in a v ertical surf ace, or an upw ard or do wnward heat f ow in a horizontal surface Space air diffusion, ceiling or sidewall inlet, or others Time of operation — summer, winter, or other seasons

Among the three modes of heat transfer , convection processes and their related coef f cients are the least understood, making analysis diff cult.

Coefficients for Radiant Heat Transfer For a radiant e xchange between the inner surf ace of an e xterior wall and the surrounding surf aces (such as the surf aces of partition w alls, ceilings, and f oors) in an air conditioned room, the sum of the shape factors F1– n can be considered as unity. If all surfaces are assumed to be black, then the radiative heat-transfer coeff cient hr, Btu/h ft2 °F (W/m2 °C), can be calculated as hr 

 (T 4Ris  T 4Rrad) TRis  TRrad

(3.26)

where TRis  absolute temperature of inner surface of exterior wall, roof, or external window glass, °R (K) TRrad  absolute mean radiant temperature of surrounding surfaces, °R (K) Often TRrad is approximately equal to the air temperature of the conditioned space TRr when both are expressed in degrees Rankine (kelvins). Radiant heat transfer coeff cients calculated according Eq. (3.26) for various surface temperatures and temperature differences between the inner surf ace of an y building envelope and the surrounded surfaces are presented in Table 3.1. From Table 3.1, hr depends on the absolute temperature of inner surface TRis and the temperature difference Tis  Trad.

HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE

3.9

TABLE 3.1 Radiative Heat Transfer Coeff cients hr Temperature difference Tis – Trad, °F

Inner surface temperature Tis, °F 60 (520°R)

70 (530°R)

75 (535°R)

80 (540°R)

85 (545°R)

0.968 0.966 0.963 0.958 0.945 0.919

0.990 0.990 0.989 0.984 0.971 0.945

Btu/h  ft2  °F 1 2 3 5 10 20

0.871 0.866 0.862 0.856 0.844 0.819

0.909 0.910 0.908 0.904 0.892 0.868

0.935 0.936 0.935 0.930 0.918 0.893

Note: Calculations made by assuming that Trad  Tr and the emissivity of inner surface i  1.

Coefficients for Forced Convection Before one can select the mean con vective heat-transfer coeff cient hc in order to calculate the rate of convective heat transfer or to determine the o verall heat-transfer coeff cient U value during cooling load calculations, the type of convection (forced or natural) must be clarif ed. Forced convection between the surf aces of the b uilding envelope and the conditioned space air occurs when the air handling system is operating or there is a wind o ver the outside surf ace. In an indoor space, free convection is al ways assumed when the air -handling system is shut of f, or there is no forced-air motion over the surface involved. The convective heat-transfer coef f cient hc also depends on the air v elocity v f owing over the surface, the conf guration of the surf ace, the type of space air dif fusion, and the properties of the f uid. According to Kays and Crawford (1980), a linear or nearly linear relationship between hc and v holds. Even though the air velocity v in the occupied zone may be only 30 fpm (0.15 m /s) or even lower when the air-handling system is operating, however, the mode of heat transfer is still considered as forced convection and can be expressed as h c  A  Bv n

(3.27)

where n  exponential index, usually between 0.8 and 1 v  bulky air velocity of f uid 0.5 to 1 ft (0.15 to 0.30 m) from surface, fpm (m/min) A, B  constants On the basis of test data from Wong (1990), Sato et al. (1972), and Spitler et al. (1991), as well as many widely used ener gy estimation computer programs, the convective heat-transfer coef f cient for forced convection hc, Btu/h ft2 °F (W/m2 °C), for indoor surfaces can be determined as h c  1.0  0.008v

(3.28)

For outside surf aces, the surface heat-transfer coef f cient ho  hc  hr, Btu /h ft °F (W/ m °C), can be calculated as 2

h o  1.8  0.004vwind where vwind  wind speed, fpm.

2

(3.29)

3.10

CHAPTER THREE

Coefficients for Natural Convection For natural convection, the empirical relationship between the dimensionless groups containing the convective heat-transfer coeff cient hc can be expressed as follows: Nu L  C(GrLPr)n  C Ra nL

(3.30)

In Eq. (3.30), GrL, called the Grashof number, is based on the characteristic length L. GrL 

 g 2(Ts  T ) (3.31)

2

where   coeff cient of volume expansion of f uid, 1/ °R (1/K) g  acceleration of gravity, ft/s2 (m/s2)   density of f uid, lb/ft3 (kg/m3) C  constant And Ra is called the Rayleigh number, and Ra  Gr Pr. Natural convective heat transfer is dif f cult to evaluate because of the comple xity of the recirculating convective stream of room air that is the result of the temperature distrib ution of the surfaces and the v ariation of temperature pro f le of the stream. Computer programs using numerical techniques have been de veloped recently. In actual practice, simplif ed calculations are often adopted. Altmayer et al. (1983) in their recent e xperiments and analyses found that “The ASHRAE free convection correlations provide a f air prediction of the heat f ux to the room air from cold and hot surfaces.” The errors are mainly caused by the v ariation of temperature of the convective airstreams after contact with cold or hot surfaces. In simplif ed calculations, this variation can only be included in the calculation of the mean space air temperature T. Many rooms ha ve a v ertical w all height of about 9 ft (2.7 m). If the temperature dif ference Ts  T  1°F (0.56°C), the natural convection f ow is often turb ulent. If Ts  T  1°F (0.56°C), as in many cooling load calculations between a partition wall and space air, the f ow is laminar. Based on data published in ASHRAE Handbook 1997, Fundamentals, the natural con vection coeff cients hc, Btu / h ft2 °F (W/m2 °C), are given as follows: Vertical plates: Large plates, turbulent f ow

hc  0.19(Tsa)0.33

Small plates, laminar f ow

h c  0.29

Tsa

0.25

L

(3.32) (3.33)

Horizontal plates, facing upward when air is heated or facing downward when air is cooled: Large plates, turbulent f ow

hc  0.22(Tsa)0.33

Small plates, laminar f ow

h c  0.27

Tsa

0.25

L

(3.34) (3.35)

Horizontal plates, facing upward when air is cooled or facing downward when air is heated: Small plates

h c  0.12

Tsa

0.25

L

(3.36)

Here, Tsa indicates the temperature difference between the surface and the air. Surface Heat-Transfer Coefficients The surf ace heat-transfer coef f cient h, sometimes called the surface conductance , Btu / h ft2 °F (W/m2 °C), is the combination of con vective and radiant heat-transfer coef f cients; that is,

HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE

3.11

TABLE 3.2 Surface Heat-Transfer Coeff cients h, Btu / h ft2  °F Surface emissivity 0.90

0.20

Indoor surface Direction of heat f ow

Description Forced convection 30 fpm 50 fpm 660 fpm (7.5 mph) 1320 fpm (15 mph) Free convection Horizontal surface Vertical surface Horizontal surface

Upward Downward

Indoor surface

Tsa  10°F Summer

Winter

Tsa  1°F

2.21 2.37

2.11 2.27

2.16 2.32

1.36 1.42

1.27 1.33

1.17 1.15 1.03

Outdoor surface

4.44 7.08

Tsa  10°F Summer

Winter

Tsa  1°F

1.46 1.62

1.43 1.59

1.44 1.60

0.62 0.68

0.60 0.66

0.44 0.30

Note: Assume space temperature Tr  74°F year-round and Tr  Trad; here Trad indicates the mean radiant temperature of the surroundings.

h  hc  hr. Table 3.2 lists the h values for v arious surf ace types at Tsa  10°F (5.6 °C) and Tsa  1°F (0.56°C) during summer and winter design conditions. The values are based upon the following: ●

●

●

Trad  Ta, where Ta is the air temperature, °F (°C). Tsa indicates temperature difference between surface and air, °F (°C). Emissivity of the surface  0.9 and  0.2.

3.4 MOISTURE TRANSFER Moisture is w ater in the v apor, liquid, and solid states. Building materials e xposed to e xcessive moisture may degrade or deteriorate as a result of physical changes, chemical changes, and biological processes. Moisture accumulated inside the insulating layer also increases the rate of heat transfer through the b uilding en velope. Moisture transfer between the b uilding en velope and the conditioned space air has a signi f cant inf uence on the cooling load calculations in areas with hot and humid climates.

Sorption Isotherm Moisture content X, which is dimensionless or else in percentage, is def ned as the ratio of the mass of moisture contained in a solid to the mass of the bone-dry solid. An absorption isotherm is a constant-temperature curve for a material in which moisture content is plotted against an increased ambient relative humidity during an equilibrium state; i.e., the rate of condensation of w ater vapor on the surface of the material is equal to the rate of e vaporation of water vapor from the material. A desorption isotherm is also a constant-temperature curve for a material. It is a plot of moisture content versus a decreased ambient relative humidity during equilibrium state. Many building materials sho w different absorption and desorption isotherms. The difference in moisture content at a speci f c relative humidity between the absorption and desorption isotherms is called hysteresis. Figure 3.2 shows absorption and desorption isotherms of a building material.

3.12

CHAPTER THREE

FIGURE 3.2 Sorption isotherms.

Temperature also has an inf uence on the moisture content of many building materials. At a constant relative humidity in ambient air, the moisture content of a b uilding material will be lo wer at a higher temperature. When a building material absorbs moisture, heat as heat of sorption is evolved. If liquid water is absorbed by the material, an amount of heat ql, Btu/lb (kJ /kg) of w ater absorbed, similar to the heat of solution, is released. This heat results from the attracti ve forces between the w ater molecules and the molecules of the b uilding material. If w ater vapor is absorbed, then the heat released qv, Btu/lb (kJ/kg), is given by qv  ql  h fg

(3.37)

where hfg  latent heat of condensation, Btu/lb (kJ /kg). Heat of sorption of liquid w ater ql varies with equilibrium moisture content for a gi ven material. The lower the X and the  of ambient air , the higher will be the v alue of ql. For pine, ql may vary from 450 Btu /lb (1047 kJ /kg) for nearly bone-dry wood to 20 Btu /lb (46.5 kJ /kg) at a moisture content of 20 percent. Man y building materials have very low ql values compared with hfg, such as a ql of 40 Btu/lb (93 kJ/kg) for sand. Moisture-Solid Relationship Many b uilding materials ha ve numerous interstices and microcapillaries of radius less than 4 106 in. (0.1 m), which may or may not be interconnected. These interstices and microcapillaries provide large surface areas to absorb w ater molecules. Moisture can be bound to the solid surf aces by retention in the capillaries and interstices, or by dissolution into the cellular structures of f brous materials. When the relative humidity  of ambient air is less than 20 percent, moisture is tightly bound to individual sites in the monomolecular layer (re gion A, in Fig. 3.2). Moisture mo ves by vapor diffusion. The binding ener gy is affected by the characteristics of the surf ace, the chemical structure of the material, and the properties of water.

HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE

3.13

When relati ve humidity  is 20 to 60 percent (re gion B, Fig. 3.2), moisture is bound to the surface of the material in multimolecular layers and is held in microcapillaries. Moisture be gins to migrate in liquid phase, and its total pressure is reduced by the presence of moisture in microcapillaries. The binding energy involved is mainly the latent heat of condensation. When   60 percent, i.e., in region C of Fig. 3.2, moisture is retained in lar ge capillaries. It is relatively free for removal and chemical reactions. The vapor pressure of the moisture is in f uenced only moderately because of the moisture absorbed in re gions A and B. Moisture moves mainly in a liquid phase. Because the density of the liquid is much greater than that of w ater vapor, the moisture content in b uilding materials is mostly liquid. Unbound moisture can be trapped in interstices ha ving a radius greater than 4 105 in. (1 m), without signi f cantly lowering the v apor pressure. Free moisture is the moisture in e xcess of the equilibrium moisture content in a solid. F or insulating materials in which interstices are interconnected, air penetrates through these open-cell structures, and, moisture can be accumulated in the form of condensate and be retained in the lar ge capillaries and pores. Moisture Migration in Building Materials Building envelopes are not constructed only with open-cell materials. The airstream and its associated w ater v apor cannot penetrate b uilding envelopes. Air leakage can only squeeze through the cracks and gaps around windo ws and joints. Ho wever, all b uilding materials are moisturepermeable; in other w ords, moisture can migrate across a b uilding envelope because of dif ferences in moisture content or other driving potentials. According to Wong and Wang (1990), many theories have been proposed by scientists to predict the migration of moisture in solids. The currently accepted model of moisture f ow in solids is based upon the following assumptions: ●

●

●

●

●

●

Moisture migrates in solids in both liquid and v apor states. Liquid f ow is induced by capillary f ow and concentration gradients; vapor diffusion is induced by vapor pressure gradients. During the transport process, the moisture content, the vapor pressure, and the temperature are always in equilibrium at any location within the building material. Heat transfer within the b uilding material is in the conduction mode. It is also af fected by latent heat from phase changes. Vapor pressure gradients can be determined from moisture contents by means of sorption isotherms. Fick’s law is applicable. Only one-dimensional f ow across the b uilding en velope is considered. Building materials are homogeneous.

If the temperature gradient is small, for simplicity, the mass f ux for m˙ w /A f ow, lb / h ft2 (kg/h m2), can be expressed as m˙w /A   w Dlv

one-dimensional

dX dx

(3.38)

where Dlv  mass diffusivity of liquid and vapor, ft2 /h (m2 /h) X  moisture content, lb/lb (kg/kg) dry solid w  density of water, lb/ft3 (kg/m3 ) x  coordinate dimension along moisture f ow, ft (m) A  area of building envelope perpendicular to moisture f ow, ft2 (m2 ) Mass dif fusivities of some b uilding materials as a function of moisture content are sho Fig. 3.3.

wn in

3.14

CHAPTER THREE

FIGURE 3.3 Mass diffusivity D lv of some b uilding materials. ( Abridged with permission from Bruin et al. Advances in Drying, Vol. 1, 1979.)

Moisture Transfer from the Surface of the Building Envelope At a certain time instant, moisture migrating from an y part in the b uilding envelope to its surf ace must be balanced by con vective moisture transfer from the surf ace of the b uilding envelope to the ambient air and the change of the moisture content as well as the corresponding mass concentration at the surf ace of the b uilding envelope. Such a con vective moisture transfer is often a part of the space latent cooling load. Analogous to the rate of con vective heat transfer [Eq. (3.7)], the rate of convective moisture transfer m˙ w, lb/h (kg/h), can be calculated as m˙ w  h m Aw (Cws  Cwr)

(3.39)

where hm  convective mass-transfer coeff cient, ft/h (m/h) Aw  contact area between moisture and ambient air, ft2 (m2 ) Cws, Cwr  mass concentration of moisture at surface of building envelope and of space air, lb/ft3 (kg/m3) It is more con venient to e xpress the mass concentration dif ference Cws  Cwr in terms of a humidity ratio difference ws  wr. Here ws and wr represent the humidity ratio at the surf ace of the building envelope and of the space air , respectively, lb/lb (kg / kg). In terms of mass concentration we can write Cws  aws

(3.40)

Cwr  awr

(3.41)

Then the rate of convective moisture transfer can be rewritten as m˙ w  ah m Aw(ws  wr)

(3.42)

In Eq. (3.42), the surface humidity ratio ws depends on the moisture content Xs, the temperature Ts, and therefore the vapor pressure pws in the interstices of the surf ace of the building envelope. From the known Xs, Ts, and the sorption isotherm, the corresponding relati ve humidity s at the surf ace can be determined. If the dif ference between relati ve humidity  and the de gree of saturation is

HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE

3.15

ignored, then ws  swss  swss

(3.43)

where s  degree of saturation at surf ace. In Eq. (3.43), wss represents the humidity ratio of the saturated air. It can be determined from Eq. (2.49) since Ts is a known value. In Eq. (3.42), the contact area Aw between the liquid w ater at the surf ace of the b uilding envelope and the space air is a function of moisture content Xs. A precise calculation of Aw is very complicated, but if the surf ace area of the b uilding material is As, then a rough estimate can be made from Aw  X s As

(3.44)

Convective Mass-Transfer Coefficients The Chilton-Colburn analogy relates the heat and mass transfer in these forms: jH  jD hc hm Pr 2/3  Sc 2/3 a cpa v h c a a cpa 

2/3



 hm

vw

(3.45)

2/3

D  aw

where v  air velocity remote from surface, ft/s (m/s) cpa  specif c heat of moist air, Btu/lb °F (J/kg°C) hm  convective mass-transfer coeff cient, ft/h (m/s)

 kinematic viscosity, ft2 /s (m2 /s)   thermal diffusivity of air, ft2 /s (m2 /s) Daw  mass diffusivity for water vapor diffusing through air, ft2 /s (m2 /s) Sc  Schmidt number Subscript a indicates dry air and w, water vapor. For air at 77 °F (25 °C), cpa  0.243 Btu /lb°F (1020 kJ /kg°C), va  1.74 104 ft2/s (1.62 105 m2 /s),   2.44 104 ft2 /s (2.27 105 m2 /s), Daw  2.83 ft 2 /s (0.263 m 2 /s), and a  0.0719 lb / ft3 (1.15 kg/m3 ), we can show that Pr 

a 



1.74 10 4  0.713 2.44 10 4

and that Sc 

a Daw



1.74 10 4  0.614 2.83 104

At a space temperature of 77°F (25°C), therefore, hm 



h c Pr 2/3

a cpa Sc 2/3 h c (0.713)2/3  63.2h c 0.0719 0.243 (0.614)2/3

3.16

CHAPTER THREE

Moisture Transfer in Building Envelopes Moisture transfer in b uilding en velopes of a typical residential b uilding can proceed along tw o paths, as shown in Fig. 3.4: 1. Moisture migrates inside the b uilding material mainly in the form of liquid if the relati ve humidity of the ambient air is more than 50 percent. It will be transported to the indoor or outdoor air by convective mass transfer. The driving potentials are the moisture content of the b uilding material, the vapor pressure gradient inside the building material, and the humidity ratio gradient between the surface and the ambient air. 2. Air leakage and the associated w ater v apor in f ltrates or e xf ltrates through the cracks and gaps around the windows, doors, f xtures, outlets, and between the joints. Air and moisture enter the cavities and the airspace in the b uilding envelope. If the sheathing is not airtight, air leakage and its water vapor penetrate the perforated insulating board and dischar ge to the atmosphere, as shown in Fig. 3.4. If the sheathing is a closed-cell, airtight insulating board, then the airstream may in f ltrate through gaps between the joints of the insulating board or through cracks between the windo w sill and the external wall and discharge to the atmosphere. The dri ving potential of the air leakage and the associated w ater v apor is the total pressure differential between indoor and outdoor air across the b uilding envelope due to wind ef fects, stack effect, mechanical ventilation, or a combination of these. Moisture is moving in the vapor form. In leak y buildings, the moisture transfer by means of air leakage is often f ar greater than the moisture migration through solids. F or better -sealed commercial b uildings, moisture migration through the building envelope may be important.

FIGURE 3.4 Moisture transfer in building envelope along two paths.

HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE

3.17

Installation of vapor retarders is an effective means to block or to reduce the moisture transfer in the building envelope. Vapor retarders are covered in greater detail later.

3.5 CONDENSATION IN BUILDINGS When moist air contacts a solid surf ace whose temperature is lower than the dew point of the moist air, condensation occurs on the surf ace in the form of liquid w ater, or sometimes frost. Condensation can damage the surf ace f nish, deteriorate the material and cause objectionable odors, stains, corrosion, and mold gro wth; reduce the quality of the b uilding envelope with dripping w ater; and fog windows. Two types of condensation predominate in buildings: 1. Visible surf ace condensation on the interior surf aces of e xternal windo w glass, below-grade walls, f oor slabs on grade, and cold surfaces of inside equipment and pipes 2. Concealed condensation within the building envelope

Visible Surface Condensation To avoid visible surf ace condensation, either the de w point of the indoor air must be reduced to a temperature below that of the surf ace, or the indoor surf ace temperature must be raised to a le vel higher than the de w point of the indoor air . The indoor surf ace temperature of the b uilding envelope, such as T1 of the composite e xterior wall in Fig. 3.1, or any cold surface where condensation may occur can be calculated from Eqs. (3.17) and (3.20) as T1  Ti 

U(Ti  To) hi

(3.46)

where q  rate of heat transfer through building envelope, Btu/h (W) hi  inside surface heat-transfer coeff cient, Btu/h ft2 °F (W/m2 °C) Increasing the thermal insulation is al ways an ef fective and economical w ay to pre vent condensation because it sa ves energy and raises the surf ace temperature of the b uilding envelope and other cold surfaces. The dew point temperature of the indoor air Tr  dew is a function of humidity ratio wr of the space air. At a speci f c space temperature Tr , the lower the relati ve humidity, the lower the Tr  dew. There are several ways to lower the dew point of the indoor space air: ●

●

●

●

●

By blocking and controlling inf ltration of hot, humid outside air By reducing indoor moisture generation By using a v apor retarder, such as polyethylene f lm or asphalt layer , to prevent or decrease the migration of moisture from the soil and the outside wetted siding and sheathing By introducing dry outdoor air through mechanical or natural v entilation when doing so is not in conf ict with room humidity criteria By using dehumidif ers to reduce the humidity ratio of the indoor air

During winter, any interior surf ace temperature of the w all, roof, or glass is al ways lower than the indoor space temperature. Better insulation and multiple glazing increase the interior surf ace temperature and, at the same time, reduce the heat loss, providing a more comfortable indoor en vironment than is possible by decreasing the relati ve humidity. To reduce e xcessive indoor humidity in man y industrial applications, a local e xhausting booth that encloses the moisture-generating source is usually the remedy of f rst choice.

3.18

CHAPTER THREE

During summer , the outer surf ace of chilled w ater pipes and refrigerant pipes (e ven the cold supply duct) is at a lo wer temperature than the indoor air temperature. Suf f cient pipe and duct insulation is needed to pre vent surf ace condensation. When an air -handling system is shut do wn, the heavy construction mass often remains at a comparati vely lower surface temperature as the indoor humidity rises. To avoid indoor surf ace condensation, the enclosure must be tight enough to prevent the inf ltration of hot, humid air from outdoors.

Concealed Condensation within the Building Envelope Usually, concealed condensation within building materials does not accumulate as a result of moisture migration. Excessive free moisture, usually in the form of liquid or frost, at any location inside the building envelope results in a higher moisture content than in surrounding areas. Therefore, it produces a moisture migration outward from that location rather than into it. During cold seasons, concealed condensation in b uilding en velopes is mainly caused by the warm indoor air , usually at a de w-point temperature of 32 °F (0 °C) and abo ve. It leaks outw ard through the cracks and openings and enters the ca vities, the gaps between component layers, or even the penetrable insulating material. It ultimately contacts a surf ace at a temperature lo wer than the dew point of the indoor air . Donnelly et al. (1976) disco vered that a residential stud w all panel with a poor v apor retarder accumulated about 3 lb (14.6 kg) of moisture per square foot (meter) of wall area during a period of 31 days in winter . This moisture accumulated in the mineral f ber insulating board adjacent to the cold side of the sheathing and at the interf ace between the insulating board and the sheathing. The rate of condensation m˙ con, lb/h (kg/h), can be calculated as m˙ con  60V˙lk r (wr  ws con )

(3.47a)

where V˙lk  volume f ow rate of air leakage, cfm (m3 /min) r  density of indoor air, lb/ft3 (kg/m3 ) ws con  saturated humidity ratio corresponding to temperature of surface upon which condensation occurs, lb/lb (kg/kg) Vapor retarders are effective for reducing moisture transfer through the b uilding envelope. They not only decrease the moisture migration in the b uilding material signi f cantly, but also block air leakage effectively if their joints are properly sealed. Vapor retarders are classif ed as rigid, f exible, and coating types. Rigid type includes reinforced plastic, aluminum, and other metal sheets. Fle xible type includes metal foils, coated f lms, and plastic f lms. Coatings are mastics, paints, or fusible sheets, composed of asphalt, resin, or polymeric layers. The vapor retarder is generally located on the warm side of the insulation layer during winter heating. In an area where summer cooling is dominant, a vapor retarder to pre vent condensation in the external brick wall and insulation layer would often be located on the outside of the insulation. The absorptive brick w all tak es substantial amounts of w ater during rainf all. The subsequent sunn y period drives the w ater vapor farther inside the w all and into the insulation layer where the w ater vapor condenses.

3.6 THERMAL INSULATION Thermal insulation materials, usually in the form of boards, slabs, blocks, f lms, or blankets, retard the rate of heat transfer in conducti ve, convective, and radiant transfer modes. They are used within building envelopes or applied o ver the surf aces of equipment, piping, or ductwork to achie ve the following benef ts: ●

Savings of energy by reducing heat loss and heat gain from the surroundings

HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE ●

●

●

3.19

Prevention of surface condensation by increasing the surf ace temperature above the dew point of the ambient air Reduction of temperature dif ference between the inside surf ace and the space air for the thermal comfort of the occupants, when radiant heating or cooling is not desired Protection of the occupant from injury due to contact with hot pipes and equipment

Basic Materials and Thermal Properties Basic materials in the manuf acture of thermal insulation for b uilding envelopes or air conditioning systems include ●

●

●

Fibrous materials such as glass f ber, mineral wool, wood, cane, or other vegetable f bers Cellular materials such as cellular glass, foam rubber, polystyrene, and polyurethane Metallic ref ective membranes

Most insulating materials consist of numerous airspaces, either closed cells (i.e., airtight cellular materials) or open cells (i.e., f brous materials, penetrable by air). Thermal conductivity k, an important physical property of insulating material, indicates the rate of heat transfer by means of a combination of gas and solid conduction, radiation, and con vection through an insulating material, expressed in Btu in./hft2 °F or Btu /hft°F (W/ m°C). The thermal conductivity of an insulating material depends on its physical structure (such as cell size and shape or diameter of the f brous materials), density, temperature, and type and amount of binders and additi ves. Most of the currently used thermal insulating materials ha ve thermal conducti vities within a range of 0.15 to 0.4 Btu in./h ft2 °F (0.021 to 0.058 W/m °C). Thermal properties of some building and insulating materials, based on data published in the ASHRAE Handbook 1993, Fundamentals, are listed in Table 3.3. The thermal conducti vity of man y cellular insulating materials made from polymers lik e polyurethane and e xtruded polystyrene is not signi f cantly affected by change in density . Other insulating materials ha ve a density at which the thermal conducti vity is minimum. De viating from this , conductivity k increases always, whether  is greater or smaller in magnitude. The thermal conductivity of an insulating material apparently increases as its mean temperature rises. F or polystyrene, k increases from 0.14 to 0.32 Btu in./h  ft2 °F (0.020 to 0.046 W/m°C) when its mean temperature is raised from 300 to 570°F (150 to 300°C). Moisture Content of Insulation Material From the point of view of moisture transfer, penetrability is an important characteristic. Closed-cell airtight board or block normally cannot ha ve concealed condensation except at the gap between the joints and at the interf ace of two layers. Only when air and its associated w ater vapor penetrate an open-cell insulating material can concealed condensation form if the y contact the surf aces of the interstices and pores at a temperature lo wer than the de w point of the penetrating air . Concealed condensation might also accumulate in open-cell insulating materials. If there is e xcess free moisture in an insulating material, the thermal insulation may be de graded. The increase of moisture in the thermal insulation layer is often due to the absorption from the ambient air , space air, or moisture transfer from adjacent layer, ground, or wetted surface. Smolenski (1996) brought forward the question of how much is too much. In roo f ng insulation, many consider the answer to be the moisture content to produce a more than 20 percent loss in thermal eff ciency or a thermal resistance ratio (TRR) of less than 80 percent. The thermal resistance ratio is def ned as the ratio of wet to dry thermal resistance of the thermal insulation, in percent, or TRR 

(wet thermal resistance) 100 dry thermal resistance

(3.47b)

3.20

CHAPTER THREE

TABLE 3.3 Thermal Properties of Selected Materials

Aluminum (alloy 1100) Asbestos: insulation Asphalt Brick, building Brass (65% Cu, 35% Zn) Concrete (stone) Copper (electrolytic) Glass: crown (soda-lime) Glass wool Gypsum Ice (32°F) Iron: cast Mineral f berboard: acoustic tile, wet-molded wet-felted Paper Polystyrene, expanded, molded beads Polyurethane, cellular Plaster, cement and sand Platinum Rubber: vulcanized, soft hard Sand Steel (mild) Tin Wood: f r, white oak, white plywood, Douglas f r Wool: fabric

Density, lb/ft3

Thermal conductivity, (Btu/h ft  °F)

Specif c heat, (Btu/lb°F)

171 120 132 123 519 144 556 154 3.25 78 57.5 450

128 0.092 0.43 0.4 69 0.54 227 0.59 0.022 0.25 1.3 27.6

0.214 0.20 0.22 0.2 0.09 0.156 0.092 0.18 0.157 0.259 0.487 0.12

23 21

0.035 0.031

0.14 0.19

58 1.25 1.5 132 1340 68.6 74.3 94.6 489 455

0.075 0.021 0.013 0.43 39.9 0.08 0.092 0.19 26.2 37.5

0.32 0.29 0.38

27 47 34 20.6

0.068 0.102 0.07 0.037

0.33 0.57 0.29

0.032 0.48 0.191 0.12 0.056

Emissivity 0.09 0.93 0.93 0.033 Highly polished 0.072 Shiny 0.94 Smooth 0.903 Smooth plate 0.95 0.435 Freshly turned

0.92 0.91 Rough 0.054 Polished 0.86 Rough 0.95 Glossy 0.06

Bright

0.90

Planed

Source: Adapted with permission from ASHRAE Handbook 1989, Fundamentals.

The equilibrium moisture content of most commomly used insulation material at 90 percent relative humidity ambient air , as well as the moisture content of insulation material at 80 percent TRR by weight (percent of dry weight) and by v olume (percent of v olume of insulating material) are listed below:

Insulation material

Density, lb/ft3

Cellular glass Expanded polystyrene Glass f ber Urethane Phenolic foam

8.4 1.0 9.2 2.1 2.6

Moisture content, %

  90% 0.2 2.0 1.1 6.0 262 23.4

80% TRR, by weight

80% TRR, by volume

23 383 42

3.1 6.1 6.2 8.8 1.0

25

For instance, for cellular glass when in contact with an ambient air of 90 percent relative humidity at normal room temperaure that reaches an equilibrium state during moisture absorption, it has

HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE

3.21

only a moisture content of 0.2 percent by dry weight, which is far lower than the moisture content of 23 percent by weight at 80 percent TRR. Most above-listed insulating materials have a far lower equilibrium moisture content of dry weight at an ambient air of 90 percent relati ve humidity than moisture content at 80 percent TRR. Also, a closed-cell structure cellular glass requires months or years to reach an equilibrium moisture content by v olume corrresponding to 80 percent TRR whereas mineral wool and calcium silica absorb moisture in only hours. Economic Thickness The economic thickness of insulation is the thickness with the minimum o wning and operating costs. Owning cost is the net investment cost of the installed insulation Cin, in dollars, less any capital investment that can be made as a result of lo wer heat loss or gain Cpt, in dollars. Theoretically, for a new plant, some small savings might be made because of the reduction of the size of a central plant; but in actual practice, this is seldom considered. F or an e xisting plant where add-on insulation is being considered, Cpt is zero because the plant investment has already been made. Operating cost Cen includes the annualized cost of ener gy o ver the life of a ne w plant, or the remaining life of an e xisting plant, in n years. It can also be tak en as the number of years o ver which an owner wishes to have a total return of the net investment, considering both the interest and the fuel escalation rate. Total cost of insulation Ct, in dollars, for any given thickness is Ct  Cin  Cpt  Cen

(3.48)

When the thickness of the insulating material increases, the quantity Cin Cpt also increases, as shown in Fig. 3.5, and Cen decreases. As a result, Ct f rst decreases, drops to a minimum, and then increases. The optimum economic thickness occurs when Ct drops to a minimum. The closest commercially available thickness is the optimum thickness. Thermal Resistance of Airspaces Thermal resistance of an enclosed airspace Ra has a signif cant effect on the total thermal resistance RT of the building envelope, especially when the v alue of RT is low. Thermal resistance Ra depends on the characteristic of the surface (ref ective or nonref ective), the mean temperature, the temperature difference of the surfaces perpendicular to heat f ow, the width across the airspace along the heat f ow,

FIGURE 3.5 Optimum thickness of insulation material.

3.22

CHAPTER THREE

TABLE 3.4 R Values of Enclosed Airspace Ra, h  ft2  °F/Btu Mean temperature, °F

Temperature difference, °F

Type of surface

Summer, 90

10

Horizontal

Upward

Vertical

Horizontal

Horizontal

Downward

Horizontal

Upward

Vertical

Horizontal

Horizontal

Downward

Winter, 50

10

Direction of heat f ow

Emissivity E †

Width of airspace, in.

0.05

0.2

0.82

0.5 3.5 0.5 3.5 0.5 3.5

2.03 2.66 2.34 3.40 2.34 8.19

1.51 1.83 1.67 2.15 1.67 3.41

0.73 0.80 0.77 0.85 0.77 1.00

0.5 3.5 0.5 3.5 0.5 3.5

2.05 2.66 2.54 3.40 2.55 9.27

1.60 1.95 1.88 2.32 1.89 4.09

0.84 0.93 0.91 1.01 0.92 1.24

† Emissity E  1/(1/e1  1/e2  1), where e1, e2 indicate the emittances on two sides of the airspace. Source: Abridged with permission from ASHRAE Handbook 1989, Fundamentals.

and the direction of air f ow. The R values of the enclosed airspaces Ra, hft2 °F/Btu (m 2 °C/W), abridged from data published in ASHRAE Handbook 1989, Fundamentals, are presented in Table 3.4.

3.7 SOLAR ANGLES Basic Solar Angles The basic solar angles between the sun ’s rays and a speci f c surface under consideration are sho wn in Figs. 3.6 and 3.7.

FIGURE 3.6 Basic solar angles and position of sun’s rays at summer solstice.

FIGURE 3.7 Solar intensity and angle of incidence.

3.23

3.24

CHAPTER THREE

●

●

●

●

●

●

Solar altitude angle  (Fig. 3.7a and b) is the angle ROQ on a vertical plane between the sun’s ray OR and its projection on a horizontal plane on the surface of the earth. Solar azimuth  (Fig. 3.7a) is the angle SOQ on a horizontal plane between the due-south direction line OS and the horizontal projection of the sun’s ray OQ. Solar declination angle  (Fig. 3.6) is the angle between the earth-sun line and the equatorial plane. Solar declination  changes with the times of the year. It is shown in Fig. 3.6 on June 21. Surface-solar azimuth  (Fig. 3.7 a and c) is the angle POQ on a horizontal plane between the normal to a vertical surface OP and the horizontal projection of the sun’s rays OQ. Surface azimuth  (Fig. 3.7a) is the angle POS on a horizontal plane between OP and the direction line SN. Latitude angle L (Fig. 3.6) is the angle SOO on the longitudinal plane between the equatorial plane and the line OO that connects the point of incidence of the sun’s ray on the surface of earth O and the center of the earth O.

Hour Angle and Apparent Solar Time Hour angle H (Fig. 3.6) is the angle SOQ on a horizontal plane between the line OS indicating the noon of local solar time tls and the horizontal projection of the sun’s ray OQ. The values of the hour angle H before noon are taken to be positive. At 12 noon, H is equal to 0. After 12 noon, H is negative. Hour angle H, in degrees, can be calculated as H  0.25 (time in minutes from local solar noon)

(3.49)

The relationship between apparent solar time tas, as determined by a sundial and expressed in apparent solar time, and local standard time tst, both in minutes, is as follows: t as  t eq  t st  4(M  G)

(3.50)

where M  local standard time meridian and G  local longitude, both in degrees. In Eq. (3.50), teq, in minutes, indicates the difference in time between the mean time indicated by a clock running at a uniform rate and the solar time due to the variation of the earth’s orbital velocity throughout the year.

Solar Angle Relationships The relationship among the solar angles is given as sin   sin  sin L  cos  cos H cos L

(3.51)

and cos  

sin  sin L  sin  cos  cos L

(3.52)

Angle of Incidence and Solar Intensity The angle of incidence  (Fig. 3.7a) is the angle between the sun ’s rays radiating on a surf ace and the line normal to this surface. For a horizontal surface, the angle of incidence H is ROV; for a vertical surface, the angle of incidence V is ROP; and for a tilted surf ace, the angle of incidence  is ROU. Here,  is the angle between the tilted surface ABCD and the horizontal surface. Let IDN be the intensity of direct normal radiation, or solar intensity, on a surface normal to the sun’s ray, in Btu /h  ft2 (W/m2). In Fig. 3.7 b, IDN is resolved into the v ertical component RQ  IDN

HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE

3.25

sin  and the horizontal component OQ  IDN cos . In Fig. 3.7c, for the right triangle OPQ, angle POQ  , and hence OP  IDN cos  cos . In Fig. 3.7 d, OPL is a right triangle. From point P, a line PT can be dra wn perpendicular to the line normal OU, and hence, two right triangles are formed: OTP and LTP. In OTP, because angle OPT  , the horizontal component of IDN along the line normal of the tilting surf ace OU is OT  IDN cos  cos  sin  . In Fig. 3.7 e, draw line OU parallel to line OU. Again, line QM can be dra wn from point Q perpendicular to OU. Then the right triangles PTL and QMR are similar. Angles PLT and QRM are both equal to . In the right triangle QMR, the component of RQ that is parallel to the line normal to the tilting surf ace is MR  IDN sin  cos  . The intensity of solar rays normal to a tilted surf ace I, Btu/h ft2 (W/m2) is the v ector sum of the components of the line normal to the tilted surface, or I  IDN cos    IDN (cos  cos  sin   sin  cos 

(3.53)

3.8 SOLAR RADIATION Solar radiation pro vides most of the ener gy required for the earth ’s occupants, either directly or indirectly. It is the source of indoor daylight and helps to maintain a suitable indoor temperature during the cold seasons. At the same time, its inf uence on the indoor environment must be reduced and controlled during hot weather . The sun is located at a mean distance of 92,900,000 mi, (149,500,000 km) from the earth, and it has a surf ace temperature of about 10,800 °F (6000°C). It emits electromagnetic w aves at w avelengths of 0.29 to 3.5 m (micrometers). Visible light has wavelengths of 0.4 to 0.7 m and is responsible for 38 percent of the total energy striking the earth. The infrared re gion contains 53 percent. At the outer edge of the atmosphere at a mean earth-sun distance, the solar intensity , called the solar constant Isc, is 434.6 Btu /hft2 (1371 W/ m2 ). The extraterrestrial intensity Io, Btu/h ft2 (W/m2) varies as the earth-sun distance changes during the earth’s orbit. Based on the data from Miller et al. (1983), the breakdo wn of solar radiation reaching the earth’s surface and absorbed by the earth is listed in Table 3.5. As listed in Table 3.5, only 50 percent of the solar radiation that reaches the outer edge of the earth ’s atmosphere is absorbed by the clouds and the earth ’s surface. At any specif c location, the absorption, ref ection, and scattering of solar radiation depend on the composition of the atmosphere and the path length of the sun ’s rays through the atmosphere, expressed in terms of the air mass m. When the sun is directly o verhead, m  1. TABLE 3.5 Components of Solar Radiation That Traverse the Earth’s Atmosphere Components

Breakdowns

Scattered by air

11%

Absorbed by water vapor, dust, etc. Intercepted by clouds

16 45

Traversed through air

28

Ref ected to space Scattered to earth Ref ected to space Absorbed by clouds Diffused through clouds and absorbed by earth Absorbed by earth Ref ected by earth

6% 5 20 4 21 24 4

3.26

CHAPTER THREE

In Table 3.5, the part of solar radiation that gets through the atmosphere and reaches the earth ’s surface, in a direction that varies with solar angles over time, is called direct radiation. The part that is diffused by air molecules and dust, arriving at the earth’s surface in all directions, is called diffuse radiation. The magnitude of solar radiation depends on the composition of the atmosphere, especially on the cloudiness of the sk y. Therefore, different models are used to calculate the solar radiation reaching the surface of building envelopes. Solar Radiation for a Clear Sky ASHRAE recommends use of the follo wing relationships to calculate the solar radiation for a clear sky. The solar intensity of direct normal radiation IDN, Btu/h ft2 (W/m2 ), can be calculated as IDN 

ACn

(3.54) exp (B/sin ) where A  apparent solar radiation when air mass m  0 (its magnitudes are listed in Table 3.6), Btu/h ft (W/m2 ) B  atmospheric extinction coeff cient (Table 3.6), which depends mainly on the amount of water vapor contained in the atmosphere In Eq. (3.54), the term 1/sin  in exponential form denotes the length of the direct radiation path through the atmosphere. The term Cn is the clearness number of the sk y; Cn takes into account the dryness of the atmosphere and the dust contained in the air at a geographic location. Estimated Cn values for nonindustrial locations in the United States are shown in Fig. 3.8. From Eq. (3.53), the direct radiation radiated onto a horizontal surf ace through a clear sk y IDH, Btu/h ft2 (W/m2), is the vertical component of IDN, that is, IDH  IDN cos H  IDN sin 

(3.55)

and the direct radiation irradiated onto a v ertical surface for clear sky IDV, Btu/h ft (W/m ), is the horizontal component of IDN, or 2

2

IDV  IDN cos V  IDN cos  cos 

(3.56)

TABLE 3.6 Extraterrestrial Solar Radiation and Related Data for 21st Day of Each Month, Base Year 1964 A

January Febuary March April May June July August September October November December

I0, Btu / h ft2

Equation of time, min

Declination, deg

Btu/h  ft2

448.8 444.2 437.7 429.9 423.6 420.2 420.3 424.1 430.7 437.3 445.3 449.1

 11.2  13.9  7.5 1.1 3.3  1.4  6.2  2.4 7.5  15.4  13.8 1.6

 20.0  10.8 0.0  11.6  20.0  23.45  20.6  12.3 0.0  10.5  19.8  23.45

390 385 376 360 350 345 344 351 365 378 387 391

B

C

(Dimensionless ratios) 0.142 0.144 0.156 0.180 0.196 0.205 0.207 0.201 0.177 0.160 0.149 0.142

A: apparent solar radiation; B: atmospheric extinction coeff cient; C: diffuse radiation factor. Source: ASHRAE Handbook 1997, Fundamentals. Reprinted with permission.

0.058 0.060 0.071 0.097 0.121 0.134 0.136 0.122 0.092 0.073 0.063 0.057

HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE

3.27

FIGURE 3.8 Estimated atmospheric clearness numbers in the United States for nonindustrial localities.

From Eq. (3.53), for the direct radiation radiated onto a tilted surf ace at an angle  with the horizontal plane through clear sky ID, Btu/h ft2 (W/m2 ), can be evaluated as ID  I  IDN cos    IDN (cos  cos  sin   sin  cos )

(3.57)

Most of the ultra violet solar radiation is absorbed by the ozone layer in the upper atmosphere. Direct solar radiation, through an air mass m  2, arrives on the earth ’s surface at sea le vel during a clear day with a spectrum of  3 percent in the ultraviolet, 44 percent in the visible, and 53 percent in the infrared. The diffuse radiation Id, Btu/h ft2 (W/m2 ), is proportional to IDN on cloudless days and can be approximately calculated as Id 

CIDN Fss C 2n

(3.58)

where C  diffuse radiation f actor, as listed in Table 3.6, and Cn  clearness number of sk y from Fig. 3.8. In Eq. (3.58), Fss indicates the shape factor between the surface and the sky, or the fraction of shortwave radiation transmitted through the sk y that reaches the surf ace. For a v ertical surface Fss  0.5, for a horizontal surface Fss  1, and for any tilted surface with an angle  Fss 

1.0  cos  2

The total or global radiation on a horizontal plane

(3.59)

IG, Btu/h ft2 (W/m2), recorded by the U.S.

3.28

CHAPTER THREE

National Climatic Data Center (NCDC), can be calculated as



IG  ID  Id  IDN sin  

C C 2n



(3.60)

The ref ection of solar radiation from any surface Iref, Btu/h ft2 (W/m2 ), is given as Iref  sFsr (ID  Id)

(3.61)

where s  ref ectance of the surf ace and Fsr  shape factor between the recei ving surface and the ref ecting surf ace. The ground-re f ected dif fuse radiation f alling on an y surf ace Isg, Btu /hft2 (W/m2), can be expressed as Isg  g Fsg IG

(3.62)

where Fsg  shape factor between the surf ace and the ground and g  ref ectance of the ground. For concrete, g  0.23, and for bitumen and gra vel, g  0.14. A mean re f ectance g  0.2 is usually used for ground. The total intensity of solar radiation It, Btu /h  ft2 (W/m2 ), falling on a surf ace at a direction normal to the surface on clear days, is given by It  ID  Id  Iref. S

(3.63)

 (IDN  Iref.DN) cos   Id

where I  component of ref ected solar intensity in direction normal to surface, Btu/h ft2 (W/m2 ) Iref. DN  component of ref ected solar intensity in direction of sun ray, Btu/h  ft2 (W/m2 )

Solar Radiation for a Cloudy Sky For cloudy skies, the global horizontal irradiation IG*, Btu/h ft2 (W/m2 ), usually can be obtained from the NCDC. If it is not available, then it can be predicted from the following relationship:



I* G  1 

Ccc Q C 2ccR  IG P P



(3.64a)

Here Ccc indicates the cloud cover, on a scale of 0 to 10, and can be calculated by Ccc  CT  0.5

4

 Ccir. j j1

(3.64b)

where CT  total cloud amount Ccir. j  clouds covered by cirriforms, including cirrostratus, cirrocumulus, and cirrus, in j  1 to 4 layers Both CT and Ccir. j values can be obtained from the major weather stations. The values of coeff cients P, Q, and R, according to Kimura and Stephenson (1969), are listed below:

Spring Summer Autumn Winter

P

Q

R

1.06 0.96 0.95 1.14

0.012 0.033 0.030 0.003

 0.0084  0.0106  0.0108  0.0082

HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE

3.29

The direct radiation for a cloudy sky ID*, Btu/h ft2 (W/m2), can be calculated as I*D 

I* G sin  (1  Ccc /10) sin   C/C 2n

(3.65)

The diffuse radiation for a cloudy sky Id*, Btu/h ft2 (W/m2 ), is calculated as Id*  IG*  ID*

(3.66)

3.9 FENESTRATION Fenestration is the term used for assemblies containing glass or light-transmitting plastic, including appurtenances such as framing, mullions, dividers, and internal, external, and between-glass shading devices, as shown in Fig. 3.12 a. The purposes of fenestrations are to (1) pro vide a view of the outside world, (2) permit entry of daylight, (3) admit solar heat as a heating supplement in winter , (4) act as an emergency exit for single-story buildings, and (5) add to aesthetics. Solar radiation admitted through a glass or windo w pane can be an important heat gain for commercial buildings, with greater energy impact in the sun belt. HV AC&R designers are ask ed to control this solar load while pro viding the required visibility , daylight, and winter heating bene f ts as well as f re protection and safety features.

Types of Window Glass (Glazing) Most window glasses, or glazing, are vitreous silicate consisting of silicon dioxide, sodium oxide, calcium oxide, and sodium carbonate. They can be classif ed as follows: ●

●

●

●

●

Clear plate or sheet glass or plastic . Clear plate glass permits good visibility and transmits more solar radiation than other types. Tinted heat-absorbing glass. Tinted heat-absorbing glass is fabricated by adding small amounts of selenium, nickel, iron, or tin oxides. These produce colors from pink to green, including gray or bluish green, all of which absorb infrared solar heat and release a portion of this to the outside atmosphere through outer surf ace con vection and radiation. Heat-absorbing glass also reduces visible light transmission. Insulating glass . Insulating glass consists of tw o panes — an outer plate and a inner plate — or three panes separated by metal, foam, or rubber spacers around the edges and hermetically sealed in a stainless-steel or aluminum-allo y structure. The dehydrated space between the glass panes usually has a thickness of 0.125 to 0.75 in. (3.2 – 19 mm) and is f lled with air , argon, or other inert gas. Air- or gas-f lled space increases the thermal resistance of the fenestration. Ref ective coated glass . Re f ective glass has a microscopically thin layer of metallic or ceramic coating on one surface of the glass, usually the inner surface of a single-pane glazing or the outer surface of the inner plate for an insulating glass. F or a single pane, the coating is often protected by a layer of transparent polyester . The chromium and other metallic coatings gi ve excellent ref ectivity in the infrared re gions but reduced transmission of visible light compared to clear plate and heat-absorbing glass. Re f ections from b uildings with highly re f ective glass may blind drivers, or even kill grass in neighboring yards. Low-emissivity (low-E) glass coatings . Glazing coated with lo w-emissivity, or low-E, f lms has been in use since 1978. It is widely used in retro f t applications. A low-emissivity f lm is usually a v acuum-deposited metallic coating, usually aluminum, on a polyester f lm, at a thickness of about 4 107 in. (0.01 m). Because of the fragility of the metal coating, protection by another polyester f lm against abrasion and chemical corrosion must be pro vided. Recently, copper and

3.30

CHAPTER THREE

silver coatings on polyethylene and polyprop ylene f lm for protection ha ve been used for better optical transmission. A low-E f lm coating reduces the U value about 25 to 30 percent for single panes. When combined with other solar control de vices, these f lms can reduce solar heat gain further.

Optical Properties of Sunlit Glazing When solar radiation inpinges on the outer surf ace of a plate of glass with an intensity of I at an angle of incidence , as shown in Fig. 3.9, a portion of the solar radiation is transmitted, another portion is ref ected from both inner and outer surfaces, and the remaining portion is absorbed. Let r indicate the portion re f ected and a the portion absorbed. Also let  be the angle of incidence. In Fig. 3.9, it can be seen that the portion of the solar radiation transmitted through the glass is actually the sum of the transmittals after successi ve multiple ref ections from the outer and inner surfaces. The decimal portion of I transmitted through the glass, or , represents the transmittance of the window glass. Similarly, the portion of solar radiation re f ected from the window glass is the sum of the successi ve ref ections from the outer surf aces after multiple re f ections and absorptions,

FIGURE 3.9 Simplif ed representation of multiple transmissions, ref ections, and absorptions of solar radiation at glass surfaces.

HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE

3.31

FIGURE 3.10 Spectral transmittance for various types of window glasses. (Source: ASHRAE Handbook 1989, Fundamentals. Reprinted by permission. )

and it is identi f ed as , the ref ectance. The portion absorbed is the sum of the successi ve absorptions within the glass , or its absorptance. Figure 3.10 illustrates the spectral transmittance of se veral types of glazing. All are transparent for shortwave solar radiation at a w avelength between 0.29 and 3 m and are opaque to longw ave radiation in the infrared range with a w avelength greater than 3 m. Most interior furnishings, equipment, and appliances ha ve an outer surf ace temperature lo wer than 250 °F (120°C), emitting almost all longwave radiation. At such temperatures, glass is opaque to longw ave radiation emitted from inside surfaces and lets only shortwave radiation through. This trapping of longwave radiation indoors is called the greenhouse effect. In Fig. 3.10 one can also see that clear plate glass has a high transmittance   0.87 for visible light, and that   0.8 for infrared from 0.7 to 3 m. Heat-absorbing glass has a lower  and higher absorptance  for both visible light and infrared radiation. Bluish-green heat-absorbing glass has a higher  in the visible light range and a lo wer  in the infrared range than gray heat-absorbing glass. Some ref ective glazing has a high ref ectance  and a signif cantly lower  in the visible light range and is opaque to radiation at w avelengths greater than 2 m. Such characteristics for heatabsorbing and re f ective glasses are ef fective for reducing the amount of solar heat entering the conditioned space during cooling as well as heating seasons. This f act sometimes presents a dilemma for the designer , who must f nally compromise to get the optimum combination of conf icting properties. Another important property of glass is that both  and  decrease and  increases sharply as the incident angle  increases from 60 to 90°. At 90°,  and  are 0 and  is equal to 1. That is why the solar radiation transmitted through vertical glass declines sharply at noon during summer with solar altitudes   70°. For all types of plate glass, the sum of these radiation components is al ways equal to 1, that is,

1 and I  I  I  I

(3.67)

3.32

CHAPTER THREE

3.10 HEAT ADMITTED THROUGH WINDOWS For e xternal glazing without shading, the heat gain admitted into the conditioned space through each square foot of sunlit area As of window Qwi /As, Btu /hft2 (W/m2), can be calculated as follows: inward heat f ow from glass Heat gain through each ft2 solar radiation transmitted   of sunlit window through window glass inner surface into conditioned space That is,

It  Q RCi Q wi  As As

(3.68)

where QRCi  inward heat f ow from the inner surface of an unshaded sunlit window, Btu/h (W). Heat Gain for Single Glazing For an e xternal, sunlit single-glazed windo w without shading, the inward heat f ow from the inner surface of the glass, as shown in Fig. 3.11, can be evaluated as QRCi  inward absorbed radiation  conductive heat transfer  UAs

It

h

o



 To  Ti

FIGURE 3.11 Heat admitted through a single-glazing window glass.

(3.69)

HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE

3.33

where ho  heat-transfer coeff cient for outdoor surface of window glass, Btu/h ft2 °F (W/m2 °C) To  outdoor air temperature, °F (°C) Ti  indoor air temperature,°F (°C) Heat admitted through a unit area of the single-glazing window glass is Q wi  It  U As

It

h

o



 To  Ti

 SHGCIt  U(To  Ti)

(3.70)

Solar heat gain coef f cient (SHGC) is the ratio of solar heat gain entering the space through the window glass to the incident solar radiation, total shortwave irradiance for a single-glazed windo w is given as SHGC Q ws / It As 

U

(3.71)

ho

In these last three equations, U indicates the o verall heat-transfer coef f cient of the windo w, in Btu/h  ft2 °F (W/m2 °C), and can be calculated as U

Uwg Awg  Ueg Aeg  Uf Af

(3.72)

Awg  Aeg  Af

In Eq. (3.72), A represents area, in square feet (square meters); subscript wg indicates glass, and eg signif es the edge of the glass including the sealer and spacer of the insulating glass. The edge of glass has a width of about 2.5 in. (64 mm). The subscript f means the frame of the window. Some U values for various types of windows at winter design conditions are listed in Table 3.7. For summer

TABLE 3.7 Overall Heat-Transfer Coeff cient U Values for Windows at Winter Conditions* with Commercial Type of Frame, Btu/h ft2  °F)

Type

Gas between glasses

Space between glasses, in.

Emittance† of low-E f lm

Glass

Edge

Aluminum frame of Uf  1.9

Aluminum frame with thermal break, with Uf  1.0

Wood or vinyl frame of Uf  0.41

Overall coeff cient Btu/h  ft2  °F Single glass Double glass Double glass Double glass Double glass Triple glass Triple glass Triple or double glass with polyester f lm suspended in between

Air Air Air Argon Air Air

3

Argon

3

/8 3 /8 3 /8 3 /8 3 /8 3 /8

/8

0.40

1.11 0.52 0.43 0.36 0.30 0.34 0.30

0.62 0.55 0.51 0.48 0.50 0.48

1.23 0.74 0.67 0.62 0.57 0.60 0.57

1.10 0.60 0.54 0.48 0.43 0.46 0.44

0.98 0.51 0.45 0.39 0.34 0.38 0.35

0.15

0.17

0.43

0.47

0.34

0.25

0.40 0.15 0.15

* Winter conditions means 70°F indoor, 0°F outdoor temperature and a wind speed of 15 mph. † Low-E f lm can be applied to surface 2 for double glass (see Fig. 3.12b) or surface 2, 3, 4, or 5 for triple glass (any surface other than outer and inner surfaces). Source: Abridged with permission from ASHRAE Handbook, 1989, Fundamentals.

3.34

CHAPTER THREE

design conditions at 7.5 mph (3.3 m /s) wind speed, the listed U values should be multiplied by a factor of 0.92. The U values of windows depend on the construction of the windo w, the emissivity of the surf aces of glass or plastic sheets, and the air v elocity f owing over the outdoor and indoor surfaces. Qws indicates the solar heat gain entering the space, in Btu/h (W). As shown in Fig. 3.11, solar radiation having a total intensity I, Btu/h ft2 (W/m2), when it is radiated on the outer surf ace of a vertical pane with an angle of incidence v, the line normal indicating total shortwave irradiance It, in Btu/h ft2 (W/m2) actually consists of I  IDN  Id  Iref. DN It  I cosv

(3.73)

Heat Gain for Double Glazing For a double-glazed window, the inward heat f ow per square foot of the inner surf ace of the glass, as shown in Fig. 3.12b, is Q RCi  NioIo  NiiIi  U(To  Ti) As U

Io

h

o



 h1

o



1 ha

 I  T  T  i

o

i

(3.74)

FIGURE 3.12 Heat f ow through an insulating glass (double-pane) windo w. (a) Construction of a typical insulating glass; ( b) heat f ow and temperature prof les.

HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE

3.35

where Nio  inward fraction of solar radiation absorbed by outdoor glass Nii  inward fraction of solar radiation absorbed by indoor glass Io, Ii  solar intensity irradiated on outdoor and indoor glass, Btu/hft2 (W/m2) The solar radiation absorbed by the outdoor and indoor panes, in Btu/h ft2 (W/m2), is

Io  o It

Ii  i It

(3.75)

Absorptance of the outdoor panes o is

o  1 

2oi 3

(3.76)

1  2 3

Absorptance of the indoor pane is

i 

3o

(3.77)

1  2 3

Subscript o indicates outdoor; i, indoor; and 1, 2, 3, and 4, the surfaces of the panes as sho wn in Fig. 3.12b. The airspace heat-transfer coef f cient ha, Btu / hft2 °F (W/ m2 °C), is the reciprocal of the R value of the airspace Ra listed in Table 3.4, that is, ha 

1 Ra

(3.78)

The transmittance of solar radiation through both outdoor and indoor panes as oi oi  1  2 3

oi can be calculated

Then the heat admitted per square foot (square meter) through a double-glass windo Btu/h  ft2 (W/m2), can be calculated as

(3.79) w Qwoi /As

o i i Q woi  oi It  U    It  U(To  Ti) As h h ha





 SHGCoi It  U(To  Ti)

(3.80)

The SHGCoi for a double-glass window can be calculated as SHGCoi  oi  U

o

h

o



i



(3.81)

ha

Because a glass plate is usually between 0.125 and 0.25 in. (3 and 6 mm) thick, with a thermal conductivity k of about 0.5 Btu /h ft °F (W/m °C), there is only a small temperature dif ference between the inner and outer surf aces of the plate when solar radiation is absorbed. F or the sak e of simplicity, it is assumed that the temperature of the plate Tg is the same in the direction of heat f ow. For a double-glazed windo w, the glass temperature of the outdoor pane Tgo, °F ( °C), can be calculated as



Tgo  To  Io  Ii 

Q RCi As

 h1

o



R go 2



(3.82)

3.36

CHAPTER THREE

The temperature of the indoor pane Tgi, °F (°C), is calculated as Tgi  Tr  Q RCi

 h1

i



R gi 2



(3.83)

where Rgo, Rgi  R values of outdoor and indoor panes, h ft2 °F/Btu (m2 °C/W) hi  heat-transfer coeff cient of inside surface 4 (see Fig. 3.12b) of inside plate, Btu/h ft2 °F (W/m2 °C) Shading Coefficients The shading coeff cient is def ned as the ratio of solar heat gain of a glazing assembly of speci f c construction and shading de vices at a summer design solar intensity and outdoor and indoor temperatures, to the solar heat gain of a reference glass at the same solar intensity and outdoor and indoor temperatures. The reference glass is double-strength sheet glass (DSA) with transmittance   0.86, ref ectance   0.08, absorptance   0.06, and FDSA  0.87 under summer design conditions. The shading coef f cient SC is an indication of the characteristics of a glazing and the associated shading devices, and it can be expressed as SC  

solar heat gain of specific type of window glass solar heat gain of double-strength sheet glass SHGCw SHGCw   1.15 SHGCwi SHGCDSA 0.87

(3.84)

where SHGCw  solar heat gain coeff cient of specif c type of window glass SHGCDSA  solar heat gain coeff cient of standard reference double-strength sheet glass Shading coeff cients of various types of glazing and shading devices are presented in Table 3.8.

TABLE 3.8 Shading Coeff cients for Window Glass with Indoor Shading Devices

Type of glass Clear Heat-absorbing

Thickness of glass, in.

Solar transmittance

3 32

0.87 to 0.79

3 8

0.34 0.24

3 16

1 or 0.46 4

Ref ective-coated Insulating glass Clear out Clear in Heat-absorbing out Clear in Ref ective glass

3 32 ,

or 81 1 4

Outer Inner 0.87 0.87 0.46

Venetian blinds

Roller shade

Draperies

Glass

Med.*

Light

Opaque, white

Translucent

Med.‡

Light§

0.67 0.53 0.52 0.40 0.23 0.29 0.38

0.39 0.30 0.28 0.28

0.44 0.36 0.32 0.31

0.62 0.46

0.52 0.44

0.30 0.40 0.50

0.74 0.57 0.54 0.42 0.25 0.33 0.42

0.38

0.36

0.62

0.58

0.35

0.40

0.39

0.36

0.22

0.30

0.19 0.27 0.34

0.18 0.26 0.33

0.80 0.20 0.30 0.40

†

*Med. indicates medium color. † Light indicates light color. ‡ Draperies Med. represents draperies of medium color with a fabric openness of 0.10 to 0.25 and yarn ref ectance of 0.25 to 0.50. § Draperies Light repesents draperies of light color with a fabric openness below 0.10 and yarn ref ectance over 0.50. Source: Adapated with permission from ASHRAE Handbook 1989, Fundamentals.

HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE

3.37

Solar Heat Gain Factors and Total Shortwave Irradiance The solar heat gain f actor (SHGF), Btu/hft2 (W/m2), is designated as the a verage solar heat gain during cloudless days through DSA glass. In the ASHRAE Handbook 1993, Fundamentals are tabulated SHGF v alue for v arious latitudes, solar times, and orientations for load and ener gy calculations. F or calculating the summer cooling peak load, the concept of maximum SHGF has been introduced. This is the maximum v alue of SHGF on the 21st of each month for a speci f c latitude, as listed in Table 3.9. F or high ele vations and on v ery clear days, the actual SHGF may be 15 percent higher than the v alue, listed in Table 3.9. In dusty industrial areas or at v ery humid locations, the actual SHGF may be lower. According to ASHRAE Handbook 1997 , Fundamentals, Gueymard (1995) pro vides a comprehensive model for calculating the spectral and broadband total shortw ave irradiance It, in Btu /h ft2 (W/m2) for cloudless sk y conditions and allo ws the input of the concentrations of a v ariety atmospheric constituents. Example 6.1. A double-glass windo w of a commercial b uilding facing west consists of an outdoor clear plate glass of 0.125-in. and an indoor re f ective glass of 0.25-in. thickness with a re f ective f lm on the outer surface of the indoor glass, as shown in Fig. 3.12b. This building is located at

TABLE 3.9 Maximum Solar Heat Gain Factors (Max SHGF) Max SHGF, Btu/h  ft2 N (shade)

NNE/NNW

NE/NW

ENE/WNW

Jan. Feb. Mar. Apr. May June July Aug. Sept. Oct. Nov. Dec.

20 24 29 34 37 48 38 35 30 25 20 18

20 24 29 71 102 113 102 71 30 25 20 18

20 50 93 140 165 172 163 135 87 49 20 18

74 129 169 190 202 205 198 185 160 123 73 60

Jan. Feb. Mar. Apr. May June July Aug. Sept. Oct. Nov. Dec.

24 27 32 36 38 44 40 37 33 28 24 22

24 27 37 80 111 122 111 79 35 28 24 22

29 65 107 146 170 176 167 141 103 63 29 22

105 149 183 200 208 208 204 195 173 143 103 84

E/W

ESE/WSW

SE/SW

SSE/SSW

S

HOR*

205 234 238 223 208 199 203 214 227 225 201 188

241 246 236 203 175 161 170 196 226 238 237 232

252 244 216 170 133 116 129 165 209 236 248 249

254 241 206 154 113 95 109 149 200 234 250 253

133 180 223 252 265 267 262 247 215 177 132 113

229 242 237 219 199 189 194 210 227 234 225 218

249 248 227 187 155 139 150 181 218 239 245 246

250 232 195 141 99 83 96 136 189 225 246 252

246 221 176 115 74 60 72 111 171 215 243 252

176 217 252 271 277 276 273 265 244 213 175 158

North latitude, 40° 154 186 218 224 220 216 216 216 203 180 151 135

North latitude, 32° 175 205 227 227 220 214 215 219 215 195 173 162

*Horizontal surface Source: Abridged with permission from ASHRAE Handbook 1989, Fundamentals.

3.38

CHAPTER THREE

40° north latitude. The detailed optical properties of their surfaces are as follows:   0.80 i  0.16

1  0.08 2  0.08 3  0.68 4  0.08

1  0.12 2  0.12 3  0.16 4  0.76

e1  0.84 e2  0.84 e3  0.15 e4  0.84

The R value of the 0.25-in. (6-mm) thickness indoor glass is Rg  0.035 h ft2 °F/Btu (0.0063 m2 °C/W), and that of the enclosed airspace is Ra  1.75 h ft2 °F/Btu (0.32 m2 °C/W). At 4 P.M. on July 21, the outdoor temperature at this location is 100 °F (37.8°C), the indoor temperature is 76°F (24.4°C), and the total solar intensity at a direction normal to this west windo w is 248 Btu / hft2 (782 W/ m2). The outdoor surf ace heat-transfer coef f cient ho  4.44 Btu /hft2 °F (25.2 W/m2°C), and the heat-transfer coeff cient of the inner surf ace of the indoor glass hi  2.21 Btu/h ft2 °F (12.5 W/m2 °C). Calculate the following: 1. 2. 3. 4.

The inward heat f ow of this window The temperatures of the outdoor and indoor glasses The shading coeff cient of this double-glass window The total heat gain admitted through this window Solution 1. From Eq. (3.76), the absorption coeff cient of the outdoor glass can be calculated as

o  1 

2o3 1  23

 0.12 

0.12 0.8 0.68  0.189 1  0.08 0.68

And from Eq. (3.77), the absorption coeff cient of the indoor glass is

i 

3o 1  23



0.16 0.8  0.135 1  0.08 0.68

From Eq. (3.75), the heat absorbed by the outdoor glass is

Io  o It  0.189 248  46.9 Btu / h ft 2 Also, the heat absorbed by the indoor glass is calculated as

Ii  a iIt  0.135 248  33.5 Btu / h ft 2 And then, from Eq. (3.22), the overall heat-transfer coeff cient of this double-glass window is U 

1 1/h i  R g  R a  R g  1/h o 1  0.400 Btu / h ft 2 F 1/2.21  0.035  1.75  0.035  1/4.44

HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE

3.39

The inward heat f ow from the inner surface of the indoor glass, from Eq. (3.74), is given as Q RCi U As

o It

 h   I h1  h1  T  T  46.9 1  0.400   33.5  1.75  100  76  40.29 Btu / h ft (127.1 W/ m ) 4.44 4.44 o

i t

o

o

a

r

2

2

2. From Eq. (3.82), the temperature of the outdoor glass is



Tgo  To  o It  i It 

Q RCi As

Rg

h1  2  o

1 0.035   109.7F 4.44 2 

 100  (46.9  33.5  40.29) and the temperature of the indoor glass is Tgi  Ti 

Rg Q RCi 1  As h i 2



 76  40.29



0.035 1   94.9F (34.9C) 2.21 2 

3. From Eq. (3.79), the transmittance for both panes is

oi 

oi 1  23



0.8 0.16  0.135 1  0.08 0.68

and from Eq. (3.81), the solar heat gain coeff cient for the double-glass window is SHGCoi  oi 

Uo ho

 0.135 



hU  hU   o

a

i

0.40 0.189 0.40   0.40 1.75 0.135  0.259 4.44 4.44





Then, from Eq. (3.84), the shading coeff cient of this double-glass window is given by SC 

SHGCoi 0.259   0.297 SHGCDSA 0.87

4. From Eq. (3.80), the total heat gain admitted through this double-glass windo w per square foot of sunlit area is Q wi  SHGCoi It  U(To  Ti)  0.259 248  0.40(100  76) As  64.2  9.6  73.8 Btu / h ft 2 (233 W/ m2) Selection of Glazing During the selection of glazing, the following factors should be considered: visual communication, use of daylight, thermal comfort, summer and winter solar heat gain, street noise attenuation, safety

3.40

CHAPTER THREE

and f re protection, and life-cycle cost analysis. Ener gy conservation considerations, including the control of solar heat with the optimum combination of absorbing and re f ective glass and v arious shading devices, are covered in the next section. To reduce the heat loss through glass during winter , one can install double or triple glazing, storm windows, or low-emission f lm coating on the surf ace of the glass. Elmahdy (1996) and de Abreu et al. (1996) tested the thermal performance of se ven insulating glass units. If a clear , double-glazed insulating glass unit with a silicone foam spacer of 0.5 in. (13 mm) between tw o panes is taken as the base unit, the U values of these seven insulating glass units are as follows:

Unit

Glazing

Airspace, in.

Spacer

U value, Btu/h ft2  °F

1 2 3 4 5 6 7

Clear, double-glazed Clear, double-glazed Clear, double-glazed Clear, double-glazed Low-e, double-glazed Clear, Triple-glazed Clear, Triple-glazed

0.5 0.5 0.25 0.75 0.5 0.5 0.25

Foam Aluminum Foam Foam Foam Foam Foam

0.51 0.51 0.58 0.51 0.36 0.32 0.39

Low-e coating reduces the U value by about 30 percent, and the triple-glazing drops about onethird compared with a clear , double-glazed insulating glass unit. If the width of the airspace is decreased to 0.25 in. (6.5 mm), its U value will increase 15 percent. When the airspace is wider than 0.5 in. (13 mm), regardless of whether a metal or silicone foam spacer is adopted, neither has any signif cant effect on U value of the insulation glass unit.

3.11 SHADING OF GLASS Shading projected o ver the surf ace of glass signi f cantly reduces its sunlit area. Man y shading devices increase the re f ectance of the incident radiation. There are tw o types of shading: deliberately installed shading de vices, which include indoor and e xternal shading de vices, and shading from adjacent buildings.

Indoor Shading Devices Indoor shading de vices not only pro vide privacy but also are usually ef fective in re f ecting part of the solar radiation back to the outdoors. They also raise the air temperature of the space between the shading de vice and the windo w glass, which in turn reduces the conducti ve heat gain in summer. Indoor shading devices are easier to operate and to maintain and are more f exible in operation than external shading devices. Three types of indoor shading de vices are commonly used: venetian blinds, draperies, and roller shades. Venetian Blinds. Most horizontal venetian blinds are made of plastic or aluminum slats, spaced 1 to 2 in. (25 to 50 mm) apart, and some are made of rigid woven cloth. The ratio of slat width to slat spacing is generally 1.15 to 1.25. F or light-colored metallic or plastic slats at a 45 ° angle, typical optical properties are   0.05,   0.55, and   0.40. Vertical venetian blinds with wider slats are widely used in commercial buildings. Consider a single-glazed window combined with indoor venetian blinds at a slat angle   45°, as shown in Fig. 3.13. Let the subscripts g, v, and a represent the glass, the venetian blinds, and the

HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE

3.41

FIGURE 3.13 Heat transfer through a single-glazed windo w combined with venetian blinds.

air between the v enetian blinds and the glass. Also let o indicate the outw ard direction and i the inward direction. Of the solar radiation transmitted through the glass and radiating on the surf ace of the slats g It, ●

●

●

A fraction vog It is ref ected outward from the surface of each slat. Another fraction vi g It is ref ected from the slat surface into the conditioned space. A third fraction absorbed by the slat is either con vected away by the space air or re f ected from the surface to the indoor surroundings in the form of longwave radiation.

If the glass has a high transmittance, the slat temperature Tv1 will be higher, as shown by the lower solid temperature curve in Fig. 3.13. If the glass has a high absorptance, its temperature Tg2 will be higher. In Table 3.8 the shading coeff cients of various combinations of venetian blinds and glazing are listed. According to results of a f eld survey by Inoue et al. in four of f ce buildings in Japan in 1988, 60 percent of the v enetian blinds were not operated during the daytime. The incident angle of the direct solar radiation had greater in f uence on operation of the blinds than the intensity . Automatic control of slat angle and of raising and lo wering the blinds is sometimes used and may become more popular in the future. Draperies. These are f abrics made of cotton, regenerated cellulose (such as rayon), or synthetic f bers. Usually the y are loosely hung, wider than the windo w, and pleated; and the y can be dra wn open or closed as required. Drapery-glass combinations reduce the solar heat gain in summer and increase the thermal resistance in winter . Re f ectance of the f abric is the dominant f actor in the reduction of heat gain, and visibility is a function of the openness of the weave. Roller Shades. These are sheets made of treated fabric or plastic that can be pulled down to cover the windo w or rolled up. Roller shades ha ve a lo wer SC than do v enetian blinds and draperies.

3.42

CHAPTER THREE

When glass is co vered with shades, any outdoor visual communication is block ed, and the visible light transmittance is less than that of other indoor shading devices.

External Shading Devices External shading de vices include o verhangs, side f ns, egg-crate louv ers, and pattern grilles, as shown in Fig. 3.14. They are effective in reducing the solar heat gain by decreasing the sunlit area. However, the external shading devices do not al ways f t into the architectural requirements and are less f exible and more diff cult to maintain. Pattern grilles impair visibility signif cantly. Figure 3.15 shows the shaded area of a glass pane constructed with both o verhang and side f n. The prof le angle  is de f ned as the angle between a horizontal plane and a tilted plane that includes the sun ’s rays. We see that tan   UQ/OR  tan  /cos . Let SW be the width of the shadow and SH be the height of the shadow projected on the plane of the glass by direct solar radiation, in feet (meters). Also, let W be the width of the glass and H be the height. Then the shado w width on the plane of the glass is SW  PV tan 

(3.85)

and the shadow height is SH  PH tan   PH

tan cos 

(3.86)

where PV  projection of side f n plus mullion and reveal, ft (m) PH  horizontal projection of overhang plus transom and reveal, ft (m) The projection factor due to the external shading device Fpro can be calculated as Fpro 

PH H

FIGURE 3.14 Types of e xternal shades. ( a) Ov erhang; ( b) e gg-crate louver; (c) pattern grilles.

(3.87)

HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE

3.43

FIGURE 3.15 Shaded area of a window glass constructed with an overhang and side f n.

The net sunlit area of the glass As, ft2 (m2), can then be calculated as As  (W  SW)(H  SH)

(3.88a)

and the shaded area of the glass Ash, ft (m ), is given by 2

2

Ash  Ag  As  WH  (W  SW)(H  SH)

(3.88b)

where Ag  area of the glass, ft2 (m2). Shading from Adjacent Buildings Shadows on glass cast by adjacent b uildings signi f cantly reduce the sunlit area of the glass. For e xample, in Fig. 3.16 we see the area on b uilding A shaded because of the presence of building B. Let tw o sides of the shaded b uilding A coincide with the X and Y axes on the plan view shown in Fig. 3.16. In the ele vation view of Fig. 3.16, the shadow height on the f acade of building A is SH  HB  L AB tan 

(3.89)

and the shadow width SW on the facade of building A is SW  WOB  WB  L AB tan 

(3.90)

3.44

CHAPTER THREE

FIGURE 3.16 Shading from adjacent building.

where HB  height of building B, ft (m) WB  width of building B, ft (m) LAB  distance between two buildings along X axis, ft (m) WOB  distance between X axis and building B, ft (m) Subscript A indicates the shaded building and B the shading building. Because of the sun’s varying position, the solar altitude  and the surface solar azimuth  change their v alues from time to time. Table 3.10 lists the solar altitude  and solar azimuth  at north latitudes 32° and 40°. To evaluate the shaded area of the outer surface of a building, a computer program can determine which of several hundred representative points on this outer surface are sunlit or shaded at specif c times of the day. Then a ratio of shaded area to total area can be calculated. At a certain time instant, a window glass area can be shaded due to the ef fect of the o verhang and vertical f ns, or the effect of the adjacent building, or both. For the calculation of the combined shading effect of overhangs, vertical f ns, and adjacent building, it is recommended that the number of sunlit and shaded windo ws of an outer surf ace of a b uilding under the ef fect of the adjacent building be calculated f rst. According to Eq. (3.88 a), the total sunlit area can be calculated as the product of the net sunlit area for each windo w because of the o verhang and v ertical f ns and the number of windows in the sunlit area of that outer surface.

HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE

3.45

TABLE 3.10 Solar Altitude (ALT) and Solar Azimuth (AZ) North latitude 32° Solar time

40° Solar time

Solar time

Date

AM

Solar position ALT

AZ

PM

AM

ALT

AZ

PM

Dec.

8 9 10 11 12

10 20 28 33 35

54 44 31 16 0

4 3 2 1 12

8 9 10 11 12

5 14 21 25 27

53 42 29 15 0

4 3 2 1 12

Jan.  Nov.

7 8 9 10 11 12

1 13 22 31 36 38

65 56 46 33 18 0

5 4 3 2 1 12

8 9 10 11 12

8 17 24 28 30

55 44 31 16 0

4 3 2 1 12

Feb.  Oct.

7 8 9 10 11 12

7 18 29 38 45 47

73 64 53 39 21 0

5 4 3 2 1 12

7 8 9 10 11 12

4 15 24 32 37 39

72 62 50 35 19 0

5 4 3 2 1 12

Mar.  Sept.

7 8 9 10 11 12

13 25 37 47 55 58

82 73 62 47 27 0

5 4 3 2 1 12

7 8 9 10 11 12

11 23 33 42 48 50

80 70 57 42 23 0

5 4 3 2 1 12

Apr.  Aug.

6 7 8 9 10 11 12

6 19 31 44 56 65 70

100 92 84 74 60 37 0

6 5 4 3 2 1 12

6 7 8 9 10 11 12

7 19 30 41 51 59 62

99 89 79 67 51 29 0

6 5 4 3 2 1 12

May  July

6 7 8 9 10 11 12

10 23 35 48 61 72 78

107 100 93 85 73 52 0

6 5 4 3 2 1 12

5 6 7 8 9 10 11 12

2 13 24 35 47 57 66 70

115 106 97 87 76 61 37 0

7 6 5 4 3 2 1 12

June

5 6 7 8 9 10 11 12

1 12 24 37 50 62 74 81

118 110 103 97 89 80 61 0

7 6 5 4 3 2 1 12

5 6 7 8 9 10 11 12

4 15 26 37 49 60 69 73

117 108 100 91 80 66 42 0

7 6 5 4 3 2 1 12

Source: ASHRAE Handbook 1981, Fundamentals. Reprinted with permission.

Solar position

Solar time

3.46

CHAPTER THREE

3.12 HEAT EXCHANGE BETWEEN THE OUTER BUILDING SURFACE AND ITS SURROUNDINGS Because atmospheric temperature is lower at high altitudes, there is always a radiant heat loss from the outer surface of the building to the sky vault without clouds. However, it may be offset or partly offset by ref ected solar radiation from the ground on a sunn y day. Radiant heat loss from the building needs to be calculated during nighttime and included in year-round energy estimation. In commercial and institutional b uildings using glass, concrete, or face brick on the outside surface of the b uilding en velope, the migration of moisture through the glass pane is rather small. Because of the hea vy mass of the concrete w all, the inf uence of the diurnal c yclic variation of the relative humidity of outdoor air on moisture transfer through the b uilding envelope is also small. Therefore, for simplicity, the moisture transfer between the b uilding envelope and the outside air can be ignored. The heat balance at the outer b uilding surf ace, as sho wn in Fig. 3.17, can be e xpressed as follows: Q sol  Q ref  Q os  (Q rad  Q at)  Q oi

(3.91)

In Eq. (3.91), Qsol represents the solar radiation absorbed by the outer surf ace of the building envelope, in Btu/h (W). It can be calculated as Q sol  os [As (ID  Id )  Ash Id]

(3.92)

where os  absorptance of the outer surf ace of the building envelope. From Eq. (3.61), the ref ection of solar radiation from an y ref ecting surface to the outer surf ace of the b uilding and absorbed by it, or qref , Btu/h (W), is given by Q ref  Aos Iref  Asos Fsr (ID  Id )

(3.93)

where A  total area of the outer surf ace of the b uilding en velope, A  As  Ash, ft (m ). The term qos indicates the con vective heat transfer from the outer surf ace of the b uilding outw ard, in 2

FIGURE 3.17 Heat balance at the outer surface of a building.

2

HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE

3.47

Btu/h (W). From Eq. (3.7), it can be calculated as Q os  h c A(Tos  To)

(3.94)

where Tos , To  outer surface temperature of the building and the outdoor air temperature, respectively,°F (°C) hc  convective heat-transfer coeff cient of outer surface, Btu/h ft2 °F (W/m2 °C) The term Qoi denotes the inward heat f ow from the outer b uilding surface, in Btu /h (W). The term Qrad  Qat indicates the net heat emitted from the outer surf ace of the building because of the radiation e xchange between the surf ace and the atmosphere, in Btu /h (W). Here, Qrad represents the longwave radiation emitted from the surf ace, and Qat indicates the atmospheric radiation to the surface. Kimura (1977) found that atmospheric radiation can be expressed as Q at  (1  Ccc K cc ) T 4RoBr  Ccc K cc  T 4Rg

(3.95)

where Ccc  cloud cover factor, which can be obtained from local climate records, dimensionless Kcc  cloudy reduction factor Usually, the smaller the v alue of Ccc, the higher the clouds and the smaller the Kcc value. For simplicity, it can be calculated as K cc  0.83  0.4Ccc

(3.96)

And Br is actually the emissi vity of the atmosphere, and it can be e xpressed by an empirical formula developed by Brunt: Br  0.51  0.55√pw

(3.97)

where pw  water vapor pressure, psia. In Eq. (3.95), TRg represents the absolute ground temperature, °R (K), and TRo the absolute outdoor temperature, °R (K). Then, radiant heat loss can be written as Q rad  Q at  os AFsat{T 4Ros  [(1  Ccc K cc )T 4Ro Br  Ccc K cc T 4Rg]}

(3.98)

where os  emissivity of outer surface of building Fsat  shape factor between outer surface and atmosphere TRos  absolute outer surface temperature of building, °R (K) Sol-Air Temperature For a sunlit outer surf ace of a b uilding, if Qref is mainly from ground-re f ected solar radiation, it is offset by Q rad  Q at, and Eq. (3.91) becomes Q sol  Q os  Q oi Let Q oi  h o A(Tsol  Tos). Substituting for Qoi , then, gives

os It  h o(Tos  To)  h o(Tsol  Tos) and Tsol  To 

os It ho

(3.99)

In Eq. (3.99), Tsol is called sol-air temperature, in °F (°C). It is a f ctitious outdoor temperature that combines the ef fect of the solar radiation radiated on the outer surf ace of the b uilding and the inward heat transfer due to the outdoor – indoor temperature difference.

3.48

CHAPTER THREE

Example 3.2.

At midnight of July 21, the outdoor conditions of an of f ce building are as follows:

Outdoor temperature Water vapor pressure of outdoor air Ground temperature Cloud cover factor Ccc Emissivity of the outer surface Shape factor between the surface and sky

75°F (23.9°C) 0.215 psia 72°F (22.2°C) 0.1 0.90 0.5

If the outer surf ace temperature of this b uilding is 76 °F (24.4 °C) and the Stef an-Boltzmann constant   0.1714 108 Btu/h ft2 °R4 (5.67 108 W/m2 K4), f nd the radiant heat loss from each square foot (square meter) of the vertical outer surface of this building. Solution. From Eq. (3.97) and given data, we can see Br  0.51  0.55√pw  0.51  0.55√0.215  0.765 And from Eq. (3.96), the cloudy reduction factor is K cc  0.83  0.4Ccc  0.83  0.4(0.1)  0.79 From the gi ven, TRos  76  460  536°R, TRo  75  460  535°R, and TRg  72  460  532°R; then, from Eq. (3.98), the radiant heat loss from the outer surf ace of the b uilding due to radiant exchange between the surface and the atmosphere is Q rad  Q at  os Fsat {T 4Ros  [(1  Ccc K cc )T 4RoBr  Ccc K cc T 4Rg]}  0.90 0.5 0.174{(5.36)4  [(1  0.1 0.79)(5.35)4 0.765  0.1 0.79 (5.32)4]}  14.27 Btu / h ft 2 (45.02 W/ m2)

3.13 COMPLIANCE WITH ASHRAE/IESNA STANDARD 90.1-1999 FOR BUILDING ENVELOPE An energy-eff cient and cost-ef fective building envelope should meet the requirements and design criteria in ASHRAE/IESNA Standard 90.1-1999, Energy Standard for Buildings Except Lo w-Rise Residential Buildings; the DOE Code of Federal Regulations, Title10, Part 435, Subpart A, Performance Standard for New Commercial and Multi-Family High-Rise Residential Buildings; and local energy codes. The DOE Code is very similar to Standard 90.1. The design and selection of the building envelope are generally the responsibility of an architect with the assistance of a mechanical engineer or contractor. Building envelopes are usually designed, or e ven constructed, before the HV AC&R system is designed. A speculati ve b uilding is b uilt of known use and type of occupanc y, but the e xact tenants are unkno wn. A shell b uilding is b uilt before the use and occupancy are determined. Compliance with the ASHRAE/IES Standard 90.1-1999 for building envelopes includes the following. Refer to Standard 90.1-1999 for exceptions and details. General Requirements The requirements apply to the e xterior building envelope which separates conditioned space from the outdoors and the semie xterior b uilding envelope which separates the conditioned space from semiheated space or unconditioned space, or separated semiheated space from the unconditioned space or outdoors. Semiheated space is an enclosed space, but not a conditioned space, within a b uilding that is heated by a heating system whose output capacity is greater than or equal to 3.4 Btu /h ft (10.71

HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE

3.49

W/m2). Compliance includes mandatory pro visions, the prescriptive building envelope option, and the building envelope trade-off option. Mandatory Provisions The standard mandates that insulation materials shall be installed in accordance with the manuf acturer’s recommendation to achie ve rated R value of insulation. Rated R value of insulation is the thermal resistance of the insulation alone in units of h ft2 °F/Btu (m 2 °C/W) at a mean temperature of 75 °F (23.9 °C). Rated R value refers to the thermal resistance of the added insulation in framing cavities or insulated sheathing only and does not include the thermal resistance of other building materials or air f lms. Insulation shall be installed in a permanent manner in substantial contact with the inside surface. The roof insulation shall not be installed on a suspended ceiling with remo vable ceiling panels. Insulation outdoors shall be co vered with a protecti ve material to pre vent damage from sunlight, moisture, landscaping operations, equipment maintenance, and wind. Insulation materials contacts with the ground shall have a water absoption rate no greater than 0.3 percent. Fenestration performance shall be determined from production line units or representati ve units purchased. U factors shall be determined in accordance with National Fenestration Rating Council (NFRC) 100. Solar heat gain coef f cient (SHGC) for the o verall fenestration area shall be determined in accordance with NFRC 200. Visible light transmittance shall be determined in accordance with NFRC 300. The following areas of building envelope shall be sealed, caulked, gasketed, or weather-stripped until air leakage is minimal: ●

●

●

●

●

●

●

Joints around fenestration and door frames Junctions between w alls and foundations, between w all corners, between w alls and roofs, wall and f oors, or walls and panels Openings because of penetrations of utility services through roofs, walls, and f oors Fenestrations and doors built at site Building assemblies used as ducts and plenums Joints, seams, and penetrations due to vapor retarders Other openings in the building envelope

Air leakage of fenestration and doors shall be determined in accordance with NFRC 400. Air leakage shall not exceed 1.0 cfm/ft 2 (5.0 L/s m2) for glazed swinging entrance doors and re volving doors and also 0.4 cfm/ft 2 (2.0 L/s m2) for all other products. A door that separates conditioned space from the e xterior shall be protected with an enclosed v estibule with all doors into and out of the vestibule installed with self-closing devices. All mandatory pro visions in Standard 90.1-1999 will be presented in the form “standard mandates that”, whereas for nonmandatory provisions, only “standard specif cs that” will be used. Prescriptive Building Envelope Option The exterior building envelope shall comply with the requirements for the appropriate conditioned space in Table 5.3 for various climate (located in Normative Appendix B of Standard 90.1-1999). In Table B1 to B26 of I-P edition: ●

●

●

There are 26 tables, each of them has a number of heating de gree days of 65 °F (HDD65), and a number of cooling de gree days of 50 °F (CDD 50). Select a table with number of HDD 65 and CDD 50 equal to or most nearly equal to the values where the proposed building locates. There are three kinds of conditioned space: nonresidential, residential, semiheated. Total vertical fenestration area, including both f xed and operable fenestration shall be less than 50 percent of the gross w all area. The total sk ylight area, including glass and plastic sk ylights with or without a curb, shall be less than 5 percent of the roof area.

3.50

CHAPTER THREE

Fenestration, including f xed and operable vertical fenestration, shall have a U factor not greater than that speci f ed in Table 5.3, 90.1-1999. Vertical fenestration shall ha ve an SHGC not greater than that specif ed for all orientations in Table 5.3, 90.1-1999. There are only visible light transmittance criteria in the Building Envelope Trade-Off Option. ●

●

●

●

All roofs shall ha ve a rated R value of insulation not less than that speci f ed in Table 5.3,90.11999. Skylight curbs shall be insulated to the le vel of the roofs with the insulation entirely abo ve the deck or R-5, whichever is less. All abo ve-grade w alls shall ha ve a rated v alue of insulation not less than that speci f ed in Table 5.3, 90.1-1999. Mass w all heating capacity shall be determined from Table A-6 or A-7 in Standard 90.1-1999. Below-grade walls shall have a rated R value of insulation not less than that specif ed in Table 5.3, 90.1-1999. All f oors and heated or unheated slab-on-grade f oors shall have a rated R value of insulation not less than that speci f ed in Table 5.3, 90.1-1999. Slab-on-grade f oor insulation shall be installed around the perimeter of the slab-on-grade f oor to the distance specif ed. All opaque doors shall have a U factor not greater than that specif ed in Table 5.3, 90.1-1999.

Building Envelope Trade-Off Option The building envelope complies with the standard if the proposed b uilding satisf ed the pro visions of 5.1 and 5.2 of Standard 90.1-1999 and the en velope performance factor of the proposed building is less than or equal to the envelope performance factor of the budget building. The envelope performance f actor considers only the b uilding en velope components. Schedules of operation, lighting power input, equipment power input, occupant density, and mechanical systems shall be the same for both the proposed b uilding and b udget building. Envelope performance f actors shall be calculated using the procedures of Normati ve Appendix C, Standard 90.1-1999. Refer to Standard 90.11999 for details.

3.14 ENERGY-EFFICIENT AND COST-EFFECTIVE MEASURES FOR BUILDING ENVELOPE According to the economic parametric analysis of the thermal design in Johnson et al. (1989), the following are ener gy-eff cient and cost-ef fective measures for the design of b uilding envelopes for off ce buildings. Exterior Walls An increase in the mass of the e xterior w all, i.e., its thermal capacitance, reduces only the peak heating and cooling loads. The increase in electrical usage and capital in vestment often offsets the benef t of the decrease of peak loads. The variation of life-cycle costing is often negligible. Increasing the insulation to more than 2-in. (50-mm) thick decreases only the natural gas usage for heating. This is cost-effective for areas with very cold winters. In some cases, the increase in the annual cost due to the increase in the capital in vestments of the insulation may balance the reduction of gas usage. A careful analysis is required. Windows The vertical fenestration area ratio (VFR) is the single parameter that most in f uences the building life-cycle costing among building envelopes for high-rise buildings. For many off ce buildings, VFR lies between 0.2 and 0.3.

HEAT AND MOISTURE TRANSFER THROUGH BUILDING ENVELOPE

3.51

Heat-absorbing and -ref ective glasses produce a signif cant building cost savings in areas where solar heat control is important in summer . Double-panes, triple-panes, and low-emission f lms are effective in reducing the U value of the windo w assembly and therefore, the heating and cooling loads. Indoor shading de vices with windo w management systems are cost-ef fective. For example, an indoor shading de vice can be turned on when solar heat gain e xceeds 20 Btu /hft2 (63 W/ m2). Although overhangs reduce the cooling loads, they may increase the need for electric lighting for daylit b uildings. The ef fect of o verhang usage on b uilding life-c ycle costing is not signi f cant in many instances.

Infiltration Inf ltration has a signif cant inf uence on heating and cooling loads. Indoor air quality must be guaranteed by mechanical ventilation systems and suff cient outdoor air intake through these systems. It is desirable that windo ws and cracks in joints be well sealed. In multiple-story b uildings, inf ltration through elevator shafts, pipe shafts, and duct shafts should be reduced.

Energy-Efficient Measures for Commercial Buildings in the United States According to the EIA ’s Commercial Buildings Char acteristics, in 1992, the breakdo wn of energy conservation features for commercial b uildings for a total area of 67,876 million ft 2 (6308 million m2) in the United States is as follows: Roof or ceiling insulation Wall insulation Storm or multiple glazing Tinted, ref ective, or shading glass Exterior or interior shading devices Windows that are openable

74% 49% 44% 37% 50% 43%

REFERENCES de Abreu, P. F., Fraser, R. A., Sullivan, H. F., and Wright, J. L., A Study of Insulated Glazing Unit Surface Temperature Prof les Using Two-Dimensional Computer Simulation, ASHRAE Transactions, 1996, Part II, pp. 497 – 507. Altmayer, E. F., Gadgil, A. J., Bauman, F. S., and Kammerud, R. C., Correlations for Convective Heat Transfer from Room Surfaces, ASHRAE Transactions, 1983 Part II A, pp. 61 – 77. ASHRAE, ASHRAE Handbook 1989, Fundamentals, Atlanta, GA, 1989. ASHRAE, ASHRAE Handbook 1997, Fundamentals, ASHRAE Inc., Atlanta, GA, 1997. ASHRAE, ASHRAE/IESNA Standard 90.1-1999, Energy Standard for Buildings Except Low-Rise Residential Buildings, Atlanta, GA, 1999. ASHRAE, Procedure for Determining Heating and Cooling Loads for Computerizing Energy Calculations, Algorithms for Building Heat Transfer Subroutines, Atlanta, GA, 1976. Bauman, F., Gadgil, A., Kammerud, R., Altmayer, E., and Nansteel, M., Convective Heat Transfer in Buildings: Recent Research Results, ASHRAE Transactions, 1983, Part I A, pp. 215 – 233. Chandra, S., and Kerestecioglu, A. A., Heat Transfer in Naturally Ventilated Rooms: Data from Full Scale Measurements, ASHRAE Transactions, 1984, Part I B, pp. 211 – 225. Dahlen, R. R., Low-E Films for Window Energy Control, ASHRAE Transactions, 1987, Part I, pp. 1517 – 1524. Deringer, J. J., An Overview of Standard 90.1: Building Envelope, ASHRAE Journal, no. 2, 1990, pp. 30 – 34.

3.52

CHAPTER THREE

Donnelly, R. G., Tennery, V. J., McElroy, D. L., Godfrey, T. G., and Kolb, J. O., Industrial Thermal Insulation, An Assessment, Oak Ridge National Laboratory Report TM-5283, TM-5515, and TID-27120, 1976. Energy Information Administration, Commercial Buildings Characteristics 1992, Commercial Buildings Energy Consumption Survey April 1994, Washington, 1994. Elmahdy, H., Surface Temperature Measurement of Insulating Glass Units Using Infrared Thermography, ASHRAE Transactions, 1996, Part II, pp. 489 – 496. Fairey, P. W., and Kerestecioglu, A. A., Dynamic Modeling of Combined Thermal and Moisture Transport in Buildings: Effect on Cooling Loads and Space Conditions, ASHRAE Transactions, 1985 Part II A, pp. 461 – 472. Galanis, N., and Chatigny, R., A Critical Review of the ASHRAE Solar Radiation Model, ASHRAE Transactions, 1986, Part I A, pp. 410 – 419. Glicksman, L. R., and Katsennelenbogen, S., A Study of Water Vapor Transmission Through Insulation under Steady State and Transient Conditions, ASHRAE Transactions, 1983, Part II A, pp. 483 – 499. Gueymard, C. A., A Simple Model of the Atmospheric Radiative Transfer of Sunshine: Algorithms and Performance Assessment, Report FSEC-PF-270-95, Florida Solar Energy Center, Cocoa, Fla., 1995. Hagentoft, C-E., Moisture Conditions in a North-Facing Wall with Cellulose Loose-Fill Insulation: Construction with and without a Vapor Retarder and Air Leakage, ASHRAE Transactions, 1995, Part I, pp. 639 – 646. Handegrod, G. O. P., Prediction of the Moisture Performance of Walls, ASHRAE Transactions, 1985, Part II B, pp. 1501 – 1509. Inoue, T., Kawase, T., Ibamoto, T., Takakusa, S., and Matsuo, Y., The Development of an Optimal Control System for Window Shading Devices Based on Investigations in Off ce Buildings, ASHRAE Transactions, 1988, Part II, pp. 1034 – 1049. Johnson, C. A., Besent, R. W., and Schoenau, G. J., An Economic Parametric Analysis of the Thermal Design of a Large Off ce Building under Different Climatic Zones and Different Billing Schedules, ASHRAE Transactions, 1989, Part I, pp. 355 – 369. Kays, M. M., and Crawford, M. E., Convective Heat and Mass Transfer, 2d ed., McGraw-Hill, New York, 1980. Kimura, K., Scientif c Basis of Air Conditioning, Applied Science Publishers, London, 1977. Kimura, K., and Stephenson, D. G., Solar Radiation on Cloudy Days, ASHRAE Transaction, 75(1), 1969, pp. 1 – 8. Miller, A., Thompson, J. C., Peterson, R. E., and Haragan, D. R., Elements of Meteorology, 4th ed., Bell and Howell Co., Columbus, Ohio, 1983. Robertson, D. K., and Christian, J. E., Comparison of Four Computer Models with Experimental Data from Test Buildings in Northern New Mexico, ASHRAE Transactions, 1985, Part II B, pp. 591 – 607. Sato, A., Eto, N., Kimura, K., and Oka, J., Research on the Wind Variation in the Urban Area and Its Effects in Environmental Engineering No. 7 and No. 8 — Study on Convective Heat Transfer on Exterior Surface of Buildings, Transactions of Architectural Institute of Japan, no. 191, January 1972. Smolenski, C. P., Absorption in Thermal Insulation: How Much Is Too Much? HPAC no. 11, 1996, pp. 49 – 58. Spitler, J. D., Pedersen, C. O., and Fisher, D. E., Interior Convective Heat Transfer in Buildings with Large Ventilative Flow Rates, ASHRAE Transactions, 1991, Part I, pp. 505 – 514. Stewart, W. E., Effect of Air Pressure Differential on Vapor Flow through Sample Building Walls, ASHRAE Transactions, 1998, Part II, pp. 17 – 24. Verschoor, J. D., Measurement of Water Vapor Migration and Storage in Composite Building Construction, ASHRAE Transactions, 1985, Part II A, pp. 390 – 403. Wang, S. K., Air Conditioning, vol. 1, Hong Kong Polytechnic, Hong Kong, 1987. Wong, S. P. W., Simulation of Simultaneous Heat and Moisture Transfer by Using the Finite Difference Method and Verif ed Tests in a Test Chamber, ASHRAE Transactions, 1990, Part I, pp. 472 – 485. Wong, S. P .W., and Wang, S. K., Fundamentals of Simultaneous Heat and Moisture Transfer between the Building Envelope and the Conditioned Space Air, ASHRAE Transactions, 1990, Part II, pp. 73– 83.

CHAPTER 4

INDOOR AND OUTDOOR DESIGN CONDITIONS 4.1 INDOOR DESIGN CONDITIONS 4.1 4.2 HEAT EXCHANGE BETWEEN HUMAN BODY AND INDOOR ENVIRONMENT 4.2 Two-Node Model of Thermal Interaction 4.2 Steady-State Thermal Equilibrium 4.3 Transient Energy Balance 4.3 4.3 METABOLIC RATE AND SENSIBLE HEAT LOSSES FROM HUMAN BODY 4.4 Metabolic Rate 4.4 Mechanical Work 4.4 Sensible Heat Exchange 4.5 Clothing Insulation 4.7 4.4 EVAPORATIVE HEAT LOSSES 4.7 Respiration Losses 4.7 Evaporative Heat Loss from Skin Surface 4.7 Maximum Evaporative Heat Loss due to Regulatory Sweating 4.7 Diffusion Evaporative Heat Loss and Total Skin Wetness 4.8 4.5 MEAN RADIANT TEMPERATURE AND EFFECTIVE TEMPERATURE 4.9 Mean Radiant Temperature 4.9 Effective Temperature 4.14 4.6 FACTORS AFFECTING THERMAL COMFORT 4.14 4.7 THERMAL COMFORT 4.15 Fanger’s Comfort Equation 4.15 ASHRAE Comfort Zones 4.17 Comfort-Discomfort Diagrams 4.17 4.8 INDOOR AIR TEMPERATURE AND AIR MOVEMENTS 4.20

Comfort Air Conditioning Systems 4.20 Design Considerations 4.21 Indoor Design Temperatures for Comfort Air Conditioning 4.21 Process Air Conditioning Systems 4.23 4.9 HUMIDITY 4.23 Comfort Air Conditioning Systems 4.23 Process Air Conditioning Systems 4.24 4.10 SICK BUILDING SYNDROME AND INDOOR AIR QUALITY 4.27 Indoor Air Contaminants 4.27 Basic Strategies to Improve Indoor Air Quality 4.29 Outdoor Air Requirements for Occupants 4.30 4.11 AIR CLEANLINESS 4.31 4.12 SOUND LEVEL 4.32 Sound and Sound Level 4.32 Sound Power Level and Sound Pressure Level 4.32 Octave Bands 4.33 Addition of Sound Levels 4.33 Human Response and Design Criteria 4.34 4.13 SPACE PRESSURE DIFFERENTIAL 4.37 4.14 OUTDOOR DESIGN CONDITIONS 4.38 Summer and Winter Outdoor Design Conditions 4.39 The Use of Outdoor Weather Data in Design 4.39 Outdoor Weather Characteristics and Their Influence 4.42 REFERENCES 4.42

4.1 INDOOR DESIGN CONDITIONS Indoor design parameters are those that the air conditioning system influences directly to produce required conditioned indoor environment in buildings. They are shown below and grouped as follows: 1. Basic design parameters ● ●

Indoor air temperature and air movements Indoor relative humidity 4.1

4.2

CHAPTER FOUR

2. Indoor air quality ● ● ●

Air contaminants Outdoor ventilation rate provided Air cleanliness for processing

3. Specific design parameter ● ●

Sound level Pressure differential between the space and surroundings

The indoor design parameters to be maintained in an air conditioned space are specified in th design document and become the tar gets to be achie ved during operation. In specifying the indoor design parameters, the following points need to be considered: 1. Not all the parameters already mentioned need to be specified in very design project. Except for the indoor air temperature which is al ways an indoor design parameter in comfort air conditioning, it is necessary to specify only the parameters which are essential to the particular situation concerned. 2. Even for process air conditioning systems, the thermal comfort of the w orkers should also be considered. Therefore, the indoor design parameters re garding health and thermal comfort for the occupants form the basis of design criteria. 3. When one is specifying indoor design parameters, economic strategies of initial in vestment and energy consumption of the HV AC&R systems must be carefully in vestigated. Design criteria should not be set too high or too lo w. If the design criteria are too high, the result will be an e xcessively high investment and energy cost. Design criteria that are too lo w may produce a poor indoor air quality, resulting in complaints from the occupants, causing low-quality products, and possibly leading to expensive system alternations. 4. Each specified indoor design parameter is usually associated with a tolerance indicated as a  sign, or as an upper or lo wer limit. Sometimes there is a traditional tolerance understood by both the designers and the o wners of the building. For instance, although the summer indoor design temperature of a comfort air conditioning system is specif ied at 75 or 78°F (23.9 or 25.6°C), in practice a tolerance of  2 – 3°F (  1.1 – 1.7°C) is often considered acceptable. 5. In process air conditioning systems, sometimes a stable indoor environment is more important than the absolute value of the indoor parameter to be maintained. For example, it may not be necessary to maintain 68°F (20°C) for all areas in precision machinery manuf acturing. More often, a 72°F (22.2°C) or e ven a still higher indoor temperature with appropriate tolerance will be more suitable and economical.

4.2 HEAT EXCHANGE BETWEEN HUMAN BODY AND INDOOR ENVIRONMENT Two-Node Model of Thermal Interaction In 1971, Gagge et al. recommended a two-node model of human thermal interaction. In this model, the human body is composed of tw o compartments: an inner body core, including skeleton, muscle and internal or gans; and an outer shell of skin surf ace. The temperatures of the body core and the surface skin are each assumed to be uniform and independent. Metabolic heat production, external mechanical work, and respiratory losses occur only in the body core. Heat e xchange between the body core and the skin surf ace depends on heat conduction from direct contact and the peripheral blood fl w of the thermoregulatory mechanism of the human body.

INDOOR AND OUTDOOR DESIGN CONDITIONS

4.3

Steady-State Thermal Equilibrium When the human body is maintained at a steady-state thermal equilibrium, i.e., the heat storage at the body core and skin surf ace is approximately equal to zero, then the heat e xchange between the human body and the indoor en vironment can be e xpressed by the follo wing heat balance equation: M  W  C  R  Esk  Eres

(4.1)

where M  metabolic rate, Btu/hft2 (W/m2) W  mechanical work performed, Btu/hft2 (W/m2) C  R  convective and radiative, or sensible heat loss from skin surface, Btu/h ft2 (W/m2) Esk  evaporative heat loss from skin surface, Btu/h  ft2 (W/m2) Eres  evaporative heat loss from respiration, Btu/h  ft2 (W/m2) In Eq. (4.1), the ft 2 in the unit Btu /h ft2 applies to the skin surf ace area. The skin surface area of a naked human body can be approximated by an empirical formula proposed by Dubois in 1916 AD  0.657mb0.425 Hb0.725

(4.2)

where AD  Dubois surface area of naked body, ft2 (m2) mb  mass of human body, lb (kg) Hb  height of human body, ft (m) In an air conditioned space, a steady-state thermal equilibrium is usually maintained between the human body and the indoor environment.

Transient Energy Balance When there is a transient ener gy balance between the human body and the indoor en vironment, the thermal interaction of the body core, skin surface, and indoor environment forms a rate of positi ve or negative heat storage both in the body core and on the skin surface. The human body needs energy for physical and mental activity. This energy comes from the oxidation of the food tak en into the human body . The heat released from this oxidation process is called metabolic heat. It dissipates from the skin surf ace of the human body into the surroundings. In a cold environment, the thermoregulatory mechanism reduces the rate of peripheral blood circulation, lowering the temperature of the skin and pre venting any greater heat loss from the human body. Ho wever, if the heat loss and the mechanical w ork performed are greater than the rate of metabolic heat produced, then the temperatures of both the body core and the skin surf ace fall, and shivering or other spontaneous acti vities occur to increase the production of heat ener gy within the human body. On the other hand, in a hot environment, if a large amount of heat energy needs to be dissipated from the human body , the physiological control mechanism increases the blood fl w to the skin surface. This raises the skin temperature. If the heat produced is still greater than the heat actually dissipated and the temperature of the body core increased from its normal temperature of about 97.6 to about 98.6°F (36.4 to about 37.0°C), then liquid water is released from the sweat glands for evaporative cooling. For a transient state of energy balance between the human body and the indoor environment, the rate of heat storage in the body core Scr and the skin surf ace Ssk, both in Btu /h ft2 (W/m2), can be calculated as Scr  Ssk  M  W  (C  R)  Esk  Eres

(4.3)

4.4

CHAPTER FOUR

4.3 METABOLIC RATE AND SENSIBLE HEAT LOSSES FROM HUMAN BODY Metabolic Rate The metabolic rate M is the rate of ener gy release per unit area of skin surf ace as a result of the oxidative processes in the li ving cells. Metabolic rate depends mainly on the intensity of the physical activities performed by the human body. The unit of metabolic rate is called the met. One met is defined as 18.46 Bt / hft2 (58.24 W/m2) of metabolic heat produced in the body core. In Table 4.1 are listed the metabolic rates of various activities.

Mechanical Work Some of the energy released from the oxidati ve processes within the body core can be partly transformed to e xternal mechanical w ork through the action of the muscles. Mechanical w ork W is usually expressed as a fraction of the metabolic rate and can be calculated as W  M

(4.4)

where   mechanical efficien y. For most office ork, mechanical efficien y is  less than 0.05. Only when there is a lar ge amount of physical acti vity such as bic ycling, lifting and carrying, or walking on a slope may increase  to a value of 0.2 to 0.24. TABLE 4.1 Metabolic Rate for Various Activities Metabolic rate Activity level

Met

Btu/h  ft2

Resting Sleeping Seated, quiet

0.7 1.0

13 18

Office ork Reading, seated Typing

1.0 1.1

18 20

Teaching

1.6

30

1.6 – 2.0 2.0 – 3.4

29 – 37 37 – 63

Walking Speed 2 mph 4 mph

2.0 3.8

37 70

Machine work Light Heavy

2.0 – 2.4 4.0

37 – 44 74

Dancing, social

2.4 – 4.4

44 – 81

Sports Tennis, singles Basketball Wrestling

3.6 – 4.0 5.0 – 7.6 7.0 – 8.7

66 – 74 90 – 140 130 – 160

Domestic work Cooking House cleaning

Source: Adapted with permission from Handbook 1989, Fundamentals.

ASHRAE

INDOOR AND OUTDOOR DESIGN CONDITIONS

4.5

Sensible Heat Exchange Sensible heat loss or, occasionally, sensible heat gain R  C represents the heat e xchange between the human body and the indoor en vironment through con vective and radiati ve heat transfer . Figure 4.1 shows the sensible heat e xchange between the human body and the en vironment. The combined convective and radiative heat transfer can be calculated as C  R  fclh c (Tcl  Ta)  fclh r(Tcl  Trad)  fclh(Tcl  To)

(4.5)

where Tcl  mean surface temperature of clothing, °F (°C). The operative temperature To is define as the weighted average of the mean radiant temperature Trad and indoor air temperature Ta, both in °F (°C), that is, To 

hrTrad  hcTa hr  hc

(4.6)

The surface heat-transfer coefficient is defined h  hc  hr

(4.7)

and the ratio of the clothed surface area to the naked surface area is fcl 

Acl AD

(4.8)

where hc, hr  convective and radiative heat-transfer coefficient Btu/h ft2 °F (W/m2 °C) Acl  surface area of clothed body, ft2 (m2) The mean radiant temperature Trad is discussed in greater detail in later sections. According to Seppenan et al. (1972), the con vective heat-transfer coef ficient hc for a person standing in moving air, when the air velocity is 30  v 300 fpm (0.15  v 1.5 m/s), is hc  0.0681v 0.69

FIGURE 4.1 Sensible heat e xchange between the human body and the indoor environment.

(4.9)

4.6

CHAPTER FOUR

When the air v elocity v 30 fpm (0.15 m /s), hc  0.7 Btu /h ft2 °F (4 W/m2 °C). F or typical indoor temperatures and a clothing emissi vity nearly equal to unity , the linearized radiati ve heattransfer coeff cient hr  0.83 Btu /hft2 °F (4.7 W/m2 °C). Let Rcl be the R value of clothing, in h ft2 °F/Btu (m2 °C/W). Then CR

Tsk  Tcl Rcl

(4.10)

where Tsk  mean skin surface temperature, °F (°C). If the human body is able to maintain a thermal equilibrium with v ery little e vaporative loss from the skin surf ace, then the skin temperature Tsk.n will be around 93°F (33.9°C). Combining Eqs. (4.5) and (4.10), and eliminating Tcl, we f nd CR

fcl h(Tsk  To) R cl fcl h  1

 Fcl fcl h(Tsk  To)

(4.11)

In this equation, the dimensionless clothing eff ciency Fcl is def ned as Fcl 

Tcl  To 1  R cl fcl h  1 Tsk  To

(4.12)

In Eq. (4.11), if To Tsk, then C  R could be negative, i.e., there could be a sensible heat gain.

TABLE 4.2 Insulation Values for Clothing Ensembles*

Ensemble description† Walking shorts, short-sleeve shirt Fitted trousers, short-sleeve shirt Fitted trousers, long-sleeve shirt Same as above, plus suit jacket Loose trousers, long-sleeve shirt, long-sleeve sweater, T-shirt Sweat pants, sweatshirt Knee-length skirt, short-sleeve shirt, pantyhose (no socks), sandals Knee-length skirt, long-sleeve shirt, full slip, pantyhose (no socks) Long-sleeve coveralls, T-shirt Overalls, long-sleeve shirts, long underwear tops and bottoms, f annel long-sleeve shirt

Rcl, clo

Rcl, h  ft2  °F Btu

fcl

0.41 0.50 0.62 0.96

0.36 0.44 0.55 0.85

1.11 1.14 1.19 1.23

1.01 0.77

0.89 0.68

1.28 1.19

0.54

0.48

1.26

0.67 0.72

0.59 0.63

1.29 1.23

1.00

0.88

1.28

*For mean radiant temperature equal to an air temperature and air velocity less than 40 fpm. † Unless otherwise noted, all ensembles included briefs or panties, shoes, and socks. Source: Adapted from McCullough and Jones (1984). Reprinted with permission.

INDOOR AND OUTDOOR DESIGN CONDITIONS

4.7

Clothing Insulation Clothing insulation Rcl can be determined through measurements on a heated manikin, a model of the human body for laboratory e xperiments. After C  R is measured from the thermal manikin in a controlled indoor en vironment, Rcl can be calculated from Eq. (4.11) since fcl, Tsk, To, and h are also known values. Clothing insulation Rcl can be e xpressed either in h ft2 °F/Btu (m 2 °C/W) or in a new unit called clo, where 1 clo  0.88 h ft2 °F/Btu (0.16 m2 °C/W). Clothing insulation Rcl values and area ratios fcl for typical clothing ensembles, taken from McCullough and Jones (1984), are listed in Table 4.2.

4.4 EVAPORATIVE HEAT LOSSES Evaporative heat loss E is heat loss due to the e vaporation of sweat from the skin surf ace Esk and respiration losses Eres. Actually, metabolic heat is mainly dissipated to the indoor air through the evaporation of sweat when the indoor air temperature is nearly equal to the skin temperature.

Respiration Losses During respiration, there is con vective heat loss Cres that results from the temperature of the inhaled air being increased to the e xhaled air temperature, or about 93°F (33.9°C). There is also a latent heat loss Lres due to e vaporation of liquid w ater into w ater v apor inside the body core. The amount of respiration loss depends mainly on the metabolic rate. In summer , at an indoor temperature of 75 °F (23.9°C) and a relati ve humidity of 50 percent, respiratory losses Eres  Cres  Lres are approximately equal to 0.09 M. In winter , Eres is slightly greater . F or simplicity , let Eres  0.1M.

Evaporative Heat Loss from Skin Surface Evaporative heat loss from the skin surface Esk consists of (1) the evaporation of sweat as a result of thermoregulatory mechanisms of the human body Ersw and (2) the direct dif fusion of liquid w ater from the skin surface Edif . Evaporative heat loss due to regulatory sweating Ersw, Btu /h ft2 (W/m2), is directly proportional to the mass of the sweat produced, i.e., E rswm˙ rswh fg

(4.13)

where mass f ow rate of sweat produced, lb/h ft (kg/s m ) m˙ rsw  hfg  latent heat of vaporization at 93°F (33.9°C), Btu/lb (J/kg) 2

2

Maximum Evaporative Heat Loss due to Regulatory Sweating The wetted portion of the human body needed for the e vaporation of a given quantity of sweat wrsw is wrsw 

Ersw Emax

(4.14)

In Eq. (4.14), Emax represents the maximum e vaporative heat loss due to re gulatory sweating when the skin surf ace of the human body is entirely wet. Its magnitude is directly proportional to the vapor pressure difference between the wetted skin surf ace and the indoor ambient air , and it can be

4.8

CHAPTER FOUR

TABLE 4.3 Moisture Permeability of Clothing Ensembles im

Ensemble description Cotton/polyester long-sleeve shirt, long trousers, street shoes, socks, briefs Cotton short-sleeve shirt, long trousers, work boots, socks, briefs, cotton gloves Cotton/nylon long-sleeve shirt, cotton/nylon trousers, combat boots, socks, helmet liner (army battle dress uniform)

0.385 0.41 0.36

*Measured with Trad  Ta, and air velocity  40 fpm. Source: Adapted with permission from ASHRAE Handbook 1989, Fundamentals.

calculated as Emax  he,c(psk, s  pa )

(4.15)

where psk, s  saturated water vapor pressure at skin surface temperature, psia (kPa) pa  water vapor pressure of ambient air, psia (kPa ) he,c  overall evaporative heat-transfer coeff cient of clothed body, in Btu/h ft2psi (W/m2 kPa) Woodcock (1962) proposed the following relationship between he, c and h s, the overall sensible heat transfer coeff cient, in Btu/h ft2 psi (W/m2 kPa): im LR 

he,c hs

(4.16)

The moisture permeability index im denotes the moisture permeability of the clothing and is dimensionless. Clothing ensembles w orn indoors usually ha ve an im  0.3 to 0.5. The moisture permeability indexes im of some clothing ensembles are presented in Table 4.3. The Lewis relation LR in Eq. (4.16) relates the e vaporative heat-transfer coef f cient he and the convective heat-transfer coeff cient hc, both in Btu /hft2 °F(W/m2°C). LR  f(he / hc) has a magnitude of 205 °F/psi (16.5°C/kPa). In Eq. (4.16), the overall sensible heat-transfer coef f cient hs can be calculated as fcl h 1  hs  Rt fcl hR cl  1 (4.17) where Rt  total resistance to sensible heat transfer between the skin and the indoor en vironment, h  ft2 °F/Btu (m2°C/W). Diffusion Evaporative Heat Loss and Total Skin Wetness The minimum le vel of e vaporative heat loss from the skin surf ace occurs when there is no re gulatory sweating and the skin wetness due to direct dif fusion Edf, min, Btu/h ft2 (W/m2), is approximately equal to 0.06Emax under normal conditions, or Edf, min  0.06Emax

(4.18)

When there is a heat loss from re gulatory sweating Ersw, the dif fusion e vaporative heat loss Edif, Btu/h ft2 (W/m2), for the portion of skin surf ace that is not co vered with sweat can be

INDOOR AND OUTDOOR DESIGN CONDITIONS

4.9

calculated as Edif  (1  wrsw)0.06Emax

(4.19)

Therefore, the total evaporative heat loss from the skin surface is Esk  Ersw  Edif  wrsw Emax  (1  wrsw)0.06Emax  (0.06  0.94wrsw)Emax  wskEmax

(4.20)

In Eq. (4.20), wsk is called the total skin wetness; it is dimensionless, and it can be calculated as wsk 

Esk Emax

(4.21)

4.5 MEAN RADIANT TEMPERATURE AND EFFECTIVE TEMPERATURE Mean Radiant Temperature Mean radiant temperature TRad is def ned as the temperature of a uniform black enclosure in which an occupant would have the same amount of radiative heat exchange as in an actual indoor environment. Mean radiant temperature TRad, °R(K), can be calculated by the expression 4 4 T Rad  T R1 F0 – 1  T 4R2 F0 –2      T 4Rn F0 – n

(4.22)

where TR1, TR2,   , TRn  absolute temperature of surrounding surfaces of indoor environment, °R(K) F0 – 1  shape factor denoting fraction of total radiant energy leaving surface of occupant’s clothing 0 and arriving on the surface 1 F0 – 2  fraction of total radiant energy leaving surface 0 and arriving on surface 2, etc. Shape factors F0 – n depend on the position and orientation of the occupant as well as the dimensions of the enclosure. One can use Figs. 4.2 and 4.3 to estimate the mean v alue of the shape f actor between a seated person and rectangular surf aces. The sum of the shape f actors of all the surf aces with respect to the seated occupant in an enclosure is unity. The temperature measured by a globe thermometer , called the globe temperature, is often used to estimate the mean radiant temperature. The globe thermometer consists of a copper hollo w sphere of 6-in. (152-mm) diameter that is coated with black paint on the outer surf ace. A precision thermometer or thermocouple is inserted inside the globe with the sensing b ulb or the thermojunction located at the center of the sphere. Because the net radiant heat recei ved at the globe surface is balanced by the con vective heat transfer from the globe surf ace in reaching a thermal equilibrium, according to Bedford and Warmer (1935), such a relationship gives 4 4 T Rad  T Rg  0.247 109 v 0.5(TRg  TRa)

(4.22a)

where TRad  absolute mean radiant temperature, °R (K) TRg  absolute globe temperature, °R (K) v  ambient air velocity, fpm (m/s) TRa  absolute air temperature, °R (K) After TRg, v, and TRa are measured, the mean radiant temperature TRad can be calculated from Eq. (4.22). The mean radiant temperature indicates the ef fect, due to the radiant ener gy from the

4.10

CHAPTER FOUR

FIGURE 4.2 Mean value of shape f actor between a sedentary person and a horizontal plane. ( Source: P. O. Fanger, Thermal Comfort Analysis and Applications in Environmental Engineering, 1972. Reprinted with permission.)

INDOOR AND OUTDOOR DESIGN CONDITIONS

4.11

FIGURE 4.3 Mean value of shape f actor between a seated person and a v ertical plane. ( Source: P. O. Fanger, Thermal Comfort Analysis and Applications in Environmental Engineering, 1970. Reprinted with permission.)

4.12

CHAPTER FOUR

FIGURE 4.4 Dimension, in feet, of a private off ce.

surroundings, on radiant e xchange between an occupant or an y substance and the enclosure. Such an inf uence may be signi f cant if the mean radiant temperature is se veral degrees higher than the temperature of the indoor air. Example 5.1. The dimensions of a pri vate off ce and the location of a person seated within it are shown in Fig. 4.4. The surface temperatures of the enclosure are as follows: West window West wall North partition wall East partition wall South partition wall Floor 78 Ceiling

88°F (31.1°C) 80°F (26.7°C) 75°F (23.9°C) 75°F (23.9°C) 75°F (23.9°C) °F (25.6°C) 77°F (25°C)

Calculate the mean radiant temperature of the enclosure that surrounds this of f ce. The orientation of the seated occupant is unknown. Solution 1. Regarding the north partition wall, the shape factor denotes the fraction of the total radiant energy that leaves the outer surf ace of the clothing of the occupant (surf ace 0) and arrives directly on the north partition wall (surfaces 1, 2, 3, and 4) and is given by F0 – 1,2,3,4  F0 – 1  F0 – 2  F0 – 3  F0 – 4 Here F0 – 1 is the shape factor for the radiation from surface 0 to surface 1 of the north partition wall. Based on the curves in Fig. 4.3, for a ratio of b/L  1.8/3  0.6 and a ratio of a/L  4.5/3  1.5,

INDOOR AND OUTDOOR DESIGN CONDITIONS

4.13

4 TABLE 4.4 Values of F0 – n and TRn F0 – n in Example 4.1

Surface

Surface Shape temperature, °R

factor

b/L

a/L

F0– n

4 T Rn F0– n 108

North partition wall

535

F0 – 1 F0 – 2 F0 – 3 F0 – 4

0.6 0.6 2.2 2.2

1.5 3.5 1.5 3.5

0.04 0.045 0.07 0.087

32.77 36.87 57.35 71.27

East partition wall

535

F0 – 17 F0 – 18 F0 – 19 F0 – 20

0.63 0.63 0.17 0.17

0.29 0.86 0.29 0.86

0.014 0.03 0.004 0.009

11.47 24.58 3.28 7.37

South partition wall

535

F0 – 9 F0 – 10 F0 – 11 F0 – 12

0.73 0.73 0.2 0.2

0.5 1.17 0.5 1.17

0.023 0.038 0.008 0.013

18.84 31.13 6.55 10.65

West wall

540 540

F0 – 13 F0 – 14

0.4 0.4

0.67 2

0.018 0.03

15.31 25.51

West window

548 548

F0 – 15 F0 – 16

1.6 1.6

0.67 2

0.04 0.07

36.07 63.13

Floor

538

F0 – 5 F0 – 6 F0 – 7 F0 – 8

1.67 1.67 5 5

2.5 5.8 2.5 5.8

0.068 0.073 0.087 0.102

56.97 61.16 72.89 85.45

Ceiling

537

F0 – 21 F0 – 22 F0 – 23 F0 – 24

1.36 1.36 0.45 0.45

0.68 1.59 0.68 1.59

0.033 0.052 0.015 0.025

27.44 43.24 12.47 20.79

0.994

8.32.56 108

the shape factor F0 –1  0.04. Here L is the horizontal distance from the occupant to the north partition wall. The shape factors F0 – 2, F0 – 3, and F0 – 4 can be calculated in the same manner. 2. To determine the shape factor F0 – 5 for the radiation from the occupant (surf ace 0) to the f oor (surface 5), we note that the vertical distance L from the center of the seated occupant to the f oor is 1.8 ft. From the curv es in Fig. 4.2, the ratio b/L  3/1.8  1.67, and the ratio a/L  4.5/1.8  2.5; thus the shape f actor F0 – 5  0.068. All the remaining shape f actors can be determined in the same manner as listed in Table 4.4. 3. For the north partition wall (surface 1), T 4R1 F0 – 1  (75  460)4 0.04  32.77 108 4 Other products T Rn F0 – n can be similarly calculated, as listed in Table 4.4. The sum of the products 4

T Rn F0– n  832.56 108. Therefore

T 4Rad  832.56 108 That is, TRad  537.2°R

or

Trad  77.2°F(25.1°C)

4. The sum of the shape factors F0 – n  0.994 is nearly equal to 1.

4.14

CHAPTER FOUR

Effective Temperature The effective temperature ET* is the temperature of an en vironment that causes the same total heat loss from the skin surface as in an actual environment of an operative temperature equal to ET* and at a relative humidity of 50 percent. And ET* can be calculated as ET*  To  wsk i m LR (0.5pET, s)

(4.23)

where pET, s  saturated w ater v apor pressure at ET*, psia (kP a abs.). The right-hand side of Eq. (4.23) describes the conditions of the indoor air re garding the total heat loss from the human body. The same value of the combination To  wsk i m LR (0.5 pET, s ) results in the same amount of total heat loss from the skin surf ace, if other parameters remain the same. Theoretically, total skin wetness wsk and clothing permeability index im are constants for a specif c ET* line. Because the effective temperature is based on the operati ve temperature To, it is a combined index of Ta, Trad, and pa. In an indoor air temperature belo w 77°F (25°C), the constant-ET* lines are nearly parallel to the skin temperature lines for sedentary occupants with a clothing insulation of 0.6 clo; therefore, ET* values are reliable inde xes to indicate thermal sensations at normal indoor air temperature during lo w activity levels. The term effective temperature was originally proposed by Houghton and Yaglou in 1923. A new def nition of ET* and its mathematical e xpression were developed by Gagge et al. in 1971. It is the en vironmental index commonly used in specifying and assessing thermal comfort requirements.

4.6 FACTORS AFFECTING THERMAL COMFORT Daily experience and many laboratory experiments all show that thermal comfort occurs only under these conditions: 1. There is a steady-state thermal equilibrium between the human body and the en vironment; i.e., heat storage of the body core Scr and the skin surface Ssk are both equal to zero. 2. Regulatory sweating is maintained at a low level. From the heat balance equation at steady-state thermal equilibrium Eq. (4.1) we have M  W  C  R  Esk  Eres Let Eres  0.1M and the mechanical ef f ciency   0.05M. From Eq. (4.11), C  R can be determined. Also, from Eqs. (4.15), (4.16), and (4.20), Esk is a kno wn v alue. If we substitute into Eq. (4.1), the heat balance equation at steady-state thermal equilibrium can be expressed as M(1  0.05  0.1)  Fcl fcl h(Tsk  To)  wsk i m LR hs (psk, s  pa) or 0.85M  Fcl fcl h(Tsk  To)  wsk im LR hs ( psk, s  pa )

(4.24)

In Eq. (4.24), the physiological and en vironmental factors that af fect the balance — the metabolic rate and the heat losses on the two sides of the equation — are as follows: 1. Metabolic rate M determines the magnitude of the heat ener gy that must be released from the human body, i.e., the left-hand side of the equation. 2. Indoor air temperature Ta is a weighted component of the operating temperature To. It also affects the sensible heat loss and the v apor pressure of indoor air pa in the calculation of the evaporative loss from the skin surface.

INDOOR AND OUTDOOR DESIGN CONDITIONS

4.15

3. Mean radiant temperature Trad is another weighted component of the operating temperature To. It affects the sensible heat loss from the human body. 4. Relative humidity of the ambient air a is the dominating f actor that determines the dif ference psk, s  pa in the evaporative loss from the skin surface. Air relative humidity becomes important when the e vaporative heat loss due to re gulatory sweating is the dominating heat loss from the human body. 5. Air v elocity va inf uences the heat transfer coef f cient h and the clothing ef f ciency Fcl in the term in Eq. (4.24) for the sensible heat loss from the human body . It also affects the overall sensible heat transfer coef f cient hs in the e vaporative heat loss term and the clothing permeability im term in Eq. (4.24). 6. Clothing insulation Rcl affects the clothing eff ciency Fcl, the area ratio fcl, the heat transfer coeff cient h, the clothing permeability inde x im, and the overall sensible heat-transfer coeff cient hs.

4.7 THERMAL COMFORT Thermal comfort is def ned as the state of mind in which one acknowledges satisfaction with regard to the thermal environment. In terms of sensations, thermal comfort is described as a thermal sensation of being neither too warm nor too cold, def ned by the following seven-point thermal sensation scale proposed by ASHRAE: 3  cold 2  cool 1  slightly cool 0  neutral 1  slightly warm 2  warm 3  hot Fanger’s Comfort Equation A steady-state energy balance is a necessary condition for thermal comfort, but is not suf f cient by itself to establish thermal comfort. F anger (1970) calculated the heat losses for a comfortable person, experiencing a neutral sensation, with corresponding skin temperature Tsk and re gulatory sweating Ersw. The calculated heat losses L are then compared with the metabolic rate M. If L  M, the occupant feels comfortable. If L M, then this person feels cool; and if L M, then this person feels warm. Using the responses of 1396 persons during laboratory e xperiments at Kansas State Uni versity of the United States and Technical University of Denmark, Fanger developed the follo wing equation to calculate the predicted mean vote (PMV) in the seven-point thermal sensation scale: PMV  (0.303e 0.036M  0.276)(M  L)

(4.25)

In Eq. (4.25), the metabolic rate M and heat losses L are both in Btu /h ft2 (W/m2). According to Fanger’s analysis, the predicted per centage of dissatisfie (PPD) v ote for thermal comfort at a PMV  0 is 5 percent, and at a PMV  1 is about 27 percent. Tables of PMV and comfort charts including v arious combinations of operating temperature To, air velocity v, metabolic rate M, and clothing insulation Rcl have been prepared to determine comfortable conditions conveniently. Fanger’s comfort charts also include relati ve humidity. Six of his

FIGURE 4.5 Fanger’s comfort charts. (Abridged with permission from ASHRAE Handbook 1981, Fundamentals.)

4.16

INDOOR AND OUTDOOR DESIGN CONDITIONS

4.17

comfort charts at v arious activity levels, wet-bulb temperatures, relative humidities, and air velocities are shown in Fig. 4.5. In the f rst four charts, the air temperatures are equal to the mean radiant temperatures. Three of these four ha ve a clothing insulation of 0.5 clo. The other is for 1-met activity level and 1 clo, because one rarely f nds an occupant with such heavy outwear at an activity level of 2 or 3 met. In the f fth and sixth charts, the air temperature could be dif ferent from the mean radiant temperature with a constant relati ve humidity of 50 percent. These comfort variations clearly show that all six f actors — air temperature T a , mean radiant temperature Trad, relative humidity , air velocity v, metabolic rate M, and clothing insulation Rcl —seriously affect the thermal comfort. For example, from Fanger’s comfort chart, a sedentary occupant at an acti vity level of 1.0 met, with a clothing insulation of 0.5 clo, in an air conditioned space at a relative humidity of 50 percent and an air v elocity less than 20 fpm (0.1 m /s), feels comfortable with an air temperature equal to the mean radiant temperature of 78°F (25.6°C). If all values were identical except for a 2-met activity level, the temperature would need to be 67°F (19.4°C) for the same level of comfort. Another factor, the duration of exposure to the indoor thermal environment, should be discussed here. If an indoor en vironment can pro vide thermal comfort for the occupant, the duration of the exposure has no signi f cant inf uence upon the physiological responses of the person ’s thermal regulatory mechanism. If the indoor en vironment is uncomfortable, subjecting the subject to a certain degree of heat or cold stress, the time e xposure will in f uence the person ’s physiological response.

ASHRAE Comfort Zones Based on results of research conducted at Kansas State Uni versity and at other institutions, ANSI/ASHRAE Standard 55-1992 speci f ed winter and summer comfort zones to pro vide for the selection of the indoor parameters for thermal comfort (see Fig. 4.6). This chart is based upon an occupant activity level of 1.2 met (69.8 W/m2). For summer, typical clothing insulation is 0.5 clo, that is, light slacks and short-slee ve shirt or comparable ensemble; there is no minimum air speed that is necessary for thermal comfort. Standard 55-1992 recommended a summer comfort zone with an effective boundary temperature ET*  73 to 79°F (22.5 to 26°C) at 68°F (20°C) wet-bulb as its upper-slanting boundary and dew-point temperature 36°F (2.2°C) as its bottom f at boundary. If the clothing insulation is 0.1 clo higher , the boundary temperatures both should be shifted 1 °F (0.6°C) lower. Rohles et al. (1974) and Spain (1986) suggested that the upper boundary of the summer comfort zone can be e xtended to 85 or 86 °F (29.4 or 30°C) ET* if the air v elocity of the indoor air can be increased to 200 fpm (1 m /s) by a ceiling fan or other means. The winter comfort zone is based upon a 0.9-clo insulation including hea vy slacks, long-sleeve shirt, and sweater or jacket at an air velocity of less than 30 fpm (0.15 m /s). Standard 55-1992 recommended a winter comfort zone with an ef fective boundary temperature ET *  68 to 74°F (20 to 23.3°C) at 64°F (17.8°C) wet-bulb as its slanting upper boundary and at de w-point 36°F (2.2°C) as its bottom f at boundary. Indoor air parameters should be f airly uniform in order to a void local discomfort. According to Holzle et al., 75 to 89 percent of the subjects tested found the en vironment within this summer comfort zone to be thermally acceptable. ASHRAE comfort zones recommend only the optimal and boundary ET* for the determination of the winter and summer indoor parameters. F or clothing insulation, activity levels, and indoor air velocities close to the v alues specif ed in Standard 55-1992, a wide range of indoor design conditions are available.

Comfort-Discomfort Diagrams A comfort diagram pro vides a graphical presentation of the total heat loss from the human body at various operative or air temperatures and indoor relati ve humidities when the activity level, level of clothing insulation, and air velocity are specif ed. The abscissa of the comfort diagram is the opera-

CHAPTER FOUR

%

70

70 %

10 0%

w, lb/lb

60

0.015

50

%

Summer

Winter

60 Dew-point temperature Tdew, F

68 64

F

we

tb

ulb

F

we

tb

0.010

ulb

% 30

50

40

0.005

20

79

74

73

30 ET* 68 F

4.18

10 0 60

70

80

90

Operative temperature T0, F FIGURE 4.6 ASHRAE comfort zones. (Adapted with permission from ANSI/ASHRAE Standard, 55 – 1992.)

tive temperature To or ambient air temperature Ta, whereas the ordinate can be either w ater vapor pressure p or humidity ratio w. Figure 4.7 shows a comfort diagram with a sedentary acti vity level, a clothing insulation v alue of 0.6 clo, and still-air conditions, i.e., an air v elocity v 20 fpm (0.1 m/s). The curv ed lines represent relati ve humidity, and the straight lines represent ef fective temperature ET*. The short dash curv es di verging from the ET* lines are total skin wetness wsk lines. The f gure is based upon To  Ta.

INDOOR AND OUTDOOR DESIGN CONDITIONS

4.19

FIGURE 4.7 Comfort-discomfort diagram.

Effective temperature ET* lines are calculated according to Eq. (4.23). Because the total skin wetness wsk is a constant for a speci f c ET* below 79°F ET*, the ET* lines and wsk lines coincide with one other. At higher ET* values, wsk lines curve to the left at high relati ve humidities. For low ambient air temperatures, evaporative heat loss from the skin surf ace Esk is small; therefore, ET* and wsk lines are nearly v ertical. As Esk becomes greater and greater, the slopes of the ET* and wsk lines decrease accordingly. The comfort diagram is divided into f ve zones by the ET* and wsk lines: 1. Body cooling zone . For the condition gi ven in Fig. 4.7, if the ef fective temperature ET* 73°F (22.8°C), the occupant will feel cold in this zone. Because the heat losses exceed the net metabolic rate, the skin and body core temperatures tend to drop gradually. 2. Comfort zone. This is the zone between the lower boundary ET*  73°F (22.8°C), and wsk  0.06, and the higher boundary ET*  86°F (30°C), and wsk  0.25. Steady-state thermal equilibrium is maintained between the occupant and the en vironment, and re gulatory sweating is at a rather low level. The occupant will feel comfortable in this zone, and the heart rate (HR) is between 76 and 87 beats per minute. ASHRAE’s winter and summer comfort zones are a part of this zone. The lower boundary of the ASHRAE winter comfort zone forms the lo wer boundary of this comfort zone. The reason that the

4.20

CHAPTER FOUR

lower boundary in this diagram is ET*  73°F (22.8 °C) whereas ET*  68°F (20 °C) in ASHRAE’s winter comfort zone is that a lower clothing insulation of 0.6 clo is used here. 3. Uncomfortable zone. In this zone, 86°F ET*  95°F (30 °C  ET*  35°C) and 0.25 wsk  0.45. Thermal equilibrium also e xists between the occupant and the en vironment, and the evaporative heat loss due to re gulatory sweating dominates. Heart rate sho ws a range between 87 and 100. The occupant feels uncomfortable, i.e., warm or hot, when his or her physiological parameters are in these ranges. 4. Very uncomfortable zone. In this zone, 95°F ET* 106.5°F (35°C ET* 41.4°C) and 0.45 wsk 1. Although thermal equilibrium is still maintained with zero heat storage at the skin and the body core, there is a danger of a heat strok e when ET* 95°F (35°C). Toward the upper boundary of this zone, the skin surface is nearly entirely wet, and the heart rate exceeds 120. Under these conditions, the occupant will feel very hot and very uncomfortable. 5. Body heating zone. When ET*  106.5°F (41.4°C) and wsk  1, thermal regulation by evaporation f ails. At a higher ET* or wsk, the environment is intolerable, and the temperatures of the body core and skin tend to rise gradually. For an air conditioned space with an occupant at lo w acti vity le vels ( M 2 met), the indoor environment is usually maintained within the comfort zone, and the physiological and thermal responses of the occupant are also in the comfort zone. Only at higher acti vity levels do the thermal responses occasionally fall into the discomfort zone.

4.8 INDOOR AIR TEMPERATURE AND AIR MOVEMENTS Comfort Air Conditioning Systems For comfort air conditioning systems, most occupants have a metabolic rate of 1.0 to 1.5 met. The indoor clothing insulation in summer is usually 0.35 to 0.6 clo, and in winter it is 0.8 to 1.2 clo. Relative humidity has a lesser in f uence on thermal comfort, and will be discussed in the ne xt section, but indoor air temperature and air velocity are discussed here. Many researchers have conducted tests to determine the ef fects of airspeed on the preferred indoor air temperature and the thermal comfort of occupants. The relationship between the preferred indoor air temperatures and v arious airspeeds is presented in Fig. 4.8. Most of the data were tak en under these conditions: metabolic rate M  400 Btu /h (117 W), clothing insulation Rcl  0.6 clo, Trad  Ta, and relative humidity of the indoor air   50 percent. The one exception is the students in Holzle ’s e xperiments, who had 0.54 clo for summer and 0.95 clo for winter . Examination of Fig. 4.8 shows the following: ●

●

Higher indoor air temperature requires greater indoor air velocity to provide thermal comfort. Variation of airspeed has a greater in f uence on preferred indoor air temperature at lo wer air temperatures.

ANSI/ASHRAE Standard 55-1992 recommended that within the thermally acceptable temperature ranges in the ASHRAE summer and winter comfort zones discussed in Sec. 4.7, there be no minimum airspeed (nondirectional) that is necessary for thermal comfort. If temperature is increased abo ve the le vel allowed for the comfort zone, means must be provided to ele vate the airspeed. F or instance, when indoor air temperature Ta  Trad (mean radiant temperature), given that the airspeed for a summer comfort zone of 79 °F (26 °C) is 40 fpm (0.2 m/s), if Ta has an increase of 2 °F (1.1 °C) from 79 to 81 °F (26 to 27.2 °C), according to ANSI/ASHRAE Standard 55-1992, there must be a rele vant increase of airspeed of about 70 fpm (0.35 m/s). Before specifying Ta for summer conditions, one needs to determine whether occupants are likely to wear suit jack ets, such as members of a church congre gation or guests in a multipurpose

INDOOR AND OUTDOOR DESIGN CONDITIONS

4.21

FIGURE 4.8 Preferred indoor air temperatures at various air velocities.

hall. In such cases, a reduction of 4 °F of summer optimal ET * may be necessary because of the increase in clothing insulation of about 0.4 clo.

Design Considerations When one is specifying indoor design conditions, thermal comfort must be pro vided at optimum cost while using energy eff ciently. These principles should be considered: 1. To determine the optimum summer and winter indoor design temperatures, consider the local clothing habits and the upper and lo wer acceptable limits on clothing insulation at v arious operative temperatures To, as shown in Fig. 4.9. 2. It is al ways more ener gy-eff cient to use dif ferent indoor design temperatures for summer and winter than a year -round constant v alue. An unoccupied-period setback during winter al ways saves energy. There are also buildings in which a constant indoor temperature is required for the health and comfort of the occupants, such as in many health care facilities. 3. For short-term occupancies, or when the metabolic rate is higher than 1.2 met, a strategy of using a lower energy-use ceiling f an or a w all-mounted fan to pro vide higher air v elocity may be considered. Thus a higher indoor design temperature within the e xtended summer comfort zone may be acceptable for occupants, especially in industrial settings.

Indoor Design Temperatures for Comfort Air Conditioning According to ANSI/ASHRAE Standard 55-1992, Thermal Environmental Conditions for Human Occupancy, and ASHRAE /IES 90.1-1999, Energy-Standard for Buildings Except Low-Rise

4.22

CHAPTER FOUR

FIGURE 4.9 Relationship between clothing insulation Rcl and operating temperature To. (Source: ASHRAE Transactions 1983, Part I B. Reptinted with permission.)

Residential Buildings, the following indoor design temperature and air speed apply for comfort air conditioning systems when the activity level is 1.2 met, there is a relative humidity of 50 percent in summer, mean airspeed  30 fpm (0.15 m/s), and Ta  Trad:

Winter Summer

Typical clothing insulation, clo

Optimum operative temperature

Indoor design temperature range

0.9 0.5

71°F (22°C) 76°F (24.5°C)

69 – 74°F (20.5 – 23.5°C) 74 – 79°F (23.5 – 26°C)

According to ANSI/ASHRAE Standard 55-1992, “within the thermally acceptable temperature ranges, there is no minimum air speed that is necessary for thermal comfort. ” If the summer indoor temperature is 79°F (26°C), an airspeed of 40 fpm (0.2 m /s) is recommended. If Ta 79°F (26°C), a rele vant increase of airspeed in the indoor occupied zone should be considered. Refer to ANSI/ASHRAE Standard 55-1992. If the space relati ve humidity can be lo wered to 35 to 40 percent in summer , then the higher limit 78°F (25.5°C) is often specif ed. Indoor badminton and table tennis tournament arenas should ha ve air v elocities below 30 fpm (0.15 m/s). To avoid nonuniformity and to pre vent local discomfort, the air temperature dif ference between 4 in. from the f oor and 67 in. abo ve the f oor should not e xceed 5°F (3°C). The radiant temperature asymmetry in the vertical direction should be less than 9°F (5°C), and in the horizontal direction less than 18°F (10°C). To determine whether the speci f ed air temperature and airspeed are met, they should be measured at 4-, 24-, and 43-in. (0.1-, 0.6-, and 1.1-m) le vels for sedentary occupants, and 4-, 43-, and 67-in. (0.1-, 1.1-, and 1.7-m) le vels for standing acti vity. The duration for determining the mean value of the air movement should be 3 min. or 30 times the 90 percent response time of the measuring instrument, whichever is greater.

INDOOR AND OUTDOOR DESIGN CONDITIONS

4.23

Process Air Conditioning Systems For process air conditioning systems, indoor design temperature is usually based on pre vious experiences. For precision manuf acturing projects, a basic temperature plus a tolerance, such as 72  2°F (22.2  1.1°C) for a precision machinery assembling w orkshop, is often speci f ed. First, the tolerance should be neither too tight nor too loose. Second, either the temperature f uctuation at various times within the w orking period or the temperature v ariation within the w orking space, or both, should be included in this tolerance. In unidirectional- f ow clean rooms, the v elocity of the airstream in the w orking area is often specif ed as 90  20 fpm (0.45  0.1 m/s) to prevent contamination of the products.

4.9 HUMIDITY Comfort Air Conditioning Systems According to ANSI/ASHRAE Standard 55-1992, for the zone occupied by people engaged in light, primarily sedentary acti vity (  1.2 met), the relative humidity should conform with the limits of ASHRAE winter and summer comfort zones, as sho wn in Fig. 4.6. These limits are intended to maintain acceptable thermal conditions for the occupants’ comfort. ASHRAE/IESNA Standard 90.1-1999 mandates that where a zone is serv ed by system(s) with both humidif cation or dehumidi f cation capacity, means shall be pro vided to prevent simultaneous operation of humidif cation and dehumidif cation equipment. Standard 90.1-1999 also specif es that where humidistatic controls are pro vided, such controls shall be capable of pre venting reheating, mixing of hot and cold air streams, and simultaneous heating and cooling. Refer to Section 29.12 for more details. The following results should be considered during the design and e valuation of the performance of comfort air conditioning systems: ●

●

●

●

●

●

Maintaining an indoor space relative humidity r between 20 and 30 percent in winter prevents or reduces the condensation at the inner side of the window glass. In high-occupancy applications, it may be economical and is still comfortable to specify the summer indoor relati ve humidity at 55 to 60 percent if the indoor temperature is within the summer comfort zone. The indoor relative humidity at part load may be considerably higher than at full load in some air conditioning systems in summer. When the indoor relative humidity is below 25 percent, the incidence of respiratory infections increases signif cantly. If, simultaneously, indoor temperatures are low, such as below 70°F (21°C), the induced static electricity in carpeted rooms may cause uncomfortable shocks to occupants contacting metal furniture or decorations. Because the increase in the outdoor v entilation rates in ASHRAE Standard 62-1999 for the air conditioning system serves the building located in areas where the humidity ratio of outdoor air is very low during winter , the system has greater dif f culty maintaining an indoor relati ve humidty of 20 to 30 percent without winter humidif cation. If a humidi f er is installed, its humidifying capacity should not e xceed the actual humidifying requirements so that wet surf aces do not occur inside the air -handling unit, packaged unit, and supply ducts. Wet surface and dirt cause the gro wth of microorganisms and poor indoor air quality. Therefore, for comfort air conditioning systems, the recommended indoor relati ve humidity

4.24

CHAPTER FOUR

levels are as follows: Relative humidity, % Summer 30 Winter Commercial and public buildings Health care buildings

– 65 20 – 60 30 – 60

Process Air Conditioning Systems Humidity af fects the physical properties of man processes.

y materials and, therefore, their manuf acturing

Moisture Content. Relative humidity has a mark ed in f uence on the moisture content of hygroscopic materials such as natural textile f bers, paper, wood, leather, and foodstuffs. Moisture content affects the weight of the products and sometimes their strength, appearance, and quality. Dimensional Variation. Hygroscopic materials often extend at higher relative humidity and contract at lo wer humidity. A 2 percent increase in moisture content may result in a 0.2 percent increase in dimension of paper. That is why lithographic printing requires a relative humidity of 45  2 percent. Corrosion and Rust. Corrosion is an electrochemical process. Moisture encourages the formation of electrolytes and therefore the corrosion process. A relative humidity greater than 50 percent may affect the smooth operation of bearings in precision instruments. When indoor relati ve humidity exceeds 70 percent, rust may be visible on the surf ace of the machinery and on parts made of steel and iron. Static Electricity . Static electricity may cause minute particles to repel or attract one another , which is detrimental to man y manufacturing processes. Static electricity char ges minute dust particles, in the air, causing them to cling to equipment and w ork surfaces. Static electricity exists in an indoor environment at normal air temperatures when relative humidity is less than about 40 percent. Loss of Water. Vegetables and fruits lose w ater v apor through e vaporation from their surf aces during storage. Low temperatures and high relati ve humidities, such as   90 to 98 percent, may reduce water loss and delay desiccation. It is important to specify the e xact relati ve humidity required for product quality and cost control. For process air conditioning systems, the specif ed relative humidity is either a year -round single value or a range. A strict relati ve-humidity requirement al ways includes a basic v alue and a tolerance, such as the relati ve humidity for lithographic printing mentioned before. When temperature and relati ve humidity controls are both required, they should be speci f ed as a combination. Consider this example:

Clean room

Case Study 4.1.

Temperature, °F (°C)

Relative humidity, %

72  2 (22.2  1.1)

45  5

A factory workshop has the following environmental parameters during summer:

Indoor air temperature Indoor air relative humidity

79°F (26.1°C) 50 percent

INDOOR AND OUTDOOR DESIGN CONDITIONS

Space air velocity Clothing insulation of the workers Activity level

4.25

20 fpm (0.1 m/s) 0.5 clo 3 met

As a result of comfort complaints by personnnel, you are ask ed to recommend ef fective and economical corrective measures to impro ve the indoor en vironment. The following table includes the information required during analysis: Area ratio of the clothed body fcl Permeability index of clothing im Skin surface temperature,Tsk Saturation water vapor pressure at 92.7°F at 79°F

1.2 0.4 92.7°F (33.7°C) 0.764 psia (5.27 kPa abs.) 0.491 psia (3.39 kPa abs.)

Solution 1. When the space air velocity v  20 fpm, the convective heat-transfer coeff cient can be calculated, from Eq. (4.9), as hc  0.0681v 0.69  0.0681 20 0.69  0.538 Btu/h ft2 °F Because at normal indoor conditions the radiati ve heat-transfer coeff cient hr  0.83 Btu /hft2 °F, the surface heat-transfer coeff cient h is h  hc  hr  0.538  0.83  1.368 Btu/h ft2 °F The clothing insulation Rcl  0.5 0.88  0.44 h ft2 °F/Btu. From Eq. (4.12), the clothing ef f ciency is found to be Fcl 

1 1   0.5806 R cl fcl h  1 0.44 1.2 1.368  1

Therefore, from Eq. (4.11), we f nd the sensible heat loss from the skin surface of the worker C  R  Fcl fcl h(Tsk  To)  0.5806 1.2 1.368(92.7  79)  13.06 Btu/h ft2 From Eq. (4.17), the overall sensible heat-transfer coeff cient hs is hs  

fclh fclhR cl  1 1.2 1.368  0.9531 Btu / h ft 2 F 1.2 1.368 0.44  1

From Eq. (4.16), the overall evaporative heat-transfer coeff cient he,c is he,c  hs im LR  0.9531 0.4 205  78.16 Btu/h ft2 psi The maximum e vaporative heat loss Eq. (4.15) as

Emax due to re gulatory sweating can be calculated from

Emax  he,c(psk, s  pa)  78.16(0.764  0.5 0.491)  40.24 Btu/h ft2

4.26

CHAPTER FOUR

In Eq. (4.24), 0.85M  0.85 3 18.46  47.07 Btu/h ft2 When   0.05 and Eres  0.1M, the total e vaporative loss from the skin surf ace of the w orker, from Eq. (4.1), is Esk  0.85M  (C  R)  47.07  13.06  34.01 Btu/h ft2 (107.3 W/m2) Therefore, the total skin wetness is wsk 

Esk 34.01   0.845 Emax 40.24

In Fig. 4.7, when wsk  0.845, a person is very uncomfortable. 2. From Fanger’s comfort chart, shown in Fig. 4.5, for a person with an acti vity level of 3 met and a clothing insulation of 0.5 clo, at a relati ve humidity of 50 percent and an air v elocity of 20 fpm and with Ta  Trad for neutral thermal sensation, the indoor air temperature should be 56 °F (13.3°C). Obviously, this is not economical because too much refrigeration is required. 3. Let us analyze the results if ceiling fans or wall fans are used to increase the space air velocity v to 300 fpm (1.5 m/s). Then the convective heat-transfer coeff cient hc is hc  0.681 3000.69  3.486 Btu/h ft2 °F The surface heat-transfer coeff cient is h  3.486  0.83  4.316 Btu/h ft2 °F Also the clothing eff ciency is calculated as Fcl 

1  0.305 0.44 1.2 4.316  1

Then the sensible heat loss is equal to C  R  0.305 1.2 4.316(92.7  79)  21.64 Btu/h ft2 The overall sensible heat-transfer coeff cient is hs 

1.2 4.316  1.58 Btu / hft 2 F 1.2 4.136 0.44  1

The overall evaporative heat-transfer coeff cient he,c can be shown to be he,c  1.58 0.4 205  129.5 Btu/h ft2 psi Then the maximum evaporative heat loss due to regulatory sweating is Emax  129.5(0.7604  0.5 0.491)  66.68 Btu/hft2 (210.3 W/m2) The total skin wetness is wsk  47.07  21.64/66.68  0.381 That is, wsk has been greatly reduced compared with the v alue at v  20 fpm (0.1 m /s). In Fig. 4.7, wsk  0.381 is in the uncomfortable zone. Workers will feel w arm, but the indoor environment has been considerably improved. This may be the most cost-effective solution.

INDOOR AND OUTDOOR DESIGN CONDITIONS

4.27

4.10 SICK BUILDING SYNDROME AND INDOOR AIR QUALITY Sick building syndrome is a kind of b uilding-related illness that has recei ved public attention since the 1970s. ASHRAE (1987) de f ned the sick b uilding as “. . . a building in which a signi f cant number (more than 20 percent) of b uilding occupants report illness percei ved as being b uilding related. This phenomenon, also kno wn as ‘sick b uilding syndrome ’ is characterized by a range of symptoms including, but not limited to, eye, nose, and throat irritation, dryness of mucous membranes and skin, nose bleeds, skin rash, mental f atigue, headache, cough, hoarseness, wheezing, nausea, and dizziness. Within a given building, there usually will be some commonality among the symptoms manifested as well as temporal association between occupanc y in the b uilding and appearance of symptoms.” If there are signs of actual illnesses, these illness are classif ed as buildingrelated illnesses. Poor indoor air quality (IAQ) is the dominant f actor that causes sick building syndrome. Indoor air quality is de f ned as an indication of harmful concentrations of the indoor air contaminants that affect the health of the occupants or the de gree of satisfaction of a substantial majority (80 percent or more) of occupants e xposed to such an indoor en vironment. Poor control of the indoor air temperature and relati ve humidity are causes of discomfort symptoms. They may also increase the indoor air contaminants. Ho wever, unsatisfactory indoor temperature and indoor relati ve humidity are only indirect causes of poor indoor air quality. National Institute for Occupational Safety and Health (NIOSH) of the United States (1989), according to the results of 529 b uilding in vestigations between 1971 and 1988, and Health and Welfare Canada (HWC), according to the results of 1362 b uilding in vestigations between 1984 and 1989, classif ed the reasons for sick building syndrome as follows: NIOSH, 529 Buildings No. of buildings Inadequate ventilation Indoor contaminants Outdoor contaminants Biological contaminant Building fabric contamination Unknown sources

280 80 53 27 21 68

Percent 53 15 10 5 4 13

HWC, 1362 Buildings No. of buildings 710 165 125 6 27 329

Percent 52 12 9 0.4 2 24

Inadequate ventilation includes lack of outdoor air , poor air distribution, poor thermal control, and inadequate maintenance; and it is the primary cause of indoor air quality. The survey found 70 to 80 percent of the investigated buildings had no known problems. Effective operation and control of the HVAC&R system will be discussed in later chapters. In the United States, most people spend about 90 percent of their time indoors. The purpose of specifying the indoor design conditions in the design documents is to pro vide the occupants with a satisfactory indoor environment at optimum cost. After the energy crisis in 1973, a lower outdoor ventilation rate, a tighter building shell, and the use of variable-air-volume (VAV) systems at part-load operation may reduce the amount of outdoor air intake signif cantly. Indoor air quality therefore has become one of the critical HV AC&R problems especially in commercial buildings since the 1980s. Indoor Air Contaminants Based on the results of the f eld investigations of three off ce buildings by Bayer and Black in 1988, the indoor air contaminants that relate to indoor air quality and the symptoms of the sick b uilding syndrome are mainly the following:

4.28

CHAPTER FOUR

1. Total particulate concentration. This parameter includes particulates from b uilding materials, combustion products, and mineral and synthetic f bers. In February 1989, the U.S. En vironmental Protection Agency (EP A) speci f ed the allo wable indoor concentration le vel of particulates of 10 m and less in diameter (which can penetrate deeply into the lungs, becoming hazardous to health) as follows: 50 g/m3 (0.000022 gr/ft3): 150 g/m (0.000066 gr/ft ): 3

3

1 year 24 h

According to ASHRAE Handbook 1997, Fundamentals, particles less than 2 m in diameter are most likely retained in the lungs, and particles less than 0.1 to 0.5 m in diameter may lea ve the lungs with the e xhaled breath. P articles larger than 8 to 10 m in diameter are separated and retained in the upper respiratory tract. Particles between 2 and 8 m in diameter are deposited mainly in the conducting airways of the lungs and are swallowed or coughed out quickly. 2. Combustion products. Carbon monoxide (CO) is a colorless, odorless gas, a product of incomplete comb ustion. CO interferes with the deli very of oxygen throughout the body . NO 2 is a combustion product from gas sto ves and other sources. There is gro wing evidence that NO 2 may cause respiratory disease. Indoor concentrations for CO and NO 2 are the same as speci f ed in the National Primary Ambient-Air Quality Standard by the EPA later in this section. 3. Volatile organic compounds (V OCs). These include formaldehyde and a v ariety of aliphatic, aromatic, oxygenated, and chlorinated compounds. Mucous membrane irritation caused by formaldehyde is well established. U.S. Department of Housing and Urban De velopment (HUD) specif es a tar get le vel of indoor concentration of formaldehyde for manuf acturing homes of 0.4 ppm. 4. Nicotine. Environmental tobacco smok e is clearly a discomfort f actor to man y adults who do not smoke. Nicotine and other components of tobacco smoke are also a health risk for human beings. 5. Radon. Radon is a colorless, odorless, inert radiative gas widely found in soil, rocks, and water, created by the decay of the radium and uranium. It tra vels through the pores of rock and soil and inf ltrates into a house along cracks and other openings in the basement slab or w alls; pressure-driven radon containing soil gas is caused by thermal stack, wind, and the mechanical ventilation system. The annual average concentration of radon in residential b uildings in the United States is about 1.25 pCi /L. Only about 6 percent of U.S. homes ha ve an annual a verage radon concentration e xceeding the EPA recommended annual a verage indoor concentration of 4 pCi /L for residential and school occupancies. Pressure-driven f ow of radon containing soil gas is the primary source for elevated concentrations. At various locations in the United States, the indoor radon concentrations may vary hourly, daily, and seasonally, sometimes by as much as a f actor of 10 to 20 on a daily basis. The radiative decay of radon produces a series of radioacti ve isotopes called progeny. These progeny are chemically active. They can deposit directly into the lung, or attach to airborne particles and then deposit into the lung. Some of the progen y are alpha particle emitters and may lead to cellular changes and initiate lung cancer. 6. Occupant-generated contaminants and odor s. These include odors and emissions (bioef f uents) from the human body, particulates, and other contaminants. 7. Bioaerosols. These contaminants include bacteria, mold and mide w, viruses, and pollens. Bacteria and viruses are airborne, carried by dust or transmitted by people and animals; standing water (wet surf ace) and dirt (nutrients) can become the breeding ground for mold, mildew, and other biological contaminants. Pollens originate from plants. In addition to the preceding indoor air contaminants, others such as sulfur dioxide and ozone can ha ve a signi f cant ef fect on occupants. Carbon dioxide (CO 2) is a kind of gas released from human beings and is not an indoor contaminant at the concentrations found in most b uildings. ASHRAE Standard 62-1999 specif es guidelines for indoor concentration for ozone during continuous exposure time as 100 g/m3.

INDOOR AND OUTDOOR DESIGN CONDITIONS

4.29

Basic Strategies to Improve Indoor Air Quality There are three basic strate gies to impro ve indoor air quality: control the contaminated source, remove air contaminants from the indoor air by air cleaner , and use outdoor ventilation air to dilute the concentrations of indoor air contaminants. To eliminate or to reduce the emissions of air contaminants from the contaminated source is often the most ef fective way to impro ve the IA Q as well as periodically cleaning the duct ’s interior surface and coil’s condensate pan and using building materials and carpets that do not release or release only ne gligible volatile organic compounds and dust. Ho wever, emissions and odor released from occupants are diff cult to eliminate or to reduce, and smoking is now prohibited in many public places and limited to specif ed areas in many commercial buildings. The volatile organic compounds and combustion products contain numerous minute particles of size between 0.003 and 1 m. Only high-ef f ciency air f lters and acti vated carbon f lters can remove these minute particles and odors from the airstream ef fectively. High eff ciency air f lters and carbon f lters are expensive to install, operate, and maintain. Adequate outdoor ventilation air to dilute the air contaminations in practice has been pro ved an essential, practical, and cost-effective means to impro ve the indoor air quality . ASHRAE Standard 62-1999, Ventilation for Acceptable Indoor Air Quality, specif es two alternative procedures to obtain acceptable IAQ: the ventilation rate procedure and indoor air quality procedure. ●

●

In the v entilation rate procedure, acceptable indoor air quality is achie ved by pro viding ventilation air of specif ed quality and quantity to the space. In the IA Q procedure, acceptable air quality is achie ved within the space by controlling kno wn and specif able contaminants.

The ventilation rate and IAQ procedurs are discused again in Sec. 23.2. ASHRAE Standard 62-1999 de f ned acceptable indoor air quality as air in which there are no known contaminants at harmful concentrations as determined by cognizant authorities and with which a substantial majority (80 percent or more) of the people e xposed do not e xpress dissatisfaction. ASHRAE Standard 62-1999 de f nes ventilation air as that portion of supply air that is outdoor air plus an y recirculated air that has been treated for the purpose of maintaining acceptable indoor air quality. Ventilation is the process of supplying and removing ventilation by natural and mechanical means, and the v entilation rate means the rate of v entilation air supplied to the conditioned space through the air system. If outdoor air is used to dilute the concentration of indoor contaminants, its quality must meet the National Primary Ambient-Air Quality Standard provided by the EPA. Part of the time a verage concentrations are shown below: Long-term concentration Pollutants

g/m3

ppm

Average period of exposure

Particulate matter Sulfur oxides Carbon monoxide

50 80

0.03

1 year 1 year

Nitrogen dioxide Oxidants (ozone) Lead

100

0.055

1 year

1.5

3 months

Short-term concentration g/m3

ppm

Average period of exposure

150 365 40,000 10,000

0.14 35 9

24 hours 24 hours 1 hour 8 hours

235

0.12

1 hour

Only particulate matter is expressed in annual geometric means; the other two are annual arithmetic means. For carbon monoxide and ozone, both values are not to be e xceeded more than once a year.

4.30

CHAPTER FOUR

Outdoor Air Requirements for Occupants For both comfort and process air conditioning systems, outdoor air is required to do the follo wing: ●

●

●

●

To meet metabolic requirements of the occupants To dilute the indoor air contaminants, odors, and pollutants to maintain an acceptable indoor air quality To support any combustion process or replace the amount of e xhaust air required in laboratories, manufacturing processes, or restrooms To pro vide mak eup for the amount of e xf ltrated air required when a positi ve pressure is to be maintained in a conditioned space

The amount of outdoor air required for metabolic oxidation processes for occupants is actually rather small. ASHRAE Standard 62-1999 noted that where only dilution v entilation is used to control indoor air quality , a CO 2 indoor-to-outdoor differential concentration (Ci,CO2  Co) is not greater than about 700 ppm, and the CO 2 production V˙CO2 of a sedentary occupant who is eating a normal diet is 0.0106 cfm (0.3 L / min). The amount of outdoor air required for each indoor occupant V˙o, oc can be calculated as

V˙o,oc 

V˙CO2 Ci,CO2  Co



0.0106 0.0007

 15 cfm (7 L / s)

(4.26)

Thayer (1982), using data from dif ferent authors, developed a dilution inde x that indicates 15 cfm (7 L /s) of outdoor air per person will satisfy more than 80 percent of the occupants in the space. Usually, for comfort air conditioning systems, the same outdoor air used for the dilution of the concentration of air contaminants including human bioef f uents is suff cient for the metabolic oxygen requirement, for exhausting air from restrooms, and for replacing e xf ltrated air lost from the conditioned space as a result of positive pressure. If the outdoor air supply is used to dilute the concentration of a speci f c indoor air contaminant, the rate of outdoor air supply V˙o, in cfm (L /min), can be calculated as V˙o 

2118m˙ par Ci  Co

(4.27)

where m˙ par  rate of generation of contaminants in space, mg/s Ci, Co  concentrations of air contaminants indoors and outdoors, respectively, mg/m3 Values of Co can be found from the EPA National Primary Ambient-Air Quality Standards. The indoor concentration of CO 2 and other contaminants should meet the speci f ed value as stated before. Some of the outdoor air requirements for v entilation, often called the v entilation rate, specif ed in ASHRAE Standard 62-1999 are indicated in Table 4.5. F or clean rooms, Federal Standard 209B specif es the rate of outdoor air, or makeup air, to be 5 to 20 percent of the supply air. When only dilution ventilation is used to control indoor air quality and CO 2 is used as an indicator of human bioef f uents, ASHRAE Standard 62-1999 noted that an indoor -outdoor dif ferential concentration not greater than about 700 ppm of CO 2 indicates that comfort (odor) criteria related to bioef f uents are lik ely to be satis f ed. The CO2 concentrations in outdoor air typically are between 300 to 350 ppm. Using CO 2 as an indicator of bioef f uents does not elimnate the need for consideration of other contaminants. The refrigeration capacity required to cool and dehumidify the outdoor air can be a major por-

INDOOR AND OUTDOOR DESIGN CONDITIONS

4.31

TABLE 4.5 Outdoor Air Requirements Recommended by ASHRAE Standard 62 – 1999 Application

cfm/person

Dining room Bar and cocktail lounges Conference rooms Off ce spaces Off ce conference rooms Retail stores Beauty shops Ballrooms and discos Spectator areas Theater auditoriums Transportation waiting rooms Classrooms Hospital patient rooms Residences Smoking lounges

20 30 20 20 20 0.2 – 0.3* 25 25 15 15 15 15 25 0.35† 60

*cfm/ft2 f oor area † Air changes/h Source: Abridged with permission from ASHRAE Standard 62-1999.

tion of the total refrigeration requirement during summer, depending upon the occupant density and the amount of e xhaust air. The amount of in f ltration depends on the wind speed and direction, as well as the outdoor and indoor temperatures and pressure dif ferences, which are v ariable. Therefore, inf ltration is not a reliable source of outdoor air supply . Inf ltration cannot replace speci f ed outdoor air ventilation requirements.

4.11 AIR CLEANLINESS The manufacturing process of semiconductors, pharmaceutical, aerospace, and operating rooms in health care facilities need clean indoor environment, clean spaces, and clean rooms. A clean room is a constructed enclosed area in which air cleanliness is e xpressed in terms of particle count of air contaminants and in which the associated temperature, humidity, air pressure, and lighting are controlled within speci f c limits. A clean space is a de f ned area in which air cleanliness and en vironmental conditions are controlled within specif c limits.The quality of their manufactured products is closely related to the size and number of particulates contained in the space air . An indoor air quality of allo wable total particulate annual-a verage concentration of 50 g/m3 cannot meet the air cleanliness requirements. Therefore, Federal Standard (FS) 209E speci f es the following classes for clean spaces and clean rooms: Class 1. Particle count not to e xceed 1 particle /ft3 (35 particles /m3) of a size of 0.5 m and larger, with no particle exceeding 5 m. Class 10. Particle count not to exceed 10 particles /ft3 (353 particles /m3) of a size of 0.5 m and larger, with no particle exceeding 5 m. Class 100. Particle count not to exceed 100 particles /ft3 (3531 particles /m3) of a size of 0.5 m and larger. Class 1000. Particle count not to e xceed 1000 particles /ft3 (35,315 particles /m3) of a size of 0.5 m and larger. Class 10,000. Particle count not to e xceed 10,000 particles /ft3 (353,150 particles /m3) of a size

4.32

CHAPTER FOUR

of 0.5 m and larger or 65 particles/ft3 (2295 particles/ft3) of a size 5.0 m and larger. Class 100,000. Particle count not to e xceed 100,000 particles /ft3 (3,531,500 particles /m3) of a size of 0.5 m and larger or 700 particles /ft3 (24,720 particles /ft3) of a size 5.0 m and larger. Refer to Federal Standard 209E for more details. Since workers in these clean rooms wear protective gowns and hats, in order to provide these air cleanliness classes, a year-round constant temperature and associated relative humidity and a specif ed unidirectional air f ow should be maintained. High-ef f ciency particulate air (HEPA) f lters and ultra-low-penetration air (ULPA) f lters should be installed in the air conditioning systems. Special building materials with hard and clean surf aces should be used as the b uilding envelope. More details are covered in Chap. 30.

4.12 SOUND LEVEL Sound and Sound Level Sound can be def ned as a variation in pressure due to vibration in an elastic medium such as air . A vibrating body generates pressure w aves, which spread by alternate compression and raref action of the molecules within the transmitting medium. Airborne sound is a v ariation of air pressure, with atmospheric pressure as the mean v alue. Because sound is transmitted by compression and e xpansion of molecules, it cannot tra vel in a v acuum. The denser the material, the f aster the tra veling speed of a sound wave. The velocity of a sound wave in air is approximately 1130 fps. In water, it is about 4500 fps and in steel 15,000 fps. Noise is any unwanted sound. In air systems, noise should be compensated for, either by attenuation (the process of reducing the amount of sound that reaches the space) or by masking it with other, less objectionable sounds.

Sound Power Level and Sound Pressure Level Sound power is the ability to radiate power from a sound source excited by an energy input. The intensity of sound power is the power output from a sound source expressed in watts (W). Because of the wide v ariation of sound output — from the threshold hearing le vel of 10 12 W to a le vel of 108 W, generated by the launching of a Saturn rock et, a ratio of 10 20 to 1 — it is more appropriate and convenient to use a logarithmic expression to def ne sound power level, i.e., L w  10 log

w re 1 pW 10 12 W

(4.28)

where Lw  sound power level, dB w  sound source power output, W Here 10 12 W, or 1 pW (picowatt), is the international reference base, and re indicates the reference base. The human ear and microphones are pressure-sensitive. Analogous to the sound power level, the sound pressure level is def ned as L p  20 log where Lp  sound pressure level, dB p  sound pressure, Pa

p re 20 Pa 2 105 Pa

(4.29)

INDOOR AND OUTDOOR DESIGN CONDITIONS

4.33

TABLE 4.6 Typical Sound Pressure Levels Source Military jet takeoff at 100 ft Passenger’s ramp at jet airliner (peak) Platform of subway station (steel wheels) Computer printout room* Conversational speech at 3 ft Window air conditioner* Quiet residential area* Whispered conversation at 6 ft Buzzing insect at 3 ft Threshold of good hearing Threshold of excellent youthful hearing

Sound pressure, Pa

Sound pressure level dB re 20 Pa

200

140

20

120

2 0.2 0.02 0.006 0.002 0.0006 0.0002 0.00006

100 80 60 50 40 30 20 10

0.00002

0

Subjective reaction Extreme Threshold of pain

Moderate

Faint Threshold of hearing

*Ambient. Source: Abridged from ASHRAE Handbook 1989, Fundamentals. Reprinted by permission.

Here the reference sound pressure le vel is 2 10 –5 Pa (pascal), or 20 Pa, corresponding to the hearing threshold. Because sound po wer is proportional to the square of the sound pressure, 10 log p2  20 log p. Sound pressure levels of various sources are listed in Table 4.6. The sound power level of a speci f c source is a f xed output. It cannot be measured directly and can only be calculated through the measurement of the sound pressure le vel. On the other hand, sound pressure level is the sound level measured at any one point and is a function of distance from the source and characteristics of the surroundings.

Octave Bands Sound w aves, like other w aves, are characterized by the relationship between w avelength, speed, and frequency: Wavelength 

speed frequency

(4.30)

People can hear frequencies from 20 Hz to 20 kHz. To study and analyze sound, we must break it down into components. A convenient way is to subdivide the audible range into eight octa ve bands or sometimes 24 13 - octave bands. An octave is a frequency band in which the frequency of the upper band limit is double the frequency of the lower band limit. The center frequency of an octave or a 13-octave band is the geometric mean of its upper and lower band limits. An octave or 13-octave band is represented by its center frequency. The eight octave bands and their center frequencies are listed in Table 4.7.

Addition of Sound Levels Because sound levels, in dB, are in logarithmic units, two sound levels cannot be added arithmetically. If sound le vels A, B, C, . . ., in dB, are added, the combined o verall sound level L can be calculated as

L  10 log (100.1A  100.1B  100.1C  · · ·)

(4.31)

4.34

CHAPTER FOUR

TABLE 4.7 Octave Bands and Their Center Frequencies Band frequency, H2

Band number

Lower

Center

Upper

1 2 3 4 5 6 7 8

22.4 45 90 180 355 710 1,400 2,800 5,600

31.5 63 125 250 500 1,000 2,000 4,000 8,000

45 90 180 355 710 1,400 2,800 5,600 11,200

Source: Abridged with permission from ASHRAE Handbook 1989, Fundamentals with permission.

Human Response and Design Criteria The human brain does not respond in the same w ay to lo wer frequencies as to higher frequencies. At lower sound pressure levels, it judges a 20-dB sound at 1000 Hz to ha ve the same loudness as a 52-dB sound at 50 Hz. Ho wever, at high sound pressure le vels, a 100-dB sound at 1000 Hz seems as loud as 110 dB at 50 Hz. The purpose of noise control in an air conditioned space is to pro vide a background sound lo w enough to avoid interference with the acoustical requirements of the occupants. The distribution of the background sound should be balanced o ver a wide range of frequencies, without whistle, hum, rumble, or audible beats. Three types of criteria for sound control are currently used in indoor system design: 1. A-weighted sound level dBA. The A-weighted sound level dBA tries to simulate the response of the human ear to sound at lo w sound pressure levels. An electronic weighting network automatically simulates the lo wer sensitivity of the human ear to lo wer-frequency sounds by subtracting a certain number of decibels at v arious octave bands, such as approximately 27 dB in the f rst octave band, 16 dB in the second, 8 dB in the third, and 4 dB in the fourth. The A-weighted sound le vel gives a single v alue. It is simple and also tak es into consideration the human judgment of relati ve loudness at lo w sound pressure le vels. Its main dra wback is its f ailure to consider the frequenc y spectrum or the subjective quality of sound. 2. Noise criteria, or NC, curves. NC curves represent actual human reactions during tests. The shape of NC curves is similar to the equal loudness contour representing the response of the human ear, as shown in Fig. 4.10. NC curv es are intended to indicate the permissible sound pressure le vel of a broadband noise at v arious octave bands by a single NC curv e sound le vel rating. NC curv es are practical. They also consider the frequency spectrum of the broadband noise. The main problem with NC curv es is that the shape of the curv e does not approach a balanced, bland-sounding spectrum that is neither rumbly nor hissy. 3. Room criteria, or RC, curves. RC curv es, as shown in Fig. 4.11, are similar to NC curv es except that the shape of an RC curv e is a close approximation of a balanced, bland-sounding spectrum. ASHRAE recommends the indoor design RC or NC ranges presented in Table 4.8. F or sounds containing signif cant pure tones or impulsive sounds, a 5- to 10-dB lower value should be specif ed. Noise is always an annoying element and source of complaints in indoor en vironments. Attenuation

INDOOR AND OUTDOOR DESIGN CONDITIONS

FIGURE 4.10 Noise criteria (NC) curv es. (Source: ASHRAE Handbook 1989, permission.)

4.35

Fundamentals. Reprinted with

to achieve an NC or RC goal less than 30 dB for a central air conditioning system or a packaged system is v ery e xpensive. In actual practice, the NC or RC criteria range for pri vate residences and apartments varies greatly because of personal requirements and economic considerations. To meet the listed design criteria range, the measured sound pressure levels of at least three of the four octa ve bands between 250 and 2000 Hz should be within the listed NC or RC range. In industrial workshops with machinery and equipment, Occupational Safety and Health Administration

4.36

CHAPTER FOUR

FIGURE 4.11 Room criteria (RC) curv es. (Source: ASHRAE Handbook 1989, permission.)

Fundamentals. Reprinted with

(OSHA) Standard P art 1910.95, published by the U.S. Department of Labor , specif es: “Feasible administrative or engineering controls shall be utilized when emplo yees are subjected to sound levels exceeding those in Table G-16. If such controls f ail to reduce sound le vels to below those in Table G-16, personal protective equipment shall be pro vided and used. Exposure to impulsi ve and impact noise should not exceed 140 dB peak sound pressure level.” Table G-16 is reproduced in this text as Table 4.9.

INDOOR AND OUTDOOR DESIGN CONDITIONS

4.37

TABLE 4.8 Recommended Criteria for Indoor Design RC or NC Range Type of area

Recommended criteria for RC or NC range

Private residence

25 – 30

Apartments

25 – 30

Hotels/motels Individual rooms or suites Meeting/banquet rooms Halls, corridors, lobbies Service/support areas

30 – 35 25 – 30 35 – 40 40 – 45

Off ces Executive Conference Private Open-plan Computer equipment rooms Public circulation

25 – 30 25 – 30 30 – 35 30 – 35 40 – 45 40 – 45

Hospitals and clinics Private rooms Wards Operating rooms Corridors Public areas

25 – 30 30 – 35 35 – 40 35 – 40 35 – 40

Churches

25 – 30

Schools Lecture and classrooms Open-plan classrooms

Duration Sound level per day, h dBA slow response

25 – 30 30 – 35

Libraries

35 – 40

8 6 4 3 2 1 12 1

Concert halls

†

Legitimate theaters

†

Recording studios Movie theaters

† 30 – 35

Note: These are for unoccupied spaces, with all systems operating. *Design goals can be increased by 5 dB when dictated by budget constraints or when intrusion from other sources represents a limiting condition. † An acoustical expert should be consulted for guidance on these critical spaces. Source: ASHRAE Handbook 1987, HVAC Systems and Applications. Reprinted with permission.

TABLE 4.9 Occupational Noise Exposure (Occupational Health and Safety Administration Table G-16)

1 2 1 4

(or less)

90 92 95 97 100 102 105 110 115

Note: If the variations in noise level involve maxima at intervals of 1 s or less, it is to be considered continuous. In all cases where the sound levels exceed the values shown herein, a continuing, effective hearing conservation program shall be administered. Source: Occupational Safety and Health Standard, Part 1910.95. Reprinted with permission.

4.13 SPACE PRESSURE DIFFERENTIAL Most air conditioning systems stri ve to maintain a slightly higher indoor pressure than that of the surroundings. This positive pressure tends to eliminate or reduce in f ltration, the entry of untreated air to the space. Ne gative space pressure may cause space high humidity le vels, mold and milde w growth, combustion equipment backdraft, and entry of se wer gas. F or rooms where toxic,

4.38

CHAPTER FOUR

hazardous, contaminated, or objectionable gases or substances are produced, a slightly lo wer pressure, or an appropriate negative room pressure, should be maintained to prevent the diffusion of these substances to the surroundings and at the same time pre vent and reduce the damage of the uncontrolled outdoor airf ow. The magnitude of the positive or negative pressure to be maintained in the space must be carefully determined. A higher positi ve pressure al ways means a greater amount of e xfiltrated air , which creates requirements for greater v olumes of outdoor air intak e. A higher negative pressure means a greater infiltration. Both result in higher initial and operating costs. Space pressurization is closely related to the amount of ef fective leakage area, or tightness of the b uilding. Cummings et al. (1996) reported that the space pressure dif ferentials between indoor and outdoor air and building tightness v ary widely for se ven field-surveyed restaurants. F or comfort systems in lo wrise b uildings, the recommended pressure dif ferential is 0.005 to 0.03 in. WC ( 1.25 to 7.5 P a), which is often adopted for a verage b uilding tightness. F or a 44-story high-rise building, the pressure dif ferential between the entrance lobby outdoors and rooftop floor because of the stack ef fect during winter may e xceed 0.3 in. WC (75 Pa). The phrase in. WC represents the pressure at the bottom of a w ater column that has a height of the specif ied number of inches. Another type of space pressurization control for the health and comfort of passengers in an aircraft is cabin pressurization control. If an airplane is f ying at an altitude of 28,000 ft (8540 m) with an ambient pressure of 4.8 psia (33.1 kP a abs.), the minimum cabin pressure required is 10.8 psia (75 kPa abs.); and a cabin pressurization control system must be installed to maintain a pressure differential of 10.8  4.8  6 psi (41 kPa) between the cabin and the ambient air. Process air conditioning systems, such as for clean rooms, need properly speci f ed space pressurization to pre vent contaminated air from entering the clean, uncontaminated area from surrounding contaminated or semicontaminated areas. According to Federal Standard 209B, Clean Rooms and Work Station Requirements, Controlled Environment, published by the U.S. government in 1973, the minimum positive pressure differential between the clean room and an y adjacent area of less clean requirements should be 0.05 in. WC (12.5 Pa) with all entryways closed. When the entryways are open, an outward f ow of air is to be maintained to minimize migration of contaminants into the room. Space pressurization is discussed in Chaps. 20 and 23.

4.14 OUTDOOR DESIGN CONDITIONS In principle, the capacity of air conditioning equipment is selected so that indoor design conditions can be maintained when the outdoor weather does not e xceed the design v alues. The outdoor weather affects the space cooling load and the capacity of the air system to condition the required amount of outdoor air. It is not economical to choose the annual maximum or annual minimum v alues as the design data. Outdoor design conditions are usually determined according to statistical analysis of the weather data of the pre vious 30 years so that either 99.6 or 99 percent of the winter indoor design conditions can then be attained annually , or only 0.4, 1.0, or 2.0 percent of the time annually the summer indoor design conditions are e xceeded. The greater the need for strict control of indoor environmental parameters, the greater the percentage co verage of the winter outdoor design hours or the smaller the percentage occurrence of the total hours that e xceed the outdoor design values. The recommended design v alues are based on data from the National Climatic Data Center (NCDC) from 1961 to 1991, or from 1982 to 1993, and Canadian Weather Energy and Engineering Data Sets (CWEEDS) from 1953 through 1993. Per ASHRAE Handbook 1997, Fundamentals, and ASHRAE/IESNA Standard 90.1-1999, the design conditions recommended for lar ge cities in the United States are listed in Table 4.10.

INDOOR AND OUTDOOR DESIGN CONDITIONS

4.39

Summer and Winter Outdoor Design Conditions In Table 4.10, note the following: ●

●

●

●

●

●

Summer outdoor design dry-bulb temperature for a specif c locality To,s, in °F (°C), is the rounded higher integral number of the statistically determined summer design temperature To,ss such that the average number of hours of occurrence of outdoor temperatures To higher than To,ss annually is on average 0.4 percent (35 h), 1 percent (88 h), or 2 percent (175 h). The summer outdoor mean coincident design wet-b ulb temperature To, s, in °F (°C), is the mean of all wet-b ulb temperatures occurring at the speci f c summer outdoor design dry-b ulb temperature To,s during the summer. The 1.0 percent summer design wet-b ulb temperature To1 is the design v alue having an a verage annual occurrence of To  To1 of 88 h. Variable To1 is used for e vaporative cooling systems, cooling towers, and evaporative condensers, which are covered in later chapters. The mean daily range, in °F (°C), indicates the mean of the dif ference between the daily maximum and minimum temperatures for the warmest month. The winter outdoor design dry-bulb temperature To,w, in °F, is the rounded lower integral value of the statistically determined winter outdoor design temperature To,ws, such that the annual a verage number of hours of occurrence of outdoor temperature at v alues To  To,ws should be equaled or exceeded 99.6 or 99 percent of the total number of annual hours. The annual a verage number of hours in winter when To 0.99To,ws is 35 h, and To 0.99To,ws is 88 h. The number of degree-days is the difference between a base temperature and the mean daily outdoor air temperature for an y one day Tbase  To,m, both in °F. Annually, the total number of heating degree-days with a base temperature of 65°F, or HDD65, is HDD65 

 (65  To, m)

n1

(4.32)

where n  number of days whose To,m 65°F per annum. The total number of cooling de greedays with a base temperature of 50°F, or CDD50, is CDD50 

 (To, m  50)

m1

(4.33)

where m  number of days whose To,m 50°F per annum. Heating and cooling degree-days with different base temperatures ha ve been used as climatic parameters to calculate the ener gy f ux through b uilding en velope, or to determine the U value and the con f guration of the b uilding envelope. ●

In the last tw o columns, MWS/MWD to DB 99.6% indicates mean coincident wind speed (MWS)/mean coincident wind direction (MWD), i.e., most frequently occurring with the heating dry-bulb 99.6 percent. In MWD, wind direction is expressed in degrees; 270° represents west, and 180° south.

The Use of Outdoor Weather Data in Design During the design of air conditioning systems, the following parameters are often adopted: 1. Indoor and outdoor design conditions are used to calculate the space cooling and heating loads. 2. Summer outdoor dry-b ulb and coincident wet-b ulb temperatures are necessary to e valuate the coil load. The summer outdoor wet-b ulb temperature is used to determine the capacity of the evaporative coolers, cooling towers, and evaporative condensers.

4.40

TABLE 4.10 Climatic Conditions for the United States and Canada Winter

Summer Design dry-bulb/ mean coincident wet-bulb

Design dry-bulb City Albuquerque Anchorage Atlanta Atlantic City CO Baltimore Billings Birmingham Bismarck Boise Boston Bridgeport Buffalo Burlington Caribou Casper Charleston Charleston Charlotte Chicago, Cincinnati CO Cleveland Concord Dallas/ Fort Worth Denver Des Moines Detroit Honolulu Houston Indianapolis Jackson Kansas City Lansing Las Vegas

99.6%

99%

0.4%

1%

2%

°F

°F

°F

°F

°F

Mean daily range, °F

5,315 131 1,033

13 14 18

18 9 23

96/60 71/59 93/75

93/60 68/57 91/74

91/60 65/56 88/73

25.4 12.6 17.3

64 58 76

39.45 39.18 45.80 33.57 46.77 43.57 42.37 41.17 42.93 44.47 46.87 42.92 32.90 38.37 35.22 41.98 39.10 41.42 43.20

66 154 3,570 630 1,660 2,867 30 16 705 341 623 5,289 49 981 768 673 482 804 344

8 11 13 18 21 2 7 8 2 11 14 13 25 6 18 6 5 1 8

13 15 7 23 16 9 12 12 5 6 10 5 28 11 23 1 12 6 2

91/74 93/75 93/63 94/75 93/68 96/63 91/73 86/73 86/70 87/71 85/69 92/59 94/78 91/73 94/74 91/74 93/74 89/73 90/71

88/73 91/74 90/62 92/75 90/67 94/63 87/71 84/72 84/69 84/69 82/67 89/58 92/77 88/73 91/74 88/73 90/75 86/72 87/70

86/72 88/73 87/61 90/74 86/66 91/62 84/70 82/71 81/68 82/68 79/66 86/58 90/77 86/71 89/73 86/71 88/73 84/71 84/68

18.1 18.8 25.8 18.7 26.5 30.3 15.3 14.1 17.7 20.4 19.5 30.4 16.2 19.1 17.8 19.6 20.0 18.6 24.1

76 76 64 77 70 64 74 74 72 72 70 61 79 75 76 75 76 74 73

32.90 39.75 41.53 42.23 21.35 29.97 39.73 32.32 39.32 42.77 36.08

597 5,331 965 663 16 108 807 331 1,024 873 2,178

17 3 9 0 61 27 3 21 1 3 27

24 3 4 5 63 31 3 25 4 2 30

100/74 93/60 93/76 90/73 89/73 96/77 91/75 95/77 96/75 89/73 108/66

98/74 90/59 90/74 87/72 88/73 94/77 88/74 93/76 93/75 86/72 106/66

96/74 87/59 87/73 84/70 87/73 92/77 86/73 92/76 90/74 84/70 103/65

20.3 26.9 18.5 20.4 12.2 18.2 18.9 19.2 18.8 21.7 24.8

77 63 76 74 75 79 77 79 77 74 70

Lat, degree

Elevation, ft

New Mexico Alaska Georgia

35.05 61.17 33.65

New Jersey Maryland Montana Alabama North Dakota Idaho Massachusetts Connecticut New York Vermont Maine Wyoming South Carolina West Virginia North Carolina Illinois Ohio Ohio New Hampshire Texas Colorado Iowa Michigan Hawaii Texas Indiana Mississippi Missouri Michigan Nevada

State

Annual Annual average daily cooling Design incident degreewet- solar days bulb, energy on base 1% east or west, 50°F, °F Btu / ft2  day CDD50

Annual heating degreedays base 65°F, HDD65

MWS MWD

4,425 10,570 2,991

8 8 12

360 290 320 310 290 230 340 290 130 320 320 270 70 270 260 20 250 50 270 260 230 320 350 180 320 240 320 340 230 340 320 290 250

MWS/MWD to DB 99.6%

1,105 538 807

3,908

739 814 789 766 916 659

3,709 2,466 5,206 2,144 2,807 2,897 2,997 2,468 2,228 1,470 2,082 6,188 3,655 4,704 2,941 3,733 2,755 2,087

4,707 7,164 2,918 8,968 5,861 5,641 5,537 6,747 7,771 9,651 7,682 2,013 4,646 6,536 4,988 6,201 7,554

9 10 10 7 7 6 17 14 12 6 10 9 7 7 6 10 9 12 4

6,587 2,732 3,371 3,046 9,949 6,876 3,453 5,900 3,852 2,449 6,745

2,259 6,020 6,497 6,167 0 1,599 5,615 2,467 5,393 7,101 2,407

13 6 11 11 5 8 8 7 10 8 7

609 698 649 961 796 667 809 729

630 875 971 788 676 953 805 692 833

1,136

5,038

Lincoln CO Little Rock Los Angeles CO Louisville Memphis Miami Milwaukee Minneapolis/ St. Paul New Orleans New York Norfolk Phoenix Pittsburgh Portland CO Providence Rapid City St. Louis Salt Lake City San Antonio San Diego San Francisco CO Seattle CO Shreveport Spokane Syracuse Tucson Tulsa Washington Wichita Wilmington Canada Calgary Montreal Regina Toronto Vancouver Winnipeg

Nebraska Arkansas California Kentucky Tennessee Florida Wisconsin

40.85 34.92 33.93 38.18 35.05 25.82 42.95

1,188 312 105 489 285 13 692

7 16 43 6 16 46 7

2 21 45 12 21 50 2

97/74 97/77 85/64 93/76 96/78 91/77 89/74

94/74 95/77 81/64 90/75 94/77 90/77 86/72

91/73 92/76 78/64 88/74 92/77 89/77 83/70

22.3 19.5 10.9 18.2 16.8 11.4 16.6

76 79 69 77 79 79 74

831 962 727 806 874 724

3,455 5,299 4,777 4,000 5,467 9,474 2,388

6,278 3,155 1,458 4,514 3,082 200 7,324

9 9 6 10 10 10 13

350 360 70 290 20 340 290

Minnesota Louisiana New York Virginia Arizona Pennsylvania Oregon Rhode Island South Dakota Missouri Utah Texas California California

44.88 29.98 40.65 36.90 33.43 40.50 45.60 41.73 44.05 38.75 40.78 29.53 32.73 37.62

837 30 23 30 1,106 1,224 39 62 3,169 564 4,226 794 30 16

16 30 11 20 34 2 22 5 11 2 6 26 44 37

11 34 15 24 37 7 27 10 5 8 11 30 46 39

91/73 93/79 91/74 93/77 110/70 89/72 90/67 89/73 95/65 95/76 96/62 98/73 85/67 83/63

88/71 92/78 88/72 91/76 108/70 86/70 86/66 86/71 91/65 93/75 94/62 96/73 81/67 78/62

85/70 90/78 85/71 88/75 106/70 84/69 83/64 83/70 88/64 90/74 92/61 94/74 79/67 74/61

19.1 15.5 13.9 15.3 23 19.5 21.6 17.4 25.3 18.3 27.7 19.1 8.9 16.7

74 80 75 77 75 73 67 74 68 78 65 77 71 63

709 838 650 792 1,116 642 647 677 819 797 975 878 950 941

2,680 6,910 3,342 1,586 8,425 2,836 2,517 2,743 2,412 4,283 3,276 7,142 5,223 2,883

7,981 1,513 5,027 3,609 1,350 5,968 4,522 5,884 7,301 4,758 5,765 1,644 1,256 3,016

7 17 12 5 10 13 12 9 12 7 10 3 5

340 320 340 90 260 120 340 350 290 160 350 70 160

Washington Louisiana Washington New York Arizona Oklahoma DC Kansas Delaware

47.45 32.47 47.62 43.12 32.12 36.20 38.85 37.65 39.68

449 259 2,461 407 2,556 676 66 1,339 79

23 22 1 3 31 9 15 2 10

28 26 7 2 34 14 20 8 14

85/65 97/77 92/62 88/72 104/65 100/76 95/76 100/73 91/75

81/64 95/77 89/61 85/71 102/65 97/76 92/76 97/73 89/74

78/62 93/76 85/60 83/70 100/65 94/75 89/74 94/73 86/73

18.3 19.1 26.1 20.3 29.4 19.5 16.6 22.2 17.0

65 79 63 73 71 78 78 76 76

621 843 758 611 1,112 820 724

2,120 6,166 2,032 2,399 6,921 5,150

4,611 2,264 6,842 6,834 1,678 3,691

4,351 3,557

4,791 4,937

10 9 7 7 7 11 11 13 11

10 360 50 90 140 360 340 360 290

Alberta Quebec Saskatchewan Ontario British Columbia Manitoba

51.12 45.47 50.43 43.67 49.18 49.90

3,556 118 1,893 568 7 784

22 12 29 4 18 27

17 7 24 1 24 23

83/60 85/71 89/64 87/71 76/65 87/68

80/59 83/70 85/64 84/70 74/64 84/67

77/57 80/68 82/62 81/68 71/62 81/66

22.0 17.6 23.6 20.2 14.0 20.5

61 72 66 72 64 70

7 7 9 9 6 7

0 250 270 340 90 320

726

Note: CO designates locations within an urban area. Most of the data are taken from city airport temperature observations. Some semirural data are comparable to airport data. Source: Abridged with permission from ASHRAE Handbook 1997, Fundamentals and ASHRAE/IES Standard 90.1-1999.

4.41

4.42

CHAPTER FOUR

3. Outdoor weather data presented consecuti vely for a whole year of 8760 h, or other simpli f ed form, are sometimes used for year-round energy estimations. 4. Outdoor climate has a signif cant inf uence on the selection of an air conditioning system and its components.

Outdoor Weather Characteristics and Their Influence The following are characteristics of outdoor climate in the United States: 1. For clear days during summer , the daily maximum temperature occurs between 2 and 4 P.M. In winter, the daily minimum temperature usually occurs before sunrise, between 6 and 8 A.M. 2. The maximum combined in f uence of outdoor temperature and solar radiation on load calculations usually occurs in July or August. January is often the coldest month in many locations. 3. The mean daily temperature range of the w armest month v aries widely between dif ferent locations. The smallest mean daily range, 10°F (5.6 °C), occurs at Galv eston, Texas. The greatest, 45°F (25°C), occurs at Reno, Nevada. Many cities have a mean daily range of 15 to 25 °F (8.3 to 13.9°C). Usually, coastal areas have smaller mean daily ranges, and continental areas and areas of high elevation have greater values. 4. Extremes in the difference between 1 percent design dry-bulb temperature in summer and the 99 percent design temperature in winter in the United States are ● ●

A maximum of 119°F (66°C) at Bettles, Alaska A minimum of 18°F (10°C) at Kaneohe, MCAS, Hawaii

5. Snelling (1985) studied outdoor design temperatures for dif ferent locations all o ver the country. He found that extremely cold temperatures may last for 3 to 5 days. Extremely hot temperatures seldom last more than 24 h. 6. Among the 538 locations listed in Table 4.10, 138 locations ha ve a summer mean coincident wet-bulb temperature corresponding to 1 percent dry-b ulb, Tos  68°F (20 °C). There it is advantageous to use evaporative cooling systems to replace part of or all the refrigeration capacity . 7. When commercial buildings are only occupied after 10 A.M. or even after 12 noon, a winter outdoor design temperature higher than the 99.0 percent design dry-b ulb temperature should be considered.

REFERENCES ANSI/ASHRAE, ANSI/ASHRAE Standard 55-1992, Thermal Environmental Conditions for Human Occupancy, Atlanta, GA, 1992. ASHRAE, ASHRAE Standard 62-1999, Ventilation for Acceptable Indoor Air Quality, Atlanta, GA, 1999. ASHRAE, ASHRAE Handbook 1996, HVAC Systems and Equipment, Atlanta, GA, 1996. ASHRAE, ASHRAE Handbook 1997, Fundamentals, Atlanta, GA, 1997. ASHRAE Evironmental Health Committee 1987, Indoor Air Quality Position Paper, Atlanta, GA, 1987. Bayer, C. W., and Black, M. S., IAQ Evaluations of Three Off ce Buildings, ASHRAE Journal, July 1988, pp. 48 – 52. Bedford, T., and Warmer, C. G., The Globe Thermometer in Studies of Heating and Ventilation, Journal of the Institution of Heating and Ventilating Engineers, vol. 2, 1935, p. 544. Berglund, L., Mathematic Models for Predicting the Thermal Comfort Response of Building Occupants, ASHRAE Transactions, 1978, Part I, pp. 735 – 749. Cena, K., Spotila, J. R., and Avery, H. W., Thermal Comfort of the Elderly Is Affected by Clothing, Activity, and Psychological Adjustment, ASHRAE Transactions, 1986, Part II A, pp. 329 – 342.

INDOOR AND OUTDOOR DESIGN CONDITIONS

4.43

Cummings, J. B., Withers, Jr., C. R., Moyer, N. A., Fairey, P. W., and McKendry, B. B., Field Measurement of Uncontrolled Airf ow and Depressurization in Restaurants, ASHRAE Transactions, 1996, Part I, pp. 859 – 869. DeBat, R. J., Humidity: The Great Equalizer, HPAC, no. 10, 1996, pp. 66 – 71. Eade, R., Humidif cation Looming Larger in IAQ, Engineered Systems, no 1, 1996, pp. 49 – 58. Fanger, P. O., Thermal Comfort Analysis and Applications in Environmental Engineering, McGraw-Hill, New York, 1970. Federal Supply Service, Federal Standard No. 209B, Clean Rooms and Work Station Requirements, Controlled Environment, General Services Administration, Washington, 1973. Gagge, A. P., Stolwijk, J. A. J., and Nishi, Y., An Effective Temperature Scale Based on a Simple Model of Human Physiological Regulatory Response, ASHRAE Transactions, 1971, Part I, p. 247. George, A. C., Measurement of Sources and Air Concentrations of Radon and Radon Daughters in Residential Buildings, ASHRAE Transactions, 1985, Part II B, pp. 1945 – 1953. Holzle, A. M., Munson, D. M., and McCullough, E. A., A Validation Study of the ASHRAE Summer Comfort Envelope, ASHRAE Transactions, 1983, Part I B, pp. 126 – 138. Janssen, J. E., Ventilation for Acceptable Indoor Air Quality, ASHRAE Journal, October 1989, pp. 40 – 46. Kirkbride, J., Lee, H. K., and Moore, C., Health and Welfare Canada’s Experience in Indoor Air Quality Investigation, Indoor Air ’90, vol. 5, D. S. Walkinshaw, ed., Ottawa: International Conference on Indoor Air and Climate, 1990, pp. 99 – 106. McCullough, E. A., and Jones, B. W., A Comprehensive Data Base for Estimating Clothing Insulation, IER Technical Report 84 – 01, Kansas State University, 1984. National Institute for Occupational Safety and Health, Indoor Air Quality: Selected Reference, Division of Standards Development and Technology Transfer, Cincinnati, Ohio, 1989. O’Sullivan, P., Energy and IAQ Can Be Complementary, Heating/Piping/Air Conditioning, February 1989, pp. 37 – 42. Persily, A., Ventilation Rates in Off ce Buildings, ASHRAE Journal, July 1989, pp. 52 – 54. Rohles, Jr., F. H., Woods, J. E., and Nevins, R. G., The Effect of Air Movement and Temperature on the Thermal Sensations of Sedentary Man, ASHRAE Transactions, 1974, Part I, pp. 101 – 119. Seppanen, O., McNall, P. E., Munson, D. M., and Sprague, C. H., Thermal Insulating Values for Typical Clothing Ensembles, ASHRAE Transactions, 1972, Part I, pp. 120 – 130. Snelling, H. J., Duration Study for Heating and Air Conditoning Design Temperature, ASHRAE Transactions, 1985, Part II, pp. 242 – 249. Spain, S., The Upper Limit of Human Comfort from Measured and Calculated PMV Values in a National Bureau of Standards Test House, ASHRAE Transactions, 1986, Part I B, pp. 27 – 37. Sterling, E. M., Arundel, A., and Sterling, T. D., Criteria for Human Exposure to Humidity in Occupied Buildings, ASHRAE Transactions, 1985, Part I B, pp. 611–622. Thayer, W. W., Tobacco Smoke Dilution Recommendations for Comfort Ventilation, ASHRAE Transactions, 1982, Part II, pp. 291 – 304. Woodcock, A. H., Moisture Transfer in Textile Systems, Textile Research Journal, vol. 8, 1962, pp. 628 – 633.

CHAPTER 5

ENERGY MANAGEMENT AND CONTROL SYSTEMS 5.1 AUTOMATIC CONTROL SYSTEM AND HISTORICAL DEVELOPMENT 5.2 Air Conditioning Automatic Control System 5.2 Historical Development 5.2 Distribution of Energy Management and Control Systems 5.3 5.2 CONTROL LOOP AND CONTROL METHODS 5.5 Control Loop 5.5 Sequence of Operations 5.5 Control Methods 5.7 Comparison of Control Methods 5.3 CONTROL MODES 5.9 Two-Position Control 5.9 Step Control and Modulating Control 5.10 Floating Control 5.11 Proportional Control 5.11 Proportional plus Integral (PI) Control 5.13 Proportional-Integral-Derivative (PID) Control 5.14 Compensation Control, or Reset 5.15 Applications of Various Control Modes 5.15 5.4 SENSORS AND TRANSDUCERS 5.16 Temperature Sensors 5.18 Humidity Sensors 5.18 Pressure Sensors 5.19 Flow Sensors 5.19 Carbon Dioxide and Air Quality Sensors 5.20 Occupancy Sensors 5.20 Wireless Zone Sensors and Intelligent Network Sensors 5.21 Transducers or Transmitters 5.21 5.5 CONTROLLERS 5.21 Direct-Acting or Reverse-Acting 5.21 Normally Closed or Normally Open 5.22 Pneumatic Controllers 5.22 Electric and Electronic Controllers 5.23 Direct Digital Controllers 5.23 5.6 WATER CONTROL VALVES AND VALVE ACTUATORS 5.26 Valve Actuators 5.26 Types of Control Valves 5.27 Valve Characteristics and Ratings 5.27 Valve Selection 5.29 Valve Sizing 5.30 5.7 DAMPERS AND DAMPER ACTUATORS 5.32

Types of Volume Control Dampers 5.32 Damper Actuators (Motors) 5.33 Volume Flow Control between Various Airflow Paths 5.33 Flow Characteristics of Opposed- and Parallel-Blade Dampers 5.35 Damper Selection 5.37 Damper Sizing 5.37 5.8 SYSTEM ARCHITECTURE 5.38 Architecture of a Typical EMCS with DDC 5.38 System Characteristics 5.40 Future Development 5.41 5.9 INTEROPERABILIITY AND OPEN PROTOCOL BACnet 5.41 Interoperability 5.41 BACnet — Open Data Communication Protocol 5.41 Application Layer 5.42 Network Layer 5.43 Data Link/Physical Layer — Network Technology 5.43 Connection between BACnet and Proprietary Network 5.44 LonTalk Protocol 5.44 5.10 CONTROL LOGIC AND ARTIFICIAL INTELLIGENCE 5.45 Fuzzy Logic 5.45 Knowledge-Based Systems and Expert Systems 5.47 Artificial Neural Networks 5.51 5.11 PROGRAMMING FOR DDC SYSTEMS 5.53 Evolution of DDC Programming 5.53 Graphical Programming 5.53 Templates 5.54 Graphical Programming for Mechanical Cooling Control 5.55 5.12 TUNING DDC UNITS 5.55 Tuning PI Controllers 5.55 Self-Tuning PI and PID Controllers 5.55 Adaptive Control 5.56 5.13 FACTORS AFFECTING CONTROL PROCESSES 5.56 Load 5.56 Climate Change 5.56 System Capacity 5.57 Performance of Control Processes 5.57 Thermal Capacitance 5.58 5.14 FUNCTIONAL CONTROLS 5.58 Generic Controls 5.59

5.1

5.2

CHAPTER FIVE

Specific Controls 5.60 Commissioning and Maintenance 5.15 FAULT DETECTION AND DIAGNOSTICS 5.61 Basics 5.61 Expert System Rule-Based Diagnostics 5.62

5.61

ARX and ANN Model-Based Diagnosis 5.63 5.16 CONTROLS IN ASHRAE/IESNA STANDARD 90.1-1999 5.66 General 5.66 Off-Hour Controls 5.66 REFERENCES 5.67

5.1 AUTOMATIC CONTROL SYSTEM AND HISTORICAL DEVELOPMENT Air Conditioning Automatic Control System An air conditioning automatic control system or simply a control system, primarily modulates the capacity of the air conditioning equipment to maintain predetermined parameter(s) within an enclosure or for the fluid entering or le ving the equipment to meet the load and climate changes at optimum ener gy consumption and safe operation. The predetermined parameter or v ariable to be controlled is called the controlled variable. In air conditioning or in HV AC&R, the controlled variable can be temperature, relative humidity, pressure, enthalpy, fluid f w, contaminant concentration, etc. Because of the variation of the space load and the outdoor climate, a control system is one of the decisive factors for an air conditioning, or an HVAC&R, system to achie ve its goal: to effectively control the indoor en vironmental parameters, provide an adequate amount of outdoor air , be energy-efficient and pro vide better security and safety . Today nearly all HV AC&R systems are installed with a control system to provide effective operation and energy conservation.

Historical Development Early controls for comfort air conditioning systems used in the Capitol b uilding since 1928 were pneumatic controls that included high- and lo w-limit thermostats in the supply air dischar ge of the air-handling unit. These controls were used to maintain the desired supply air temperature. De wpoint control of the supply air maintained room humidity within a desirable range. A thermostat in the return-air passage was used to control the air volume fl w rate passing through the cooling coil, and a 16-point recorder connected with resistance thermometers w as used to record the room temperature at various locations. Shavit (1995) stated that the use of thermocouples of fers the possibility of remote monitoring of the space temperature. In coordination with pneumatic local control systems, the first centralize monitoring system with a central panel and remote set point change w as installed in the White House in 1950. The concept of a building automation system (BAS) which centralizes the monitoring management of b uilding services w as devised and de veloped in 1950s. The primary impro vement of the pneumatic control system occurred in 1963. The vapor and b ulb-based sensors were replaced by a remote pneumatic sensor using a bimetal strip and a nozzle flapper to r gulate the compressed air pressure between 3 and 15 psi (20.7 and 103.4 kPa). Shavit also indicated that the first on-line computer as introduced in 1967 in One Main Place in Dallas, Texas. This system had the first set ene gy conservation software including enthalpy control, demand limit, optimum start /stop, reset according to the zone of highest demand, chiller optimization, and night purge. In 1970, the introduction of all-electronic, solid-state centralized control system w as another significant milestone. Solid-state components impr ved the scanning process and serial transmission as well as reduced many wires in the trunk wiring to a single pair.

ENERGY MANAGEMENT AND CONTROL SYSTEMS

5.3

The 1973 energy crisis greatly boosted the control industry . Many energy management systems (EMSs) were installed for the purpose of sa ving ener gy. The microprocessor -based direct digital control (DDC) was first introduced in 1981 and the DDC unit controller in 198 – 1987. By means of the incorporated hardw are and softw are, it tremendously increased the control functions, developed the control logic, and accelerated the data processing and analysis.

Distribution of Energy Management and Control Systems A building automation system is a centralized monitoring, operation, and management system of the building services in a building including air conditioning (HVAC&R), lighting, fire protection and security. Only air conditioning control systems and smoke control systems are covered in this handbook. In the United States, a single-zone, two-position electric control system, as shown in Fig. 5.1, is now widely used in residential and light commercial b uildings that adopt DX packaged systems. Single-zone means that the load characteristics of the whole conditioned space are similar , it is controlled by one controller , and two-position means on-off control. This kind of control system is simple, easy to manage, and low in cost. For a large high-rise building constructed in the 1990s, an energy management and control system (EMCS), as shown in Fig. 5.2, is often the choice. Variable air volume indicates that airfl w is v aried to match the v ariation of the system load. Today, an EMCS is a microprocessor -based system with direct digital control which optimizes the operation and the indoor environmental parameters of an HVAC&R system in order to maintain a healthy and comfortable environment at optimum ener gy use. Today, an EMCS is an adv anced-generation energy management and control system and is also a part of a building automation system. According to the surv eyed results in Commercial Buildings Char acteristics 1992 by EIA, the percentages of the floor area ser ed by the installed EMCS among the total conditioned area in commercial buildings are as follows:

Refrigerant pipes

Furnace

Condensing unit

Thermostat

DX-coil

Supply fan FIGURE 5.1 A single-zone, two-position electric control system.

Compressor

CHAPTER FIVE

3

Year built

Building with EMCS, million ft2

Percentage of f oor space

Before 1970 1970 – 1979 1980 – 1989 1990 – 1992

5431 3313 4343 1236

18 29 36 58

1

1 Outdoor air damper

1

3

5

ru

DM1

T P SD

1

2

P2

DM4 T3

T2 SD T4

cc

sf

Cooling coil

Inlet vanes

Recirculating air damper

DM2

Relief air damper

4

Air-handler

T1

C MPS

6

2

P3

M

DM3

S

Relief fan Outdoor air reference

Conditioned space r

P1

Temperature sensor Pressure sensor Smoke detector

0

14.0

Manual positioning switch DM Damper actuator Multicompensator M

0.020

MPS

50 %

70

0.016

.5

13

0.012

60 r

50 40

cc

sf S

r

20%

ru

ru

m

0.004

sf S

0 40

50

0.008

w, lb/lb

5.4

60

70 T, F

80

90

100

0

FIGURE 5.2 Control diagram and psychrometric analysis of an energy management and control system for a single-zone, VAV cooling packaged system.

ENERGY MANAGEMENT AND CONTROL SYSTEMS

5.5

The later the building is constructed and the larger the conditioned area of the commercial building, the greater is the chance to install an energy management and control system (EMS).

5.2 CONTROL LOOP AND CONTROL METHODS Control Loop The basic element of a control system is a control loop. A control loop often consists of a sensor (such as T2 in Fig. 5.2), which senses and measures the controlled v ariable of recirculating air; a DDC unit controller which compares the sensed input signal with the predetermined condition (the set point) and sends an output signal to actuate the third element, a controlled device or control element (such as a damper or a valve in Fig. 5.2). Modulation, minute adjustments, or on-off control of dampers and v alves will change the controlled de vices position or operating status, which affects the controlled v ariable by changing the air f ow and w ater f ow, electric power supply, and so on. The controlled v ariable is thus v aried toward the predetermined v alue, the set point. Air and w ater are control media or control agents. Airf ow and w ater f ow are the manipulated v ariables. The equipment that v aries the output capacity by changing the opening position of the dampers and valves is called the process plant. There are tw o types of control loops: open and closed. An open-loop system assumes a f xed relationship between the controlled v ariable and the input signal being recei ved. The sensed v ariable is not the controlled v ariable, so there is no feedback. An e xample of an open-loop system would be a ventilating fan that turns on when the outdoor temperature e xceeds a specif ed set point. The sensed variable is the outdoor air temperature, and the controlled variable is the state of the fan (on or off). A closed-loop system depends on sensing the controlled v ariable to v ary the controller output and modulate the controlled de vice. In Fig. 5.2, temperature sensor T2 senses the controlled variable, the recirculating air temperature entering the air handler Tru; the DDC unit controller receives this sensed signal input and produces an output according to the softw are stored to modulate the fan speed, or the position of the inlet v anes of the supply f an. As the supply fan speed changes, or inlet v anes open and close, the supply air v olume f ow rate v aries and the space temperature Tr and recirculating temperature Tru change accordingly. This change in Tru is sensed again by T2 and fed back to the controller for further modulation of f an speed or inlet v anes to maintain v alues of Tru that approach the set points. These components form a closed-loop system. Figure 5.3 shows a block diagram of this closed-loop system. It sho ws a secondary input to the controller, such as the outdoor air temperature To, which may reset the set point in the controller to provide better and more economical control. The disturbances that af fect the controlled v ariable are load variations and changes in the outdoor weather . After the controller senses the signal feedback, it sends a correcti ve signal to the controlled de vice based on the dif ference between the sensed controlled variable and the set point. Thus, Tru is under continuous comparison and correction. A control system or its component, control subsystem, used to control the controlled v ariable(s) in a conditioned space, or within a mechanical de vice or equipment, may contain only one control loop; or it may contain two or more loops.

Sequence of Operations The sequence of operations is a description of the sequential order of the functional operations that a control system is supposed to perform, which plans and guides the operation and control of an air conditioning system. For a single-zone variable-air-volume (VAV) cooling packaged system sho wn in Fig. 5.2, when cold air supply is required during full occupanc y, the sequence of operations of cooling mode is as follows:

5.6

CHAPTER FIVE

FIGURE 5.3 Block diagram for a closed-loop feedback control system.

●

●

●

●

●

The supply f an is started and stopped by the scheduling softw are stored in the DDC unit controller. Manual o verride is possible. When the time schedule puts this system in cooling occupy mode, the microprocessor-based controller goes through a short initiation period, such as a 2-min period. During this period, dampers are dri ven to fully open, minimum open, and fully closed positions. These determine the effective range of the economizer potentiometer range. The smoke detector in the return air or the lo w-temperature limit sensor of the mix ed air will stop the supply fan if necessary. The supply fan status (on or off) is determined by the pressure differential switch across the fan. When the initiation period is completed, the supply fan is turned on. The control system tends to maintain the recirculating temperature Tru at around the cold set point, and it uses the 100 percent all outdoor air free cooling economizer c ycle as the f rst-stage cooling. If the outdoor temperature To  Tru, the outdoor air damper is fully opened and the recirculating air damper closed. If the outdoor air temperature To  Tru, the outdoor air damper is closed to a minimum position to pro vide required outdoor v entilation air, and the recirculating damper is fully opened. When the initiation period is completed, the supply fan is turned on from its zero speed. If Tru, sensed by the temperature sensor T2, is at a v alue above the set point, Tru  Tc, set, also To  Tru, the speed of the supply f an is gradually increased by the v ariable-speed drive inverter, resulting in a higher supply v olume flow rate. When the fan speed is raised to its upper limit, the supply volume flow rate is then at its maximum v alue. A still higher space load further raises Tru , and it e xceeds the cold set point Tc, set, and if To  Tru, chilled water starts to flo w to the cooling coil to cool the air simultaneously with the outdoor air free cooling, in order to mak e Tru  Tc, set. Only when the outdoor air damper is fully open and the static pressure dif ference between the space air and outdoor air is greater than a preset value, such as ps – o  0.03 in. WC (7.5 Pa), will the relief fan be energized. Its speed is modulated to maintain ps – o  0.03 in. WC (7.5 Pa). When Tru drops, the chilled water f ow to the cooling coil is reduced f rst, and then the outdoor air free cooling, outdoor air volume f ow rate.

Refer to Sec. 21.2 for the details of the sequence of operations of single-zone VAV cooling systems. In the design and operation of an EMCS system, the necessary documentation includes the sequence of operation, control diagrams, specif cations, operation, and maintenance manual.

ENERGY MANAGEMENT AND CONTROL SYSTEMS

5.7

Control Methods According to the types of control signal and the dif ferent kinds of ener gy used to transmit the signals, as well as whether a softw are is used during control operation, control methods can be classif ed as direct digital, pneumatic, electric, and electronic. Analog and Digital. There are tw o types of control signals: analog and digital. An analog signal is in the form of a continuous v ariable. It often uses the magnitude of electric v oltage or pneumatic pressure to represent the air temperature. A digital signal is a series of on and of f pulses used to transmit information. A con ventional analog controller recei ves a continuous analog signal, such as a v oltage or a pneumatic signal, that is proportional to the magnitude of the sensed v ariable. The controller compares the signal received from the sensor to the desired v alue (i.e., the set point) and sends a signal to the actuator in proportion to the difference between the sensed value and the set point. A digital controller , or microprocessor -based controller , receives an electric signal from sensor(s). It converts the electric signal to digital pulses of dif ferent time interv als to represent the signals values. The microprocessor of the digital controller performs the mathematical operations and kno wledge processing on these v alues. The output from the microprocessor can be either in digital form to actuate relays or con verted to an analog signal (say , a voltage or a pneumatic pressure) to operate the actuator(s). Direct Digital Control (DDC). A control system using DDC in volves adopting a microprocessorbased digital controller to perform mathematical operations and kno wledge processing according to the predetermined control algorithms or computer programs. The key element of DDC compared to analog control is the software and hardware contained in the direct digital controller which e xpands the control functions tremendously and adopts recently developed control logic. ADDC unit usually has more precise sensors and uses the same type of controlled devices as other control methods. Figure 1.2 shows an energy management and control system using DDC for an air -handling unit in a typical f oor of the NBC Tower, and Fig. 5.2 sho ws an EMS with DDC for a single-zone VAV system. Pneumatic Control. In a control system using pneumatic control, compressed air is used to operate the sensors, controllers, and actuators and to transmit the signals. It consists of: a compressed air supply and distrib ution system, sensors, controllers, and actuators. Figure 5.4 sho ws a typical pneumatic control system. In Fig. 5.4, a f lter is used to remove the dust particles, including submicrometer-size particles, contained in the air . The function of the pressure-reducing v alve is to reduce the pressure of compressed air dischar ged from the air compressor to the required v alue in the main supply line. The discharged compressed air is usually at a gauge pressure of 18 to 25 psig (124 to 172 kP ag), and the pressure signal required to actuate the v alve or damper actuator is 3 to 13 psig (20.7 to 89.6 kPag). The advantages of pneumatic control are as follows: ●

●

●

●

The compressed air itself is inherently a proportional control signal. The cost of modulating actuators is low, especially for large valves and dampers. Pneumatic controls require less maintenance and have fewer problems. Pneumatic controls are explosionproof.

The disadv antages stem mainly from comparati vely fe wer control functions, the high cost of sophisticated pneumatic controllers, and the comparati vely higher f rst cost of a clean and dry , compressed air supply for small projects. Electric or Electronic Control. Both types of control use electric ener gy as the energy source for the sensors and controllers. A control system using electric control often of fers two-position on-off control, as sho wn in Fig. 5.1. Switches, relays, contactors, and electromechanical de vices are

5.8

CHAPTER FIVE

FIGURE 5.4 A typical pneumatic control system.

system components for electric control systems. They are generally used for lo w-cost, small, and simpler control systems. In addition to the switches and relays, a control system using electronic control has transistors, diodes, capacitors, and printed-circuit boards as system components. Electronic control systems al ways have more accurate sensors and solid-state controllers with sophisticated functions and can be easily interfaced with the building automation system. Electronic control has a faster response and more accurate processing of data than electric control systems. Electronic control systems cost more and need skilled personnel for maintenance and troubleshooting.

Comparison of Control Methods Because of the increasing demand for more complicated controllers to satisfy the needs of better indoor environmental control, satisfactory indoor air quality , improved energy saving, lower cost, and greater reliability, the recent trend is to use EMCS with direct digital control for more demanding and large projects. A modern DDC system consists of electronic sensors, microprocessor-based controllers incorporated with electronic components, and electronic or pneumatic actuators. It is often more ef fective and cheaper to operate a lar ge pneumatic actuator than to use a lar ge electronic actuator in a DDC system. An EMCS using DDC offers the following advantages: 1. Flexibility of pro viding required control functions and the ability to coordinate multiple functions directly from complicated softw are programs (it can e ven mimic a human e xpert within a certain knowledge domain and offer artif cial intelligence) 2. More precise and f aster-response control actions pro vided by the microprocessor -based controller 3. The possibility of using high-le vel self-checking and self-tuning system components, which increases the system reliability

ENERGY MANAGEMENT AND CONTROL SYSTEMS

5.9

One of the main disadv antages of DDC is a still-higher f rst cost. However, the cost per control action is comparati vely lower. As more and more DDC systems are installed, DDC will become familiar to us. The cost of the microprocessor-based controllers and other DDC system components will drop further in the future. In the 1990s, the trend in HV AC&R control for ne w and retro f t projects is to use EMCS with DDC systems except for small projects.

5.3 CONTROL MODES Control modes describe ho w the correcti ve action of the controller tak es place as well as its ef fect on the controlled v ariable. For applications in HV AC&R, control modes can be classi f ed as tw oposition, step, f oating, proportional, proportional-integral, and proportional-integral-derivative. Before we discuss the control modes in detail, the term lag, or time lag, should be introduced. According to ASHRAE terminology, lag is (1) the time delay required for the sensing element to reach equilibrium with the controlled v ariable; or (2) an y retardation of an output with respect to the causal input, including the delay , because of the transport of material or the propagation of a signal.

Two-Position Control In tw o-position control, the controller controls the f nal control element at one of tw o positions: maximum or minimum (except during the short period when it changes position). Examples of twoposition control include starting and stopping the motor of a f an, pump, or compressor and turning on or off an electric heater. Sometimes it is called on-off control. Figure 5.5 shows a two-position control mode for an electric heater installed in a branch duct. In the middle diagram, the ordinate indicates the controlled v ariable, the space temperature Tr; in the lower diagram, the ordinate denotes the output capacity of the f nal control element, the electric heater. If the controller turns on the electric heater when the sensor senses a space temperature 69.5°F (20.8 °C), and turns of f at 70.5 °F (21.4 °C), the result is a c yclic operation of the electric heater and the rise and f all of Tr toward the two positions of 69.5 °F (20.8°C) and 70.5°F (21.4°C). The thermal storage ef fect of the electric heater , the branch duct, the building envelope that surrounds the space, and the sensor itself will ha ve a time lag ef fect on Tr. The rise of Tr is a convex curve with an overshoot higher than 70.5°F (21.4°C), and the drop of Tr is a concave curve with an undershoot lower than 69.5°F (20.8°C). The difference between the two points, on and off, is called the differential. If the heating capacity of the electric heater that results from an on-and-of f cyclic operation is represented by Qe,t, and if the actual space heating load Q rh  12Q e,t , the slope of the rising Tr curve will then be greater than the slope of the f alling Tr curve. As a consequence, the “on” period will be shorter than the “off” period. Such a cyclic Tr curve is shown in the upper left corner of Fig. 5.5. If Q rh  12Q e,t , the condition will be reversed (see the upper right corner of Fig. 5.5). The two-position control varies the ratio of on and off periods to meet any variation in the space heating load. To reduce the o vershoot and undershoot of the controlled v ariable in a tw o-position control mode, a modi f cation called timed two-position contr ol has been de veloped. A small heating element attached to the temperature sensor is ener gized during on periods. This additional heating effect on the sensor shortens the on timing. During of f periods, the heating element is deener gized. The differential of two-position control, the overshoot, and the undershoot all result in a greater f uctuation of the controlled variable. A suitable differential is always desirable in two-position control in order to prevent very short cycling, which causes hunting, a phenomenon of short c ycling of the controlled v ariable. Two-position control is not suitable for precise control of the controlled variable, but it is often used for status control, such as opening or closing a damper, turning a small single piece of equipment on or off for capacity control, etc., and for lower-cost control systems.

5.10

CHAPTER FIVE

FIGURE 5.5 A typical two-position control mode.

Step Control and Modulating Control In step control, the controller operates the relays or switches in sequence to vary the output capacity of the process plant in steps or stages. The greater the deviation of the controlled v ariable from the set point and the f aster the rate of change of the controlled v ariable, the higher will be the output capacity of the process plant. In modulating control, the controller activates the control device continuously and thus the change of the output capacity gradually. During the earlier stage of step control, there must be a differential of controlled variable between two on or tw o of f points of a particular piece of equipment of the process plant. The result is a greater f uctuation of the controlled variable. In the DDC controller -activated new generation of step control, the software of the DDC unit determines ho w fast each capacity step is added or subtracted according to the deviation from the set point and the rate of change of the controlled variable. Figure 5.6 shows step control of the cooling capacity of refrigeration compressors using DDC to maintain the dischar ge air temperature Tdis within predetermined limits in a packaged unit. When the air economizer alone can no longer balance the DX coils load, and Tdis f oats to the upper limit of the control band, point 1, the f rst stage of cooling is ener gized. If the c ycling and the continuously energized f rst-stage cooling capacity still cannot balance the coils load, then Tdis continues to rise until it reaches point 2, which is 1 °F (0.56°C) higher than the upper limit of the control band. The f rst-stage cooling is then lock ed on, and the second stage is ener gized to c ycling. When Tdis rises to point 2, all the present cooling stages will be lock ed on, and an additional cooling stage is added for cycling. When Tdis drops to point 7, the current cycling cooling stage will be deenergized, and the one cooling stage ne xt to the deener gized stage will be c ycling. It al ways needs a time delay, say at least 4 min, to turn on or off again in order to prevent hunting and other damage.

ENERGY MANAGEMENT AND CONTROL SYSTEMS

5.11

FIGURE 5.6 DDC-activated discharge air temperature control for a VAV rooftop packaged system.

Step control is another kind of on-and-of f control ha ving smaller v arying capacity and thus a f uctuating controlled v ariable. It has been widely used for the capacity control of refrigeration compressors and the electric heaters.

Floating Control In a f oating control mode, the controller moves the control device by means of the actuator to ward the set point only when the control point is out of the dif ferential, or dead band, as sho wn in Fig. 5.7. A control point is the actual v alue of the controlled v ariable at a certain time instant. Figure 5.7 represents a duct static pressure control system using a f oating control mode by opening and closing the inlet v anes of a supply f an. The control device — inlet vanes — can be moved in either an opening or a closing direction depending on whether the control point is over the upper limit of the differential or under the lo wer limit. In Fig. 5.7, when the controlled variable, the duct static pressure ps, is above the upper limit of the differential, the controller then closes the inlet vanes. If ps is below the lower limit of the differential, the controller opens the inlet vanes. A f oating control mode is more suitable for control systems with a minimal lag between the sensor and the control medium. A control medium is the medium in which the controlled v ariable exists.

Proportional Control In a proportional control mode, the controller mo ves the controlled de vice to a position such that the change in its output capacity is proportional to the de viation of the controlled v ariable from the set point. The position of the controlled de vice is linearly proportional to the magnitude of the controlled variable. Figure 5.8 represents the control of the space temperature through the modulation of a two-way valve of a cooling coil using proportional control. In Fig. 5.8, the controlled variable is the space temperature Tr , and the controlled device is the valve.

5.12

CHAPTER FIVE

FIGURE 5.7 Floating control mode.

In a proportional control mode, the throttling range is the change in the controlled variable when the controller moves the controlled de vice from the position of maximum output to the position of minimum output. The controlled v ariable’s range of v alues that will mo ve the proportional controller through its operating range is called the proportional band. In a proportional control system, the throttling range is equal to the proportional band.

FIGURE 5.8 Proportional control mode.

ENERGY MANAGEMENT AND CONTROL SYSTEMS

5.13

The set point is the desired v alue of the controlled v ariable, or the desired control point that the controller seeks to achie ve. The dif ference between the control point and the set point is called the offset, or deviation. In proportional control, when the controlled variable is at the bottom line of the throttling range, the controller will position the actuator at the closed position. At the set point, the actuator will be at 50 percent of the open position. When the controlled variable is at the top of the throttling range, the actuator will be at 100 percent of the open position. In a proportional control mode, since the output signal V of the controller is proportional to the deviation of the control point from the set point, their relationship can be expressed as: V  Kpe M

(5.1)

where Kp  proportional gain, proportional to 1 /(throttling range) e  error signal, i.e., deviation or offset M  output value when deviation is zero (usually, output value at middle of output range of controller) For space and dischar ge air temperature control using proportional control, the of fset is directly proportional to the space load and the coil load. The space or discharge temperature T, °F (°C), can be calculated as: T  R loadTt,r Tmin

(5.2)

where Tt,r  throttling range, expressed in terms of space or discharge temperature, °F (°C) Tmin  space and discharge temperature when space load or coil load is zero, °F (°C) In Eq. (5.2), Rload represents the load ratio of the space load or coil load, which is dimensionless and can be evaluated as R load 

actual load design load

(5.3)

The actual load and design load must be in the same units, such as Btu/h (W). For a proportional control mode, a certain de gree of of fset, or deviation, is inherent. The position of the controlled de vice is a function of the of fset. Only when of fset exists will the controller position the actuator and the v alve or damper at a position greater or smaller than 50 percent open. The throttling range is primarily determined by the HV AC&R system characteristics and cannot be changed after the system is designed. A proportional control mode is suitable for an HVAC&R system that has a lar ge thermal capacitance, resulting in a slow response and stable system and allo wing comparatively narrower throttling range and thus a smaller of fset. Fast-reacting systems need a large throttling range to avoid instability and short cycling, or hunting. Proportional plus Integral (PI) Control In a proportional plus integral control mode, a second component, the integral term, is added to the proportional action to eliminate the offset. The output of the controller can thus be expressed as V  K pe K ie dt M

(5.4) where Ki  integral gain and t  time. In Eq. (5.4), the second term on the right-hand side (the inte gral term) indicates that (1) the error, or offset, is measured at regular time intervals and (2) the product of the sum of these measurements and Ki is added to the output of the controller to eliminate the of fset. The longer the of fset exists, the greater the response of the controller . Such a control action is equi valent to resetting the set point in order to increase the controller output to eliminate the of fset. As such, PI control is sometimes called proportional plus reset control. Figure 5.9 a shows the v ariations of a controlled

5.14

CHAPTER FIVE

FIGURE 5.9 Proportional-integral and proportional-inte gral-derivative control modes. (a) Proportional-integral; (b) proportional-integral-derivative.

variable for a proportional-integral control mode. In PI control, proportional-integral control signals are additive. For PI control, the controlled variable does not have any offset once it has achieved a stable condition, except due to an y inaccuracy of the instrumentation and measurement. Proper selection of the proportional gain Kp and integral gain Ki as well as proper tuning is important for system stability and control accurac y. PI control may be applied to f ast-acting control systems with a greater throttling range setting at the controller for better system stability , e.g., discharge air temperature, discharge chilled water temperature, and duct static pressure control systems.

Proportional-Integral-Derivative (PID) Control A PID control mode has additional control action added to the PI controller: a derivative function that opposes to any change and is proportional to the rate of change. The output of such a controller can be described by the following equation: V  K pe K i e dt K

de M dt

(5.5)

ENERGY MANAGEMENT AND CONTROL SYSTEMS

5.15

where K is the deri vative gain. The ef fect of adding the deri vative function K de /dt is that the quicker the control point changes, the greater the corrective action provided by the derivative function. Figure 5.9b shows the variation of the controlled variable for a PID control mode. As with a PI control mode, PID control mode also has no offset once the controlled variable has reached a stable condition, except due to instrument inaccuracy. Compared to the PI control mode, the PID control mode, which is a combination of proportional, integral, and derivative actions, exhibits faster corrective action and a smaller o vershoot and undershoot following an of fset and a change of the controlled v ariable. The controlled v ariable is brought to the required set point in a shorter time. Ho wever, it is more dif f cult to determine properly three constants, or gains (Kp, Ki, and K).

Compensation Control, or Reset Compensation control, or reset, is a type of control mode in which a compensation sensor is generally used to reset a main sensor to compensate for a v ariable change sensed by the compensation sensor. The purpose is to achieve operation that is more effective, energy-eff cient, or both. In the design of a reset mode, the f rst things to decide are the control point at which the main sensor will be reset and the v ariable to be sensed by the compensation sensor . The main sensor , which senses the mix ed air temperature in the air -handling unit, is usually reset by a compensation sensor that senses the outdoor temperature, as shown in Fig. 5.10. Another task to decide on is the relationship between the v ariables sensed by the main and compensation sensors — the reset schedule. Man y reset schedules ha ve a different relationship between these tw o sensors at v arious stages. F or e xample, in Fig. 5.10, when the outdoor temperature To is less than 30°F ( 1.1°C), it is within the range of stage I. No matter what the magnitude of To is, the set point of the mix ed temperature Tm is 65°F (18.3°C). When 30°F  To  95°F ( 1.1°C  To  35°C), it is in stage II. The linear relationship between the temperatures sensed by the main and compensation sensors in Fig. 5.10 can be expressed by the following equation: Tm  65

65 55 95 30

(To 30)

(5.6)

When To  95°F (35°C), the set point of the mixed temperature is always 55°F (12.8°C). Compensation control modes ha ve been widely adopted to reset space temperature, discharged air temperatures from air -handling units or packaged units, or water discharge temperatures from central plants.

Applications of Various Control Modes Selection of a suitable control mode depends on ●

●

●

Operating characteristics Process or system characteristics, such as whether the thermal capacitance should be tak en into consideration Characteristics of load changes

If a simpler control mode can meet the requirements (say , two-position control versus PID control), the simpler control mode is always the f rst choice. Except for two-position and step control, all the other control modes are modulation controls. Modulation control is a control mode that is capable of increasing or decreasing a v ariable according to the de viation from the required v alue in small increments continuously. Also note that there is a signi f cant difference between the tw o-position control, step control, f oating control, and proportional control modes with an of fset and the of fset free PI and PID controls.

5.16

CHAPTER FIVE

FIGURE 5.10 Reset control and schedule.

For HVAC&R processes, PI control can satisfy most of the requirements, and it is most widely used. Microprocessor-based PID control is a v ery powerful tool. PID control mode is more appropriate for f ast-acting duct static pressure control and air f ow control, and it is recommended to set the controller with a lar ge proportional band for control system stability , a slow reset to eliminate deviation, and a derivative action to provide a quick response.

5.4 SENSORS AND TRANSDUCERS A sensor is a de vice that acts as a component in a control system to detect and measure the controlled variable and to send a signal to the controller. A sensor consists of a sensing element and accessories, as shown in Fig. 5.11. The term sensing element often refers to that part of the sensor that actually senses the controlled variable. In HVAC&R systems, the most widely used sensors are temperature sensors, humidity sensors, pressure sensors, and f ow sensors. Recently , CO2 sensors, air quality sensors, and occupancy sensors are being used in man y new and retro f t projects. Electronic sensors that send electric signals to electronic controllers, can also be used for the DDC units, and the current trend is to use solid-state miniaturizing sensing elements. Electric output from the sensors is usually expressed in 0 to 10 V dc or 4 to 20 mA. In the selection of sensors, accuracy, sensitivity of response, reliability, long-term stability or drift, required calibration interv als, maintainability, and especially the possibility of contamination by dust particles due to contact with the control medium should be considered. Drift is a kind of of fset. ASHRAE’s Terminology def nes drift as “change in an output-input relationship o ver time with the change unrelated to input, environment, or load.” A calibration period of 1 year and longer is often considered acceptable. Location of the sensor af fects the sensor output directly. A sensor should be located in the critical area where the controlled v ariable needs to be maintained within required limits. The sensing element of an air sensor should be well e xposed to the air , so that air can f ow through the sensor without obstruction. Air sensors should not be af fected directly by the supply airstream nor should space air sensors be af fected by the outdoor airstream. Sensors should be shielded from radiation

ENERGY MANAGEMENT AND CONTROL SYSTEMS

FIGURE 5.11 (a) A temperature sensor and (b) a typical pitot tube f ow-measuring station.

5.17

5.18

CHAPTER FIVE

and mounted f rmly on a structural member or duct w all, free from vibration. If a space air sensor must be located on a concrete column, thermal insulation should be pro vided between the sensor and the column.

Temperature Sensors Temperature sensors f all into tw o categories: those that produce mechanical signals and those that emit electric signals. Bimetal and rod-and-tube sensors that use sensing elements to produce a mechanical displacement during a sensed temperature change either open or close an electric circuit in an electric control system, as sho wn in Fig. 5.1, or adjust the throttling pressure by means of a bleeding nozzle in a pneumatic control system. F or a sealed bellow sensor, a change in temperature causes a change in the pressure of a liquid in a remote bulb. The expansion and contraction of vapor then move the mechanism of the controller. Temperature sensors that produce electric signals, as shown in Fig. 5.11 a, are the same as the sensors for temperature measurement and indication mentioned in Sec. 2.3. In addition to resistance temperature detectors (R TDs) and thermistors, sensors sometimes use thermocouples. A thermocouple uses wires of two dissimilar metals, such as copper and constantan, or iron and nickel, connected at two junctions, to generate an electromotive force between the junctions that is directly proportional to the temperature dif ference between them. One of the junctions is kept at constant temperature and is called the cold junction. Various systems have been developed to maintain the cold junction at a constant temperature and to pro vide compensation if the cold junction is not at 32 °F (0 °C). This task mak es the use of thermocouples more complicated and expensive. The electromotive force produced between the tw o junctions can be used as the signal input to a controller. Bimetal and rod-and-tube temperature sensors are simple and lo w in cost. However, they cannot provide temperature indication and electric signals for DDC and are often used in electric control systems. Platinum and nickel RTDs are stable, reliable, and accurate. They are very expensive compared to thermistors and need calibration to compensate for the ef fects of ha ving external leading wires. R TDs are widely adopted in DDC for commercial applications. F or a project that needs precise temperature control, RTD is often the choice. High-quality thermistors can pro vide stable, reliable, and interchangeable temperature sensors, and are also widely used in man y commercial applications. Petze (1986) reported that some high-quality thermistors e xhibited better than  0.002°F ( 0.001°C) stability for a 2-year period. A typical space air temperature sensor has the following characteristics: Sensing element

Operating range

Platinum f lm element, 3000 Positive temperature coeff cient, 4.8 /°F (8.64 /°C) 60 to 90°F (15.6 to 32.2°C)

Humidity Sensors As mentioned in Sec. 2.8, humidity sensors f all into tw o cate gories: mechanical and electronic. When the same humidity sensor is used for both monitoring and DDC, the ion-exchange resistancetype, also called the bulk polymer resistance-type, and capacitance-type humidity sensors are often adopted. Humidity sensors usually have different accuracy at very low, 20 to 80 percent mid-range, and high relati ve humidities. The bulk polymer resistance humidity sensor is not accurate at lo w relative humidities and generally pro vides stable and accurate readings within a range of 30 to 90 percent relative humidity. Its performance is affected by the air contamination. Capacitance humidity sensors are accurate at 10 to 80 percent relati ve humidity. However, they become unstable at high relative humidities.

ENERGY MANAGEMENT AND CONTROL SYSTEMS

5.19

In an EMCS, an electronic humidity sensing element measures the relati ve humidity and sends an electric signal to the DDC unit. A typical space air humidity sensor has the follo wing specif cations: Power source Humidity range (active) Nominal output range Accuracy at 70°F (21.1°C) Speed of response Mounting

24 V ac 10 to 80 percent 0 to 5 V dc  3 percent at 10 to 60 percent range  4 percent at 60 to 80 percent range 8 min (90 percent of response time) Wall mount

Pressure Sensors A pressure sensor usually senses the dif ference in pressure between the controlled medium (air and water) and a reference pressure; or the pressure dif ferential across two points, such as the pressure differential across a f lter. The reference pressure may be an absolute v acuum, atmospheric pressure, or the pressure at any adjacent point. Output signals from the pressure sensors may be electric or pneumatic, and analog or binary. Pressure sensors used for HVAC&R systems can be divided into high-pressure and low-pressure sensors. High-pressure sensors measure in pound per square inch or feet WC (kilo pascals), and low-pressure sensors measure in inches WC (pascals). The sensing elements for high-pressure sensors are usually Bourdon tube, bellows, and sometimes diaphragms. F or low-pressure air sensors, large diaphragms or f exible metal bellows are usually used. To measure the duct static pressure, if a long section of straight duct of a length greater than 10 ft (3 m) is a vailable, a single-point, pitot-tube type of duct static pressure sensor can be used. Otherwise, a multipoint pitot-tube array or f ow-measuring station with airf ow straighteners should be used. The small hole used to measure static pressure should ne ver be directly opposite to an airstream with a velocity pressure that can affect the reading. A reference pressure should be picked up at a point with lo w air velocity outside the duct, at a point serv ed by the same air system, or in the ceiling plenum. In a DDC system, low air pressure is often sensed by measuring the capacitance of tw o diaphragms; one is allowed to move toward or away from the f xed one, depending on the pressure on two sides of the diaphragm. An electric circuit is used to con vert the capacitance to a v oltage or milliampere signal. The other method used to sense the lo w air pressure in DDC systems w orks because the air pressure dif ferential compresses or stretches a diaphragm as well as a strain gauge. The change of the resistance of the gauge is detected and ampli f ed, and this electric signal in the form of voltage or milliamperes is sent to the controller. A space air pressure sensor having an operating range of 0.1 to 0.1 in WC ( 25 Pa to 25 Pa) detects the resistance of a silicon diaphragm, and thus the space air pressure is then measured. Stainless steel, rubber, etc., can also be used as the material of the diaphragms. A space air pressure sensor should be located in an open area of the conditioned space where the air v elocity is less than 40 fpm (0.2 m /s) and where its reading is not af fected by the opening of the doors. The reference pressure pickup is best located at the rooftop, at a level 10 ft (3 m) abo ve the building to avoid the inf uence of wind.

Flow Sensors Flow sensors usually sense the rate of air f ow and w ater f ow in cfm (L /s) for air and gpm (L /s) for w ater. For air f ow sensors, the average velocity pressure pv, in. WC (Pa), is often sensed and

5.20

CHAPTER FIVE

measured. The average air velocity va, fpm, is then calculated as va  4005K √pv

(5.7)

where K is the f ow coeff cient, which depends on the type of pitot-tube array used and the dimension of the round or rectangular duct. After that, the volume f ow rate can easily be determined as the product va A. Here, A represents the cross-sectional area of the air passage, perpendicular to the airf ow, in ft 2 (m2). Various forms of pitot-tube array ha ve been de veloped and tested to determine the average velocity pressure of a rectangular or circular duct section by measuring the dif ference between the total and static pressures of the airstream. A typical pitot-tube f ow-measuring station with f ow straighteners to provide more even airf ow is shown in Fig. 5.11. Electronic air v elocity sensors such as hot-wire anemometers and thermistors ha ve also been widely adopted to measure airf ow, especially for variable-air-volume (VAV) boxes. Thermistors are cheaper than hot-wire anemometers. Heated thermistors need periodic calibration. Water f ow sensors of the dif ferential-pressure type, such as ori f ce plates, f owing nozzles, and pitot tubes, have only a limited measurement range. Turbine or magnet-type f owmeters can apply to a wider range, but they are more expensive and also need periodic calibration.

Carbon Dioxide and Air Quality Sensors A CO 2 sensor detects and indicates the amount of carbon dioxide (CO 2 ) contained inside the air , which is a reliable indication of the body odor released by the occupants of the conditioned space. Because of a certain relationship that e xists between the CO 2 contained in the outdoor air and the CO2 released by the occupants, the concentration of CO 2 in space air may sometimes be used as a rough indication of amount of outdoor v entilation air supply to the conditioned space under the specif c conditions. The sensing process used in CO 2 sensors includes potentiometric and amperometric electrochemical cells, an infrared detector, etc. A CO 2 sensor usually has an operating range of 0 to 3000 ppm. F or measurements within an accurac y of  100 ppm, recalibration may be required on the order of once per year. An air quality sensor , also called a v olatile organic compound (VOC) or a mix ed-gas sensor, is used to monitor and detect the relati ve concentrations of VOC or mixed gas, or total concentration including acetone, ammonia, CO, CO2, SO2, chlorine, formaldehyde, CFC-11, CFC-12, etc. The concentration of the contaminant is often e xpressed in tested and commissioned units, such as 0 to 5 or 0 to 10 units. The sensing element has a tin dioxide surf ace which is heated to a temperature abo ve 130 °F (54°C). The change of the conductivity and thus its resistance are then amplif ed and fed to the controller in terms of 0 to 10 V dc in proportion to the contamination. Air quality sensors are less expensive than CO 2 sensors and need less maintenance. The drift of an air quality sensor is unpredictable. Both CO 2 sensors and air quality sensors are no w used for demand-based v entilation control to provide the required amount of outdoor air for acceptable indoor air quality.

Occupancy Sensors An occupancy sensor detects whether a room is occupied by occupant(s). As a result, the HVAC&R and lighting in this room can be turned off when the room is not occupied, to save energy. There are two types of occupanc y sensors: ultrasonic and infrared. An ultrasonic occupancy sensor sends out a rather lo wer le vel of ultrasonic w ave and detects mo vement of the occupant when there are changes in recei ving patterns. Ultrasonic sensors sweep o ver the area that surrounds the sensor . False signals may be received due to the motion of the papers or air mo vement from a diffuser. The sensitivity of the ultrasonic sensor should be adjusted to avoid these problems.

ENERGY MANAGEMENT AND CONTROL SYSTEMS

5.21

An infrared occupanc y sensor senses the mo vement of the occupant(s) as the sensor recei ves heat from the occupant when that person is mo ving. Infrared sensors need to see the occupant. As long as there is no obstruction between the sensor and the occupant in the room, the sensing process of an infrared sensor is ef fective. Occupancy sensors are often mounted on the ceiling or at a high level on a w all. They are used in the guest rooms of the hotels and motels and other commercial buildings. Ultrasonic occupancy sensors show a wider acceptance than infrared occupanc y sensors.

Wireless Zone Sensors and Intelligent Network Sensors In an HVAC&R system, a wireless zone sensor is an indoor spread spectrum radio-frequenc y transmitter which sends room temperature and other status information to a local recei ver located no more than 1000 ft (300 m) a way. A translator converts these data and sends them to a v ariable-airvolume DDC unit controller through a wired communications link. This type of ne w technology was de veloped recently for the sak e of pro viding f exibility for the rearrangements of the of f ce layout. Kovacs (1996) noted that intelligent netw ork sensors “. . . can linearize the sensor signal, accept an offset adjustment through the netw ork, may have alarm- or decision-making algorithms on board, and include self-test diagnostics to continuously v alidate performance.” An intelligent network sensor may have a sensing element, an analog /digital transducer, a neutron chip — an onboard microprocessor, and a transceiver. Dew-point temperature, enthalpy, wet-bulb temperature, etc. can be pro vided as a combined v ariable. An intelligent netw ork sensor will help to dri ve the system structure towards becoming a more cost-ef fective distributed control system in the future, a system that has more locally processed signal and data.

Transducers or Transmitters A transducer is a device that converts energy from one form to another or ampli f es an input or output signal. In HVAC&R control systems, a transducer may be used to convert an electric signal to a pneumatic signal (E /P transducer), e.g., a pneumatic proportional relay that v aries its branch air pressure from 3 to 15 psig (20 to 103 kP ag) in direct proportion to changes in the electrical input from 2 to 10 V. Also E /P transducers are used between microprocessor -based or electronic controllers and pneumatic actuators. However, a pneumatic signal can be converted to an electric signal in a P /E transducer. For example, a P /E relay closes a contact when the air pressure falls and opens the contact when the air pressure rises abo ve a predetermined v alue. A transmitter is used to transmit a signal, pneumatic or electric; through air, water, or other f uids. A sensor is also a transmitter. However, the difference between them is that a sensor only senses the signal of the controlled v ariable and transmits it.

5.5 CONTROLLERS A controller receives input from the sensor, compares with the set point or implemention based on its stored computer software, sends an output to or modulates the control device for maintaining a desirable indoor en vironment. A thermostat is a combination of a temperature sensor and a temperature controller, whereas a humidistat is a combination of a humidity sensor and a humidity controller.

Direct-Acting or Reverse-Acting A direct-acting controller increases its output signal upon an increase in the sensed controlled v ariable, and it decreases its output signal upon a decrease in the sensed controlled v ariable. Conversely,

5.22

CHAPTER FIVE

a re verse-acting controller decreases its output signal upon an increase in the sensed controlled variable and increases its output signal upon a decrease in the controlled variable.

Normally Closed or Normally Open A controlled device, a valve or a damper, that is said to be in the normally closed (NC) position indicates that when the input signal to the controller f alls to zero or below a critical value, the control device will be in the closed position. The closing of the v alve or damper is most lik ely due to the action of a spring or a supplementary po wer supply. A valve or a damper that is said to be normally open (NO) indicates that when the input signal to the controller f alls to zero or belo w a critical value, then the valve, damper, or associated process plant will be in the open position.

Pneumatic Controllers In a pneumatic controller , the basic mechanism used to control the air pressure in the branch line supplied to the actuators is a nozzle- f apper assembly plus a restrictor . Figure 5.12 sho ws a pneumatic controller with such a mechanism. Compressed air supplied from the main line f ows through the restrictor and dischar ges at the opening between the nozzle and the f apper, which has a spring pulling it do wnward. The nozzle and restrictor are sized in such a manner that when the f apper moves a way from the nozzle, all the air escapes from the nozzle and the branch line pressure is zero. When the f apper covers the nozzle or if there is no air f ow in the branch line, the pressure of the branch line is equal to that in the main line. If a sensor, such as a bimetal sensor , moves the f apper upward according to the magnitude of the controlled v ariable sensed by the sensor , then the input signal from the sensor determines the opening between the nozzle and f apper and hence the compressed air pressure in the branch line.

FIGURE 5.12 A pneumatic controller with a nozzleand a restrictor.

f apper assembly

ENERGY MANAGEMENT AND CONTROL SYSTEMS

5.23

For a direct-acting pneumatic temperature controller used to control the space temperature during summer, the branch line pressure may change from 9 to 13 psig (62 to 89 kP ag) when the space temperature increases from 73 to 77 °F (22.8 to 25 °C). This nozzle-f apper assembly pneumatic controller operates in proportional control mode. Man y other more complicated pneumatic controllers have been developed to perform other control modes and additional functions.

Electric and Electronic Controllers An electric controller uses switches, relays, and a bridge circuit formed by potentiometers to position the actuators in on-of f, f oating, and proportional control modes according to the input signal from the sensor and the predetermined set point. An electronic controller can pro vide f ar more functions than electric controllers can. It may recei ve input signals from both the main sensor and the compensation sensor with amplif cation and combination. In the control circuit, an electronic controller basically pro vides proportional or proportionalintegral control modes. The output signal from the controller can be used to position an actuator or to provide the sequencing of actuators, or to change to tw o-position, f oating, or even PID control modes in conjunction with additional circuits.

Direct Digital Controllers A direct digital controller has a microprocessor to implement computer programs to provide various control functions. In DDC units, there are analog-to-digital (A /D) and digital-to-analog (D /A) converters to con vert analog input to digital signals for processing, or to con vert digital signals to analog for actuators, if necessary. DDC units are stand-alone and microprocessor -based controllers. Stand-alone means that the controller has suf f cient capacity to e xecute the assigned control functions alone. Today, there are mainly tw o types of DDC units: system controllers and unit controllers. System Controllers. A system controller , also called a stand-alone panel (SAP), as sho wn in Fig. 5.13, has the ability to coordinate communications between system controllers, between the system controller and the personal computer (PC) in the w orkstation, and between the system controller and the supported unit controllers. It also has the ability to pro vide and e xecute the control programs for functional control, and to store user databases and trend log v alues. A system controller can support 50 to 200 unit controllers on separate unit controller trunk(s). Unit Controllers. A unit controller, also called a terminal controller, is shown in Fig. 5.14. It usually has limited capacity to e xecute f actory-loaded computer programs and to pro vide functional control for a terminal or a piece of HV AC&R equipment. Unit controllers are often connected in a separate netw ork and supported by a system controller . The new-generation unit controllers ha ve greater memory to handle complicated control programs, and they provide time and calendar scheduling, data storage, and other functions, such as limited programming. Hardware. Many system controllers have only a single printed-circuit board. Single-board conf guration often of fers lo wer f rst cost. Its disadv antage is that an y component f ailure requires the replacement of the complete board. Another approach is that the controller is made from v arious modules. The modular approach isolates the component f ailures and plugs on the single-module replacement quickly and inexpensively. Memory. ●

The types of memory included in DDC units are as follows:

Read-only memory (R OM), which stores the softw are provided by the manuf acturer and should not be modif ed by the user.

5.24

CHAPTER FIVE

Universal analog inputs;

Independent power supplies for each electronic module;

Tool-less installation;

Fully integrated output relays, overrides and transducers;

Wireway space

Onboard diagnostic displays and indications.

Context-sensitive touchpad; user interface.

FIGURE 5.13 A typical system controller. (Source: Johnson Controls. Reprinted by permission.)

●

●

Random-access memory (RAM), which stores custom control softw are developed during installation or prepared by the user. This type of memory is volatile (i.e., it can be read from and written to) and requires battery backup. Electric erasable programmable read-only memory (EEPROM), which stores custom control software and is v olatile. The adv antage of EEPR OM o ver RAM is that EEPR OM does not need

ENERGY MANAGEMENT AND CONTROL SYSTEMS

5.25

FIGURE 5.14 A typical unit controller . (Source: Honeywell Inc. Reprinted by permission.)

●

battery backup. Ho wever, EEPROM cannot be used for as man y writing, erasing, and rewriting cycles as RAM. Flash erasable programmable read-only memory (f ash EPROM), which is a kind of new memory technology and allo ws the stored control softw are to remain untouchable inde f nitely without power.

System controllers often use a 16-bit microprocessor. A typical system controller has the following memories: RAM: ROM:

256 Kbytes 128 Kbytes EEPROM 512 Kbytes f ash EPROM

Input/Output (I /O). There are four kinds of I /O: analog input (AI); digital or binary input (BI); analog output (AO); and pulsed or binary output (BO). Con ventionally, a sensor input, a controller output, or a control v alue, is referred to as a point or object. Current system controllers allo w their

5.26

CHAPTER FIVE

input and output connections to be con f gured with great f exibility. Each input or output point can be either analog or digital. Typical analog inputs (electric signal) are 0 to 10 V dc or 4 to 20 mA. A system controller often has a total of 20 to 50 I /O points. Some system controllers can be e xtended to 100 points if necessary. A typical system controller has the following I/O points capacity: Analog/digital inputs Universal analog/digital input/outputs Digital outputs Totalizer inputs, i.e., pulsed inputs

18 6 12 4

A unit controller or a terminal controller usually has 4 to 20 I /O points. The I /O points in a typical unit controller may take the following forms: Analog inputs Digital inputs Microbridge sensor Analog outputs Digital outputs Pneumatic Triac

0 to 10 V dc Switch, relay, and transistor 0 to 3 in. WC pressure differential (0 to 750 Pa) 0 to 10 V dc, 4 to 20 mA 30 V ac 3 to 15 psig (20.6 to 103 kPag) On/off output for electric heater, fan motor, etc.

5.6 WATER CONTROL VALVES AND VALVE ACTUATORS Water valves are used to re gulate or stop water f ow in a pipe either manually or by means of automatic control systems. Water control valves adopted in water systems can modulate water f ow rates by means of automatic control systems. Valve Actuators An actuator, sometimes called an operator, is a device which receives an electric or pneumatic analog control signal from the controller, either directly or through a digital-to-analog converter. It then closes or opens a v alve or damper , modulating the associated process plant, and causes the controlled variable to change to ward its set point. Valve actuators are used to position control v alves. They are mainly of the following types: Solenoid Actuators. These use a magnetic coil to mo ve a mo vable plunger connected with the valve stem. Most solenoid v alve actuators operate at tw o positions (on and of f). They are used mainly for small valves. Electric Actuators. These move the v alve stem by means of a gear train and linkage. Dif electric motor valve actuators can be classif ed according to the control mode they use:

ferent

1. On /off mode. For this type of actuator, the motor moves the valve in one direction, and when the electric circuit breaks, the spring returns the valve stem to the top position (either open or closed position depending on whether it is a normally open or closed valve). 2. Modulating mode. The motor can rotate in both directions, with spring return when the electric circuit breaks. 3. Modulating mode with supplementary power supply . The motor rotates in tw o directions and without a spring-return arrangement. When the power is cut of f, a bypass signal is usually sent

ENERGY MANAGEMENT AND CONTROL SYSTEMS

5.27

to the electric motor to dri ve the valve to its open or closed position, depending on whether it is a normally open or closed valve. It may tak e minutes to fully open a lar ge valve using an electric motor v alve actuator. Modern electronic actuators use solid-state control boards to determine the speed, the action, and other functions to meet more demanding requirements. Pneumatic Actuators. A pneumatic valve actuator consists of an actuator chamber whose bottom is made of a f exible diaphragm or bellows connected with the valve stem. When the air pressure in the actuator chamber increases, the downward force overcomes the spring compression and pushes the diaphragm downward, closing the valve. As the air pressure in the actuator chamber decreases, the spring compresses the diaphragm, moving the valve stem and valve upward. A pneumatic valve actuator is powerful, simple, and fast to respond. Because of the increasing popularity of the DDC systems, there is an increasing demand for electric actuators that can be interfaced with a DDC system.

Types of Control Valves Water control v alves consist mainly of a v alve body, one or tw o valve disks or plugs, one or tw o valve seats, a valve stem, and a seal packing. Based on their structure, water control valves can be classif ed into the following types: 1. Single-seated. A single-seated valve has only a single valve disk and seat, as shown in Fig. 5.15b and c. It is usually used for water systems that need a tight shutoff. 2. Double-seated. A double-seated v alve has tw o v alve disks connected to the same v alve stem and is designed so that the f uid pressure e xerted on the v alve disks is al ways balanced. Consequently, less force is required for the operation of a double-seated v alve, as sho wn in Fig. 5.15a. 3. Butterf y. A butterf y valve consists of a c ylindrical body, a shaft, and a disk that rotates on an axis, as shown in Fig. 5.15 d. When the v alve closes, it seats against a ring inside the body . A butterf y valve exhibits low f ow resistance when it is fully opened. It is compact and is usually used in large water pipes. According to the pattern of the w ater f ow, water control v alves can again be classi f ed as tw oway valves or three-way valves. A two-way valve has one inlet port and one outlet port. Water f ows straight through the two-way valve along a single passage, as shown in Fig. 5.15a. In a three-w ay valve, there are three ports: two inlet ports and one common outlet port for a three-way mixing valve, as shown in Fig. 5.15b, and one common inlet port and two outlet ports for a three-way diverting valve, as shown in Fig. 5.15 c. In a three-w ay mixing v alve, the main w ater stream f ows through the coil or boiler , and the bypass stream mix es with the main stream in the common mixing outlet port. In a three-w ay diverting valve, the supply w ater stream di vides into two streams in the common inlet port. The main water stream f ows through the coil, and the bypass stream mixes with the main water stream after the coil. A three-way mixing valve is always located downstream of the coil. But a diverting valve is always located upstream of the coil. A diverting valve should never be used as a mixing valve. The unbalance pressure difference between the two inlet ports and the outlet port at a closed position may cause disk bouncing and v alve wear when the valve disk travels between the two extremes.

Valve Characteristics and Ratings The different types of control v alves and the characteristics that are important during the selection of these valves are as follows:

5.28

CHAPTER FIVE

FIGURE 5.15 Various types of control v alves: (a) Double-seated tw o-way valve; (b) single-seated three-w ay mixing v alve; (c) singleseated three-way diverting valve; (d) butterf y valve.

1. Equal-percentage valve. This control v alve changes the w ater f ow rate by a certain percentage for that same percentage of lift in the v alve stem when the upstream v ersus downstream water pressure difference across the valve (its pressure drop) is constant. 2. Linear valve. This control valve shows a directly proportional relationship between the f ow rate and the lifting of the valve stem for a constant pressure drop.

ENERGY MANAGEMENT AND CONTROL SYSTEMS

5.29

3. Quick-opening valve. This control v alve gives the maximum possible f ow rate when the v alve disk or plug is just lifted from its seat. Rangeability is def ned as the ratio of the maximum f ow rate to the minimum f ow rate under control. An equal-percentage valve may have a very good rangeability of 50 :1. A linear valve may have a rangeability of 30 :1. The f ow characteristics of equal-percentage, linear, and quick-open v alves are sho wn in Fig. 5.16. The following control v alve ratings should be considered during the selection and sizing of a valve: 1. Body r ating. The nominal body rating of the v alve is the theoretical rating of the v alve body only, in psig. The actual body rating is the permissible safe w ater pressure for the valve body, in psig (kPag), at a specif c water temperature. 2. Close-off rating. That is the maximum pressure difference between the inlet and outlet ports that a valve can withstand without leakage when the valve is fully closed, in psi (kPa). 3. Maximum pressure and temperature. These are the maximum pressure and temperature of w ater that the whole valve, including body, disk, seat, packing, etc., can withstand.

Valve Selection Proper selection of w ater control v alves depends on w ater system performance, load variations, pipe size, control modes, etc. Today, the use of scaling f actors on analog outputs in a DDC system permits a nonlinear de vice to pro vide an output of linear response. Ho wever, select a control v alve having a linear relationship between a change in the controlled v ariable and the amount of tra vel of the v alve stem, or a linear system control characteristic over the operating range is still desirable when it is costeffective. Hence, a linear valve is often used for the water system for which the controlled variable has

FIGURE 5.16 Typical f ow characteristics of v arious types of control valve.

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a linear relationship with the w ater f ow or v alve opening; or for applications that do not ha ve wide load variations. When a control v alve is used to modulate the w ater f ow rate of a hot or chilled w ater coil, a large reduction in the f ow rate causes only a small reduction in the heating or cooling output of the coil. Given such circumstances, the nonlinear behavior of an equal-percentage valve combined with the nonlinear output performance of a hot or chilled w ater coil will pro vide more linear system behavior. When three-way valves are used, the water f ow rate before or after the common port is approximately constant, no matter how wide the openings of the v arious ports in the three-w ay valves. As such, three-way valves are used in constant or approximately constant water f ow rate systems, even the coil load changes. As a tw o-way valve closes, the f ow rate of the w ater system decreases and its pressure drop across the two-way valve increases. A two-way valve is thus used for w ater systems that have variable volume f ow during a variation in system load, as shown in Fig. 5.17.

Valve Sizing The size of a control v alve affects the controllability of a w ater system. If a control v alve is o versized, then the smallest increment possible may o vershoot the controlled v ariable. An undersized control valve needs great pumping po wer. The size of a control v alve is also closely related to its

FIGURE 5.17 A typical chilled water system.

ENERGY MANAGEMENT AND CONTROL SYSTEMS

5.31

design water f ow rate V˙ , gpm (L /s), and the pressure drop across the v alve pvv, psi (kPa), when the control valve is fully opened. Their relationship can be expressed as V˙  Cv√pvv Cv 

V˙

(5.8)

√pvv

where Cv is the f ow coeff cient for a f ow rate of 1 gpm at a pressure drop of 1 psi. The f ow coeff cient of control valves can be found in manufacturers’ catalogs. For a water system that has se veral cooling coils between supply and return mains, as shown in Fig. 5.17, modulating the w ater f ow is effective only when the opening and closing of the control valve V1 in piping section EH has a signi f cant ef fect on the change in the pressure drop pEH across piping section EH between the supply and return mains. As control v alve V1 opens wider and pEH decreases, if there is no concurrent change in f ow resistance of piping sections JM and NQ, then a greater w ater f ow will pass through piping section EH. If pEH increases, less water will f ow through EH. If pEH, pJM, and pNQ all increase, based on the characteristic curve of the centrifugal pump, the water f ow rate through pumps P2 and P3 will decrease for an increase in the system head. If pEH, pJM, and pNQ all drop, the water f ow rate through pump P2 or P3 will increase. Because of the use of DDC systems with PI or PID control modes and the v ariable-speed pumping systems, as well as the results of the previous analysis the following hold: 1. The size of a control v alve should be determined according to the f ow coef f cient calculated from Eq. (5.8) and listed in the manufacturer’s catalog. 2. The assumed pressure drop pvv across the control valve should be appropriate. It is af fected by the type of control v alve used. It is also a compromise between a higher pvv value to pro vide desirable controllability and a lower pvv value to save energy. 3. For a w ater system using v ariable-speed pumping and DDC systems with PI or PID control modes, a pressure drop across an equal-percentage control v alve pvv  5 to 10 ft WC (15 to 30 kPa) and a rangeability of 30 : 1 or greater is recommended for energy saving. Example 5.1. A water system using v ariable-speed pumping with DDC supplies chilled w ater to the cooling coils of three air -handling units (AHUs), as sho wn in Fig. 5.17. If the chilled w ater f owing through two-way valve V1 is at 100 gpm (6.31 L/s), select and size control valve V1. Solution. 1. To achieve a nearly linear relationship between the coil load and the tra vel of the valve stem, an equal-percentage valve is selected for valve V1. 2. For a w ater system using v ariable-speed pumping with DDC, it is assumed that the pressure drop across the control v alve pvv  5 ft WC, or 5 0.433  2.165 psi (15 kP a). From Eq. (5.8), the f ow coeff cient is Cv 

V

√pvv



100

√2.165

 68

From one of the manuf acturers’ catalog, for Cv  68, the size of equal-percentage control v alve V1 is 3 in. (76 mm). 3. From the friction chart of w ater in steel pipes, for a chilled w ater f ow rate of 100 gpm (6.31 L/s) and a piping head loss of 2.5 ft /100 ft (2.5 m /100 m) of length, the diameter of the branch piping section EH is also 3 in. (76 mm). The size of the control v alve and the diameter of the branch pipe are the same.

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5.7 DAMPERS AND DAMPER ACTUATORS A damper is a device that controls the airf ow in an air system or ventilating system by changing the angle of the blades and therefore the area of its f ow passage. In HVAC&R systems, dampers can be divided into volume control dampers and f re dampers. Fire dampers are co vered in a later section. In this section, only volume control dampers are discussed.

Types of Volume Control Dampers Volume control dampers can be classif ed as single-blade dampers or multiblade dampers according to their construction. Various types of volume control dampers are shown in Fig. 5.18. Butterfly Dampers A butterf y damper is a single-blade damper. A butterf y damper is made from either a rectangular sheet mounted inside a rectangular duct or a round disk placed in a round duct, as shown in Fig. 5.18a. It rotates about an axle and is able to modulate the air v olume f ow rate of the duct system by varying the size of the opening of the passage for air f ow. Gate Dampers. A gate damper is a single-blade damper . It also may be rectangular or round. It slides in and out of a slot in order to shut of f or open up a f ow passage, as shown in Fig. 5.18 b. Gate dampers are mainly used in industrial exhaust systems with high static pressure. Split Dampers. A split damper is also a single-blade damper . It is a piece of mo vable sheet metal that is usually installed at the Y connection of a rectangular duct system, as shown in Fig. 5.18 c.

FIGURE 5.18 Various types of v olume control dampers: (a) Butterf y damper; ( b) gate damper; ( c) split damper; (d) opposed-blade damper; (e) parallel-blade damper.

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5.33

The movement of the split damper from one end to the other modulates the v olume of air f owing into the tw o legs or branches. A split damper is usually modulated only during air balancing after installation or during periodic air balancing. Opposed-Blade Dampers. An opposed-blade damper is a type of multiblade damper that is often rectangular, as shown in Fig. 5.18 d. It is usually used for a f ow passage of lar ge cross-sectional area. The damper blades may be made of galvanized steel, aluminum alloy, or stainless-steel sheets, usually not exceeding 10 in. (25.4 cm) in width. Rubber or spring seals can be pro vided at the fully closed position to control the air leakage rating, which often does not exceed 6 cfm /ft2 (30 L /s m2) at a pressure drop across the damper of 4 in. WC (1000 P a). The bearing used for supporting the blade axle should be made of a corrosion-resistant material such as copper allo y or n ylon. Tef oncoated bearings may also be used to ensure smooth operation of the damper . Le ver linkages are used to open and close the damper blades. The characteristics of the opposed-blade dampers are covered later in this section. The maximum static pressure drop across closed opposed-blade dampers is 6 in. WC (1500 Pa) for a 36-in.- (914-mm-) long damper (the length of the damper blade) and 4 in. WC for a 48-in.(1219-mm-) long damper. Parallel-Blade Dampers. A parallel-blade damper is also a type of multiblade damper used mainly for lar ge cross-sectional areas, as shown in Fig. 5.18 e. The blade material and the requirement for the seals and bearings are the same as those for opposed-blade dampers.

Damper Actuators (Motors) Damper actuators, also called damper motors, are used to position dampers according to a signal from the controller . As with v alve actuators, damper motors can be classi f ed as either electric or pneumatic. Electric Damper Motors. These either are driven by electric motors in reversible directions or are unidirectional and spring-returned. A reversible electric motor is used often for more precise control. It has tw o sets of motor windings. When one set is ener gized, the motor ’s shaft turns in a clockwise direction; and when the other set is ener gized, the motor’s shaft turns in a counterclockwise direction. If neither motor winding is ener gized, the shaft remains in its current position. Such an electric motor can provide the simplest f oating control mode, as well as other modes if required. Pneumatic Damper Motors. Their construction is similar to that of pneumatic v alve actuators, but the stroke of a pneumatic damper motor is longer. They also have lever linkages and crank arms to open and close the dampers.

Volume Flow Control between Various Airflow Paths For air conditioning control systems, most of the dampers are often installed in parallel connected airf ow paths to control their f ow v olume, as sho wn in Fig. 5.19. The types of air f ow v olume control are as follows: Mixed-Air Control. In Fig. 5.19 a, there are tw o parallel air f ow paths: the recirculating path um in which a recirculating air damper is installed and the e xhaust and intak e path uom, in which e xhaust and outdoor air dampers are installed. The outdoor air and the recirculating air are mix ed together before entering the coil. Both the outdoor damper and the recirculating damper located just before the mixing box (mix ed plenum) are often called mixing dampers. The openings of the outdoor and recirculating dampers can be arranged in a certain relationship to each other . When the outdoor damper is at minimum opening for minimum outdoor air v entilation, the recirculating

5.34

CHAPTER FIVE

FIGURE 5.19 Airf ow paths: (a) mixed- air control, (b) bypass control, and (c) branch f ow control.

ENERGY MANAGEMENT AND CONTROL SYSTEMS

5.35

damper is then fully opened. If the outdoor damper is fully opened for free cooling, the recirculating damper is closed. Bypass Control. In the f ow circuit for bypass control, as shown in Fig. 5.19 b, the entering air at the common junction m1 is di vided into tw o parallel air f ow paths: the bypass path, in which a bypass damper is installed, and the conditioned path, in which the coil f ace damper is installed in series with the coil, or the w asher damper with the air w asher. The bypass and the conditioned airstreams are then mix ed together at the common junction m2. The face and bypass dampers can also be arranged in a certain relationship to each other. Branch Flow Control. In a supply main duct that has man y branch tak e offs, as shown in Fig. 5.19c, there are many parallel airf ow path combinations: paths b1s1 and b1b2s2, b2s2 and b2b3s3, etc. In each branch f ow path, there is a damper in the VAV box, and points s1, s2, s3, etc., are the status points of the supply air. Parallel airf ow paths such as those shown in Fig. 5.19 have the following characteristics: 1. The total pressure losses of the tw o air f ow paths that connect the same endpoints are al ways equal; for example, pum  puom, pm1 bym2  pm1 conm2 , etc. 2. The relationship between total pressure loss p, in. WC (Pa); f ow resistance R, in. WC /(cfm)2 (Pa s2 /m6); and volume f ow rate V˙ , cfm (m3 /s), can be expressed as p  RV˙ 2

(5.9)

Flow resistance is covered in greater detail in Chap. 10. 3. If the total pressure loss p remains constant and the f ow resistance Rn of one parallel path increases, from Eq. (5.9), the airf ow through this path V must be reduced. The airf ow in other parallel paths remains the same. 4. The total pressure loss of an airf ow path between two common junctions p determines the volume f ow rate of air passing through that path and can be calculated from Eq. (5.9) as V˙ 

√

p R

5. When the f ow resistances in most of the branches increase because of the closing of the dampers to a small opening in their VAV boxes, the f ow resistance of the supply duct system R sys and the system total pressure loss psys both tend to increase, and thus the total air v olume f ow of the supply duct system V˙ sys will reduce accordingly.

Flow Characteristics of Opposed- and Parallel-Blade Dampers A parallel-blade or an opposed-blade damper that is installed in a single air f ow path to modulate airf ow is often called a volume control damper (or throttling damper). For volume control dampers, a linear relationship between the percentage of the damper opening and the percentage of full f ow is desirable for better controllability and cost ef fectiveness. (Full f ow is the air v olume f ow r ate when the damper is fully opened at design conditions.) The actual relationship is gi ven by the installed characteristic curv es of parallel-blade and opposed-blade dampers sho wn in Fig. 5.20 a and b. For the sake of energy savings, it is also preferable to ha ve a lower pressure drop when air f ows through the damper at the fully open condition. In Fig. 5.20,  is called the damper characteristic ratio and may be calculated as



ppath pod pod



ppath pod

1

pp-od  pod

(5.10)

5.36

CHAPTER FIVE

FIGURE 5.20 Flow characteristic curves for dampers: (a) parallel-blade and (b) opposed-blade.

ENERGY MANAGEMENT AND CONTROL SYSTEMS

5.37

where ppath  total pressure loss of airf ow path, in. WC (Pa) pod  total pressure loss of the damper when it is fully opened, in. WC (Pa) pp-od  total pressure loss of air f ow path excluding damper, in. WC (Pa) Damper Selection Butterf y dampers are usually used in ducts of small cross-sectional area or in places like VAV boxes. For volume control dampers in a single air f ow path, in order to ha ve better controllability , an opposed-blade damper is recommended if man y dynamic losses other than the damper itself (such as coil or air washer, heat exchanger, and louvers) exist in the airf ow path. If the damper is the primary source of pressure drop in the airf ow path, a parallel-blade damper is often used. For mixing dampers, a parallel-blade damper is recommended for the recirculating damper as the pressure drop across the damper is often the primary source in its air f ow path. An opposed-blade damper is recommended for the outdoor damper and exhaust (relief) damper for better controllability. The parallel blades of the recirculating damper should be arranged so that the recirculating airstream will blow toward the outdoor airstream, resulting in a more thorough mixing. Man y packaged units also use parallel-blade outdoor dampers for smaller pressure drop and less energy consumption. For face and bypass dampers, an opposed-blade coil face damper in an airf ow path of greater pressure drop and a parallel-blade bypass damper will give better linear system control characteristics. For two-position control dampers, a parallel-blade damper is al ways used because of its lo wer price. Damper Sizing Damper sizing should be chosen to pro vide better controllability (such as a linear relationship between damper opening and air f ow), to avoid airf ow noise if the damper is located in the ceiling plenum, and to achieve an optimum pressure drop at design f ow to save energy. The f ace area of the damper Adam , ft2 (m2), in most cases is smaller than the duct area Ad , in ft2 (m2). Based on Alley (1988) paper, the local loss coeff cient Cdam of the damper for different setups can be determined from Fig. 5.21. Then the pressure drop across the damper when the damper is fully opened pod, in. WC (Pa), can be calculated as vdam

2

4005

pod  Cdam vdam 

V˙dam Adam

(5.11) (5.12)

where vdam  face velocity of the damper, fpm. 1. The damper is generally sized when the air f owing through the damper is at a maximum. F or an outdoor damper, the maximum airf ow usually exists when the free cooling air economizer c ycle is used. For a recirculating damper , its maximum air f ow occurs when the outdoor air damper is at minimum opening position, to provide outdoor air ventilation. 2. The face velocity of dampers vdam is usually 1000 to 3000 fpm (5 to 15 m /s), except that the face velocity of a b utterf y damper in a VAV box may drop to only 500 fpm (2.5 m /s) for ener gy savings and to avoid airf ow noise. The ratio Adam /Ad is often between 0.5 and 0.9. 3. The outdoor damper may be either made in a one-piece damper or split into tw o dampers, a larger and a smaller, to match the needs at free cooling and minimum outdoor ventilation. 4. For a bypass damper , its face area should be f ar smaller than that of an air w asher or than a water heating or cooling coil’s face damper. When the air washer or coil’s face damper is closed, the area of the bypass damper should pro vide an airf ow that does not exceed the system design airf ow.

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CHAPTER FIVE

FIGURE 5.21 Local loss coef f cient Cdam of air damper . ( Source: ASHRAE T ransactions 1988, Part I. Reprinted by permission.)

5.8 SYSTEM ARCHITECTURE Architecture of a Typical EMCS with DDC Figure 5.22 shows the system architecture of a typical ener gy management and control system with direct digital control (EMCS with DDC) for a medium or large building. Operating Levels.

Such an EMCS has mainly two operating levels:

1. Unit level. This level is controlled by unit controllers. A unit controller is a small and specialized direct digital controller which is used to control a speci f c piece of HVAC&R equipment or device such as a VAV box, a fan-coil unit, a water-source heat pump, an air-handling unit, a packaged unit, a chiller, or a boiler . For HVAC&R, most of the control operations are performed at the unit level. Since the softw are is often f actory-loaded, only the time schedules, set points, and tuning constants can be changed by the user. Some of the most recently developed unit controllers are also

ENERGY MANAGEMENT AND CONTROL SYSTEMS

5.39

PC Printer

Modem

Operator interface

BAC net

BAC net

SC

UC

UC

AHU

Unit

UC

UC

UC

UC

System

GW

Proprietary Network

BAC net

SC

UC

BAC net

BAC net

Chiller SC

System controller

UC

Unit controller

GW

Gateway

FIGURE 5.22 System architecture of a typical large EMCS.

programmable to v arious de grees. Sometimes the manuf acturer pro vides a v ariety of preprogrammed control sequences, such as monitoring, and diagnostics, and designers can specify the required control sequence that best f ts their designs. 2. System/building le vel. This le vel is controlled by system controllers. Since a system controller has an onboard capacity, programmed by an operator or f actory-preprogrammed software, to execute complicated HVAC&R and other programs, they are the brain of an EMCS. Generic control software such as scheduling, trending, alarming, diagnostics, and security is also pro vided in system controllers. Generic control is covered in detail in a later section. A system controller is used to coordinate the control operations of an HVAC&R system, such as the coordination between the duct static pressure and the total air v olume of VAV boxes in an air system, or the sequencing of three centrifugal chillers and cooling to wers in a refrigeration plant. A system controller may interf ace with sensors /transmitters by means of input /output (I /O) connections directly. Unit controllers are also con f gured on a separate subnetw ork and connected to a system controller. Operator P ersonal Computer (PC) Workstation. The operator may interf ace with the EMCS primary through an operator ’s PC w orkstation or purpose-b uilt device, either handheld or f xed to the system controller . Each w orkstation shall consist of a PC, autodial telephone modems, and printers. The central processing unit (CPU) shall be a minimum of an Intel 80486 and operated at a minimum of 33 MHz. Its memory includes a minimum of 8 Mbytes RAM and 212 Mbytes hard disk. The communication ports connected to system controllers and other control systems should be provided. A 14-in. (356-mm) color monitor also shall be provided.

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CHAPTER FIVE

The software in the workstation shall do the following: ●

●

●

●

●

Accommodate processes as well as prioritize applications based on their input/output priority level. Provide system graphics including the HV AC&R equipment (such as a display of up to 10 graphic screens at once for comparison) to monitor the operating status of the system. The operator with the proper passw ord is able to add, delete, or change the set points, time scheduling, etc. Support the editing of all system applications including the generic control softw are provided in the system controllers. The edited or custom programmed softw are shall be downloaded and executed at one or more of the system controllers. Automatically save the database and restore the database that has been lost in one of the system controllers. Provide scheduling, trends, totalization, alarm processing, security to vie w and edit data, and system diagnostics.

Communication Netw ork. A peer -to-peer data communication of a local-area netw ork (LAN) adopting either Ethernet or ARCNET will be used between system controllers and between system controllers and PC w orkstations (or other system). A peer -to-peer communication means that all system controllers or work stations have equitable access to communication resources. For the communication subnetw ork between system controllers and unit controllers, and between unit controllers themselves, a master-slave token-passing (MS /TP) technology is often used. A system controller also acts as a medium to pro vide data communication between the w ork station and the unit controllers. The network technology is covered in detail in a later section. Power Source. The temperature sensors and humidity sensors may need up to 12 V dc and 24 V ac as a power source. Many DDC units have a power source of 24 V ac or 120 V ac line voltage. Most of the valve actuators and damper motors need a power source of 24 V ac. Size of EMCS. The size of an EMCS depends on the number of points (or objects) that belong to its DDC units. An EMCS of 100 points or less can be considered a small project. An EMCS that has 1000 points or more can be considered a large project.

System Characteristics An architecture of EMCS incorporating DDC such as that in Fig. 5.22 has the follo wing characteristics: 1. All DDC units are independent and stand-alone controllers. If an y of the controller fails, there is only a limited effect. 2. The system architecture shows a distributed processing model. Since most of the control operations are performed at the unit controller le vel, and partly at the system controller le vel, such an architecture has the adv antage that it tremendously reduces the data communication between the unit controllers and the system controller, as well as between the DDC system controller and the PC workstation. 3. Zimmerman (1996) noted that since the introduction of DDC in the 1980s, “. . . microprocessor and memory ha ve declined rapidly in cost while wiring and installation costs ha ve not declined at the same rate.” To provide more powerful unit controllers, moving the controllers nearer to the sensors and control de vices will reduce a lot of wiring and installation costs as well as the overall system cost. 4. If each HVAC&R piece of equipment has its own unit controller, it surely will be benef cial to the HVAC&R equipment manuf acturers to f abricate controllers and other control system components themselves.

ENERGY MANAGEMENT AND CONTROL SYSTEMS

5.41

Future Development The development of more po werful, programmable unit controllers using a modular con f guration, will reduce the dif ference between a system controller and a unit controller . As predicted by Hartman (1993), the future architecture of an EMCS may have only a single tier of various kinds of unit controllers. The operator’s PC workstation, modems, and unit controllers will all be connected to a peer -to-peer data communication trunk. Such an architecture will simplify the DDC units and the communication netw ork, move the intelligence nearer to the control de vices, and f nally create more effective control at lower system cost.

5.9 INTEROPERABILITY AND OPEN PROTOCOL BACnet Interoperability Turpin (1999) de f ned interoperability as “the ability of systems including equipment and components from different manufacturers to share data and information for the purpose of operation with plug-and-play connectivity.” Interoperability is one of the necessary conditions for system inte gration. System integration is a strategy to integrate various HVAC&R systems of various manufacturers together, and to integrate HVAC&R systems with lighting, f re protection, security, elevator, and electrical systems in a building together. Robertson and Moult (1999) note that the advantages of system integration include the following: ●

●

●

●

●

It reduces the installation cost of the building automation system. It enhances energy management. It can apply building automation system features. It provides building operator training. There is a single user interface.

The greater the interoperability of a single system and the more systems you try to inte grate into a single system the more complicated and costly the process will be. Interoperability and system integration is one of the goals of the HVAC&R industry. More and more engineers, facility owners, and manufacturers recognize this need. It takes time to accomplish such a complicated process. BACnet — Open Data Communication Protocol BACnet means building automation and control network. It is an open data communication protocol. Open means that an independent institution go verns its de velopment. All contents are kno wn, f xed, and accessible. A protocol refers to the rules by which tw o or more de vices communicate data to each other that must be obeyed. BACnet enables that building automation devices from various manufacturers can talk to one another , share data, and work together following a standard way. BACnet def nes all the elements of data communication between de vices in a b uilding automation control system. It is specif cally tailored for HVAC&R control equipment, but it also provides a basis for inte grating lighting, security, and f re-detecting systems. BACnet was developed from 1987 to 1995 by ASHRAE and was adopted as a national standard in 1995 by ANSI, as ANSI/ASHRAE Standard 135-1995. For details, refer to ASHRAE’s BACnet. BACnet will assist building owners, designers, contractors, and operators in three areas: 1. It gives more freedom to select the best equipment and components from dif ferent manufacturers in order to have a more eff cient system at lower cost. 2. It operates the control system from a single w orkstation; i.e., it is more ef fective to operate and easier to maintain, and there is only one system to learn.

5.42

CHAPTER FIVE

3. It collects data from dif ferent systems and of fers greater f exibility for e xtended systems in retrof t projects. Most of the EMCS manufacturers agreed to fabricate BACnet-compliant products from the late 1990s. There are man y other data communication protocols. One w as developed by Echelon Corporation, called LonTalk Protocol, and it became a w orking system in the mid-1990s and w as favored by members of the LonMark Interoperability Association. LonTalk protocol is neither an opento-public protocol nor a standard. An independent consortium called the OPC foundation, formed as a nonpro f t organization in 1996 in Boca Raton, Florida, has dedicated itself to pro vide interoperability with Microsoft technologies to de velop a global speci f cation and multi vendor interoperability in industries. In 1990, OPC is leading 140 member companies including Honeywell , Johnson Controls, and Siemens.

Application Layer Layered Structure. A data communication system often adopts a hierarchical layered structure so that a comple x problem is brok en into smaller and more easily solv ed problems. B ACnet is based on a four-layer collapsed architecture that corresponds to application, network, data link, and physical layers in an International Or ganization for Standardization (ISO) model. This is the result of careful consideration of the characteristics and requirements of the b uilding automation control (including HVAC&R) together with a constraint that protocol overhead be as low as possible. An application layer is the highest layer in B ACnet. It serv es to de f ne the objects and services (including control operations, information exchange, and control devices) in a b uilding automation control (B AC) system. It also pro vides communication services and data encoding schemes required by applications to perform monitoring and control functions. Object Types, Properties, and Devices. The BACnet def nes a set of standard object types instead of conventional points. Analog input, analog output, binary value, command, f le, program, schedule, etc., grouped in 18 types are standard object types, for every device in an EMCS must ha ve a device object. There are 123 properties that have been identi f ed by B ACnet which fully describe the device, or object type, in the netw ork. Certain properties are required to be speci f ed whereas others are optional. An object identi f er speci f es its object name, object type, etc., and optional properties such as description and device type. In BACnet, a device is def ned as any device, real or virtual, that supports digital communication using the BACnet protocol. Services. In BACnet, services are the operations by which one de vice acquires information from another de vice, commands another de vice to do something, or announces that some e vent happened. BACnet def nes 32 services that can be grouped into f ve categories: ●

●

●

●

●

Alarm and e vent services refer to changes in conditions detected by a de vice, such as ackno wledgment of an alarm and conf rmed change of value notif cation. File access services are used to read and manipulate f les that are k ept in de vices, such as only one read or write operation at a time. Object access service pro vides the means to read, to write, and to modify properties, such as to add one or more items to a property. Remote device management offers disparate operations such as to tell a de vice to stop accepting messages. Virtual terminal services are used by an operator to establish a bidirectional connection with an application program implemented in a remote device.

Services are classif ed as conf rmed when a reply is usually e xpected with data and unconf rmed when no reply is e xpected. In B ACnet, a given device is not required to implement e very service. However, “read property” is required to be executed by all the devices.

ENERGY MANAGEMENT AND CONTROL SYSTEMS

5.43

Conformance Class, Functional Groups, and Protocol Implementation Conf ormance Statement (PICS). BACnet def nes six le vels of conformance for all de vices, classes 1 to 6, that are hierarchical to indicate the dif ference in requirements that must be met to conform to B ACnet. The requirement of a class includes all the requirements of all the other classes ha ving a lower number. At the lowest level, conformance class 1 requires only that a de vice be able to e xecute (respond to) a “read property service” request, such as a sensor. For conformance class 6, a device is required to implement 21 types of service requests, such as a PC workstation. A functional group de f nes a combination of services and object types that are required to perform certain B AC functions. In B ACnet, there are altogether 13 functional groups, such as clock, workstation, and event initiation. The protocol implementation conformance statement (PICS) is a document pro vided by the manufacturer of a de vice to identify those options implemented by a particular de vice. The PICS covers the conformance class of the de vice, supported functional groups, standard and proprietary services executed and initiated, etc. Data Encoding. In BACnet, application-layer protocol data units (APDUs) are used to convey the information contained in the service primiti ves and associated parameters. ISO Standard 8824, Specif cation of Abstract Syntax Notation One (ASN.1) , has been chosen as the method to represent the data content of BACnet services. Each data element consists of three components: (1) identif er octets, (2) length octets, and (3) content octets. The f xed portion of each APDU providing protocol control information is encoded implicitly, and the variable portion of each APDU providing servicespecif c information is encoded explicitly.

Network Layer In BACnet, the purpose of the netw ork layer is to pro vide the means from which messages can be relayed from one B ACnet network to another in the internetw ork. Two or more B ACnet networks are connected by routers to form a B ACnet internetwork. A router, a BACnet device, is used to interconnect tw o disparate B ACnet local-area netw orks (LANs). A netw ork layer directs the messages to a single remote de vice or broadcasts the messages on a remote netw ork, or to all de vices on all networks. A device is located by a netw ork number and a medium access control (MA C) address. Another network layer function is message segmentation and reassembly. In an EMCS, there are often tw o netw orks: one uses Ethernet or ARCNET for high-speed message transmission between system controllers and the PC w orkstation, and another adopts lowspeed message transmission between unit controllers and system controllers. Consequently , a network layer is required for the BACnet protocol.

Data Link/Physical Layer — Network Technology In B ACnet, a data link layer has the capability to address messages to a single de vice or to all devices. At the data link layer , only incoming B ACnet messages recei ved from the physical layer are passed on to the network layer, and a code is added to the outgoing messages to identify them as BACnet messages before they are passed to the physical layer . A physical layer provides the physical medium for message transmission. For local-area netw orks, BACnet supports Ethernet, ARCNET, MS/TP, PTP, and LonTalk as alternatives. Various media are used as physical transmission entities. Typical media are twistedpair wire, f ber-optic cable, and coaxial cable. Ethernet. It is one of the commonly used LAN technologies. At any time, multiple devices may access the netw ork simultaneously, i.e., multiple access with collision detection. Ethernet is using a peer-to-peer communication with a b us network conf guration. Ethernet often runs at a speed of 10 Mbits /s. Two types of coaxial cable are often used in Ethernet: thick-wire and thin-wire. A thick-wire Ethernet segment has a maximum length of 1600 ft (488 m), and up to 100 nodes can be

5.44

CHAPTER FIVE

attached. Thick-wire Ethernet is more e xpensive. A thin-wire Ethernet se gment has a maximum length of 600 ft (183 m) and 50 attached nodes. ARCNET (Attached Resources Computer Netw ork). This is also a commonly used LAN technology and is lower in cost than Ethernet. ARCNET adopts a token-passing network access method. A tok en which indicates the permission to use the physical medium is passed from one netw ork node to the next node in a circular manner . As in Ethernet, ARCNET also uses a peer -to-peer communication network, and its network nodes reside on a bus. Both coaxial cable and twisted-pair wire are used in ARCNET and run at a speed of 2.5 Mbits/s. F or each ARCNET se gment, up to 8 nodes can be connected. All together , up to 255 nodes can be communicated over an ARCNET network. Master-Slave/Token-Passing (MS/TP). In BACnet, MS/TP divides all the nodes on the netw ork into tw o cate gories: masters and sla ves. Only masters can initiate data communication, whereas slaves cannot initiate. Slaves can only respond to requests from masters. The MS / TP also adopts a token-passing network access method. A master node may access the netw ork only when the tok en (permission to use the medium) is passed to it from the pre vious master node. The tok en ne ver passes to the slave nodes. Master nodes in an MS / TP network are at a peer-to-peer communication. The MS /TP network often uses twisted-pair wires. It can operate at a speed of 9600 bits /s, as well as 19.2, 38.4, and 76.8 kbits / s. Point-to-Point (PTP). This is a data link layer protocol which pro vides serial communication between two devices. Such a point-to-point communication typically in volves a dial-up phone modem or hardwired connection between tw o nodes. PTP has a simpler medium access mechanism and is often temporary in nature. PTP is much slo wer than a LAN. Both de vices can recei ve and transmit simultaneously. LonTalk LAN. This is an option of the physical medium in B ACnet and is at the base of the Lon Mark protocol. LonTalk supports a number of choices of physical media. Connection between BACnet and Proprietary Network For an extension project, it is possible that the original b uilding is still intended to k eep the proprietary netw ork and the e xtended part is to construct a B ACnet netw ork. The proprietary netw ork needs a gate way to connect to the B ACnet network, as shown in Fig. 5.22. A proprietary netw ork does not open to others. Typically, information can only be exchanged between EMCS components and the proprietary network of the same manufacturer. For two networks and their computers using different protocols to communicate, some translation must take place. The device that provides this translation is called a gateway. LonTalk Protocol According to Glinke (1997), the Local Operators Netw ork (LonTalk) protocol is a se ven-layer proprietary protocol de veloped by Echelon Corp. Data communication is implemented on a neuron chip. There are actually three microprocessors within the chip; tw o are used for the netw ork and one is for speci f c functions. These de vices, which allo w dif ferent types of e xisting netw orks to communicate, use multiple, low-cost media (Ethernet, ARCNET, f ber optics) and pro vide f exibility. Since this technology is essentially peer -to-peer communication without the need for a central supervisory control node, Lon Mark-certi f ed unit controllers can e xist on the same netw ork as equals. The two promising and widely used protocols, BACnet and LonTalk, were able to communicate with each other in the late 1990s. They will impro ve themselv es through actual operation in the future and produce better choices for engineers and users.

ENERGY MANAGEMENT AND CONTROL SYSTEMS

5.45

5.10 CONTROL LOGIC AND ARTIFICIAL INTELLIGENCE In an EMCS, the software in the DDC units determines the control functionality . Since the logic is separate from the hardware, the control functions and sequences are now limited only by the knowledge and innovation of the designer and operators. The development of artif cial intelligence in the DDC softw are — fuzzy logic, expert systems, artif cial neural netw ork, etc. — tremendously e xpands the control f exibility and improves the operating quality to meet the more comple x requirements for an HVAC&R energy management and control system.

Fuzzy Logic Basics. Since many systems have become more and more comple x over the past decades, an accurate analytical model based on rigorous and nonlinear mathematics is dif f cult to de velop and to be accepted in the daily management by operators. In 1965 Professor Lot f Zadeh of the Uni versity of California, Berkeley, developed fuzzy set theory which provides a new control logic for a system. Fuzzy logic has the ability to deal with the imprecision that happens in e veryday life. Conventional digital technology is based on bi valence: yes or no, on or of f, 1 or 0, and black or white. Fuzzy logic is multi valent. Things can be partly yes to some de gree, partly no to some de gree, and fuzzy logic deals with shades of gray . Conventional set theory observ es that a f act must be either true or false, whereas in fuzzy logic set theory, a fact can be partly true and partly f alse, can belong to a set and also belong to another set. According to Lehr (1996) and Scholten (1995), fuzzy logic of fers simplicity in the midst of complexity and is a real alternati ve for system operation and control. The ob vious bene f ts of fuzzy logic are that it makes things more human and more friendly to a person who is less trained, and it is more easily maintainable. Fuzzy logic really is not a v ague theory. It is comprised of a set of precise rules based on rigorous mathematics which go verns the behavior of a system by means of w ords and phrases instead of nonlinear models. Fuzzy logic controllers (FLCs) ha ve been widely used in air conditioners, humidif ers, refrigerators, and many other de vices such as ele vators. There are also man y applications for which con ventional control logic is better than fuzzy logic controller. Fuzzy Sets and Membership Function. Fuzzy sets, membership function, and production rules are three primary elements of fuzzy logic. Figure 5.23 sho ws the fuzzy sets, membership functions, and a diagram of fuzzy logic control. In con ventional bivalent crisp set theory , “sets” of thermal comfort of an occupant in an air conditioned space can be cate gorized according to the indoor temperature Tr exactly as cool, a range of 62 to 72 °F (16.7 to 22.2 °C); and just right, a range of 72 to 76°F (22.2 to 24.6 °C), etc., as shown in Fig. 5.23. At Tr  74°F (23.3°C) and Tr  75°F (23.9°C), both are “just right” — have a membership v alue of 1 with re gard to the just-right set and a membership value of 0 with regard to the cool set. On the other hand, the fuzzy set “just right” ranges between a membership v alue of 0 at 70 °F (21.1°C), a membership v alue of 1 at 74 °F (23.3°C), and a membership v alue of 0 again at 78 °F (25.6°C). F or Tr  75°F (23.9 °C), the membership v alue of fuzzy sets w ould be described as 75 percent just right and 25 percent w arm. In another w ay, the assertion of thermal comfort of just right is 75 percent true and 25 percent false. If Tr  76°F (24.4°C), then the conventional crisp set will be dif f cult to determine. In addition, for Tr  74°F (23.3°C) and Tr  75°F (23.9°C), the assertion of the conventional crisp set for both is “just right,” so there is no difference between them. And the assertion of fuzzy sets for Tr  74°F (23.3°C) is 100 percent just right, which is different from 75 percent just right and 25 percent warm when Tr  75°F (23.9°C). Production Rules. When the fuzzy set of an occupant ’s thermal comfort is integrated with another fuzzy set of percentage of f an speed v ariation (positive large, positive small, zero, negative small,

CHAPTER FIVE

0 100 80

0.5

1

PL

60 Fan speed variation range, %

5.46

40

If warm, then decrease a little

PS

20 0

ZE

20

32

40

NS

If just right, then do nothing Centroid If cold, then increase a lot

60

80

If hot, then decrease a lot

NL

If cool, then increase a little

100 PL PS ZE NS NL

Positive large Positive small Zero Negative small Negative large

1 0.5 0.45 0.2 Cold

Just right

Cool

0

62

66

7071

74

Fuzzy sets

Warm 78

82

Hot 86

90

Indoor temperature, F

1 0.5 Cold 0

Just right

Cool

62

66

70

74

Conventional crisp sets

Warm

78

82

Hot

86

90

Indoor temperature, F

FIGURE 5.23 Fuzzy sets, membership function, and fuzzy logic control.

and negative large), then the production rules, or fuzzy logic rules, can be described as follows: ●

●

●

●

●

If the indoor temperature is hot, then the fan speed rises a lot. If the indoor temperature is warm, then the fan speed rises a little. If the indoor temperature is just right, then the fan speed stays unchanged. If the indoor temperature is cool, then the fan speed reduces a little. If the indoor temperature is cold , then fan speed reduces a lot.

ENERGY MANAGEMENT AND CONTROL SYSTEMS

5.47

Fuzzy Logic Controller . An FLC consists of three parts: a fuzzi f er converts ordinary inputs to fuzzy variables, a fuzzy reasoning unit produces fuzzy control signals based on input fuzzy v ariables, and a defuzzif er converts fuzzy control signals to conventional control outputs. If the indoor temperature Tr is 71°F (21.7°C), then the membership function of the fuzzy sets of thermal comfort in Fig. 5.23 would be described as 0.2 (20 percent) just right and 0.45 (45 percent) cool. From the production rules, then, “reduce the f an speed a little ” has a membership v alue of 0.75 and “fan speed remains unchanged” has a membership value of 0.2. One way to interpret these two fuzzy outputs in a con ventional crisp output is to determine the centroid based on the area of the two truncated triangles. From Fig. 5.23, the output value is 32 (reduce the f an speed by 32 percent of its range). Huang and Nelson (1994) noted that an FLC is characterized by a set of linguistic fuzzy logic rules. The initial set of these rules is often based on past e xperience or analysis of the control process. The initial set of rules can be modi f ed by analyzing the performance trajectory on the linguistic plane to obtain an optimal rule set. This is called rule ref nement. The second signif cant inf uence on the behavior of an FLC is the choice of membership functions. The overlap of the fuzzy sets should be moderate to allo w for reasoning with uncertainty and the need for completeness of the control rules. Refer to Huang and Nelson’s paper for details.

Knowledge-Based Systems and Expert Systems Basics. A knowledge-based system (KBS) is a kno wledge-rich, logic-oriented computer program that mimics human kno wledge and reasoning in a speci f c domain to assist in solving more complex problems. Human kno wledge includes e xpert kno wledge, common f acts, and kno wledge in any form, whereas an expert system, strictly speaking, mimics only an expert’s expertise in a given domain to solve specif c problems. An expert system can be considered as the core part of a kno wledge-based system. In comparing a kno wledge-based system with con ventional programmed softw are, Hall and Deringer (1989) noted that the benef ts of a knowledge-based system are due to the following superior abilities: ●

●

●

●

Logical reasoning. Logical problems are solv ed by means of sorting, comparing, searching, and reasoning as well as to e valuate alternati ve choices to a speci f c problem. More often, several causes (or solutions) of a problem are identif ed. Resolution of uncertainty. When precise knowledge is not available and for problems that involve qualitative facts and incomplete data, the certainty of the conclusions of the KBS using heuristic rules depends on the reliability of the heuristic rules themselves. Multiple approaches. Multiple e xperts are often used to de velop a kno wledge base. Experts will rarely agree exactly on the proper approach to solve a problem. Solution of the problem with multiple approaches is one aspect of KBS. Justif cation of results. A well-developed KBS has the capability to recall the basis of each decision during the problem-solving process. It can pro vide a decision trail that will e xplain the sequence of decisions used to produce the conclusion.

The limitations of a KBS are as follows. First, a KBS can make no decision that is not explicitly contained in the kno wledge base. Second, two dif ferent KBSs de veloped using the e xpertise from different e xperts may gi ve dif ferent advice for the same problem. Third, the KBS is timeconsuming and expensive. Knowledge-based systems will not replace e xperts and engineering professionals. Instead, the systems will play the role of specialized knowledge-based advisers or consultants. KBS Structure.

An expert system comprises the follo wing four modules, as shown in Fig. 5.24.

5.48

CHAPTER FIVE

FIGURE 5.24 Structure of a knowledge-based system for an EMCS and HVAC&R design.

1. Knowledge base. This is the most important part of an e xpert system. It contains the common facts and inference rules, often in the form of if-then, and is domain-speci f c. Ev ery rule in the knowledge base is e xtracted from the e xpert’s knowledge and e xperience by the de veloper of the KBS, the kno wledge engineer , or is based on data and information from published handbooks, manufacturers’ engineer manuals, and f eld survey results. Frames are another common kno wledge representation method. Hall and Deringer (1989) de f ned frames that “. . . facilitate the description of knowledge about complex objects, which have many subelements, . . .” The content of the knowledge base should be tailored to the user’s task and requirements. The user’s task can be determined through direct interview with potential end users. 2. Inference engine . The inference engine asks for inputs from the user interf ace, executes reasoning algorithms by applying the kno wledge from the kno wledge base, and arrives at the conclusion based on the rules. There are tw o kinds of reasoning: (1) data-driven forward chaining or forward reasoning and (2) goal-dri ven backward chaining or backw ard reasoning. They are categorized according to ho w new information is inferred. F orward reasoning proceeds forw ard from the user inputs about probable outcomes. In most engineering design problems, forward reasoning is adopted. F or instance, knowledge of the b uilding envelope leads to load calculations, which forwards next to determining the equipment capacity, and so on. Backward reasoning proceeds backw ard from the user inputs about probable outcomes. In troubleshooting or diagnostics, backward reasoning is used. F or an e xample, collected f acts lead to determining the cause of the problem. F orward and backw ard reasoning were often used in conjunction to control the f ow of questioning as well as pro vide a dialog between the system and the end user. 3. User interface. This includes the f acilities to support smooth and con venient interaction with the user. Only simple instruction is needed to start the system, end a session, save the contents, and

ENERGY MANAGEMENT AND CONTROL SYSTEMS

5.49

give a printout. A friendly dialog between the KBS and the user and a simple question-and-answer format with a pro vided menu of possible answers are e xpected. An intelligent operating system which accesses v arious application tools is recommended. Modular ne w application tools can be plugged into the operating system as required. A user interface with built-in intelligence, graphically based high-resolution display with standardized menus and format may be the preferable answer. 4. Knowledge acquisition . A kno wledge acquisition module pro vides strate gies to capture the experts’ knowledge to develop a KBS. Sometimes, it also checks for consistency and completeness. Knowledge acquisitions are usually accomplished through personal intervie ws with experts and review of the application literature. Most HVAC&R-related knowledge systems have long-term plans to continue the process of kno wledge acquisition. A new trend is machine learning; i.e., the computer learns from the experience how to capture and manipulate new knowledge. Development of KBSs. Many KBSs are de veloped by using commercially a vailable development tools called shells. A shell consists of mainly a rule editor , a knowledge base, an inference engine, and user interf aces. Usually, the inference engine and user interf ace are fully de veloped within a shell. Thus, the user can focus on the collection and the input of the kno wledge. He or she can change the e xisting kno wledge base and does not need to change the entire system. KBS is e xpected to operate on a PC. The performance of a KBS and the resources needed during de velopment are highly affected by its knowledge engineer. The knowledge base is b uilt through the cooperation of the kno wledge engineer and experts in a specif c domain. The knowledge engineer has the responsibility to choose an appropriate inference strategy, and a suitable shell and to ensure compliance of the system with the task. The development cycle of a KBS is an iterati ve and incremental process. It begins with the initial prototype. The next is an improved, and expanded, one. The process is repeated and may tak e years. How the KBS Works. For e xample, if the space air is too humid and the space cooling load is only one-half of the design load, it is required to f nd the general cause of these symptoms. The problem is a diagnosis problem. Brothers and Coone y (1989) stated that if-then rules are composed of parameters (symptom, set point) and v alues (too hot, too humid, correct). In the inference engine, the computer program of the KBS will be gin to search for an if-then rule through the knowledge base that will gi ve a v alue (too humid) for a general cause. If the v alues do not match those required by the rule, the computer program will search for the ne xt rule, until the follo wing if-then rule has been found: IF symptom is too humid, and cold supply air temperature set point is correct, and cold supply air relative humidity is O.K., and sensible cooling load is 50 percent of design load THEN general cause is size of supply airf ow rate CF 90 The cause of a too humid space air is that the volume f ow of supply air is too great. The term CF is the abbreviation of certainty factor. That means the con f dence in the answer to the problem is 90 percent. Testing, Verification and Validation. In Jafar et al. (1991), testing often detects logical errors related to the kno wledge base, syntactic errors, and missing kno wledge. Testing only sho ws errors and does not e xplain their cause. Verif cation and v alidation should be performed at each stage of development, to check the knowledge base for internal inconsistencies, mismatches, etc. Applications. There are three primary areas of HVAC&R-related applications of KBS, as reported in the paper by Hall and Deringer: ●

Monitoring — interpretation of measured data in comparison to expected behavior

Input layer i

Hidden layer 1

Input

Neuron

W12,1

Hidden layer 2... L

Output layer o

Neuron

Neuron

W i1,1

W

Lo

,1

Outputs

o,n

WL

W

i1,n

W12,n (a)

W

i1,1

Wi1,2

Neuron 1n

Output

W i1,n

f(x)

1

0

i

(b) FIGURE 5.25 An artif cial neural network (ANN): (a) Structure of an ANN; (b) sigmoid function.

5.50

ENERGY MANAGEMENT AND CONTROL SYSTEMS ●

●

5.51

Diagnostics — assistance in identifying solutions to complex technical problems Design — assistance in the selection of the HVAC&R systems and subsystems

Artificial Neural Networks Basics. An artif cial neur al network (ANN) is massi ve interconnected, parallel processing, dynamic system of interacting processing elements that are in some aspect similar to the human brain. The fundamental processing element is called the neuron, which is analogous to the neural cell in human brain. The neurons are set in layers, and thus a netw ork is formed as sho wn in Fig. 5.25. Inputs representing the variables that affect the output of the network are feeding forward to each of the neurons in the follo wing layers with an acti vation depending on their weighted sum. Finally, an output can be calculated as a function of the weighted sum of the inputs and an additional factor, the biases. The ability to learn is one of the outstanding characteristics of an ANN. The weights of the inputs are adjusted to produce a predicted output within speci f ed errors. ANNs have been increasingly used in recent years to predict or to impro ve nonlinear system performance in HVAC&R. An ANN system is characterized by its net topology , neuron activations transfer, and learning method. Net Topology. The structure of the network of an ANN, or net topology, depends on the data f ow mode, the number of layers, and the number of hidden neurons. ●

●

●

In Miller and Seem (1991), there are two types of data f ow modes: state and feed-forward models. In the state models, all the neurons are connected to all the other neurons. In feed-forw ard models, the neurons are connected between layers, as shown in Fig. 5.25 a, and the information f ows from one layer to the ne xt. Feed-forward models are the most popular and most often analyzed models. In an ANN, there is always an input layer with the number of inputs equal to the number of parameters (variables) that affect the output. There may be one or more hidden layers of neurons ne xt to the input layer. The selection of number of hidden layers and the number of neurons in each hidden layer remains an art. Curtiss et al. (1996) noted that too man y hidden layers and hidden neurons tend to memorize data rather than learning. The hidden layers and hidden neurons must be suf f cient to meet the requirement during the learning process for more comple x nonlinear systems. More hidden layers and hidden units need more calculations and become a burden.

Among the 10 papers published from 1993 to 1996 in ASHRAE Transactions regarding developed ANNs, most have only one hidden layer, some have two layers, and only one has three hidden layers. None e xceeds three hidden layers. Ka washima (1994) recommended that in an ANN only one hidden layer is suff cient for load prediction. ●

●

If the relationship between the inputs and output is more complex, i.e., nonlinear, and more inputs are involved, then more neurons are needed in each hidden layer . Ka washima (1994) also suggested that the number of neurons in each hidden layer e xceed 2 m 1. Here m indicates the number of inputs. There is al ways an output layer ne xt to the hidden layer(s). It is preferable to ha ve one neuron (single output) in the output layer for simplicity . There may be tw o or more neurons for multiple outputs.

Neuron Activation Transfer. In Miller and Seem (1991) and Curtiss et al. (1996), for each neuron in the hidden and output layers: 1. The input acti vations to a neuron in the

f rst hidden layer

h, denoted by i1n, can be

5.52

CHAPTER FIVE

calculated as i 1n  i nWi1, n B  S B

(5.13)

where in  normalized input Wi1,n  weights of connection between inputs and neurons in f rst hidden layer B  biases in 

i o i min i max i min

(5.14)

where io  original input data imax, imin  maximum and minimum of original input data 2. The output from the neurons in the f rst hidden layer o1n, is expressed as a selected sigmoid activation function: o1n 

1 1 e (S B)

(5.15)

There is only one output o for a neuron. This output is transmitted through output connections in which it usually splits into multiple connections with identical activations. 3. The input activation to a neuron in the hidden layer L, denoted iLn, or a neuron in the output layer o, denoted by ion, is equal to the output of the neuron in the pre vious layer which split identically to all the neurons in hidden layer L. The output of a neuron in hidden layer L, denoted by oLn, or output layer oon can be calculated as oLn 

1 1 e (T B)

1 oon  1 e (V B)

(5.16)

and T  i 1nW1L,n V  i LnWLo,n

(5.17)

Learning Method. The learning method, also called the training process, determines the connection weights Wxyn through kno wn sets of input /output pairs. The initially assigned connection weights are adjusted repeatedly during the learning process until the error is within the speci f ed values. After training, the ANN can predict the outputs from gi ven inputs. Backpropagation is the most often used systematic method to train multilayer ANNs. Curtiss et al. (1996) and Miller and Seem (1991) recommended the following training procedure: 1. Assign initial random v alues for connection weights, often between 0.5 and 0.5. Select a training input output pair; calculate the normalized input activations in and the input activation to the neurons in the f rst hidden layer according to Eqs. (5.13) and (5.14). 2. Calculate the outputs of the 1, 2, . . ., L hidden layers o1n, o2n, . . ., oLn and then of the output layer oon from Eqs. (5.15) and (5.16).

ENERGY MANAGEMENT AND CONTROL SYSTEMS

5.53

3. Evaluate the error  between the calculated oon and the selected training output (tar get output) ot by:

  oon(1 oon)(ot oon)

(5.18)

4. Adjust the connection weights from Wj to Wj 1 to minimize the error by using the follo wing rule: Wj 1  Wj oon

(5.19)

where is the learning rate whose value lies between 0 and 1. 5. Repeat the previous steps for all the input /output pairs in the training set until the error for the entire training set is lo wer than the preset training tolerance. The training set should co ver the operating range of the inputs and outputs. Applications. An ANN can model multiple parameters simultaneously for nonlinear systems. It can also be periodically trained to update the weights. ANNs are no w widely used for predicti ve control, such as ener gy use prediction, energy optimization, adaptive control, data trending, and optimum start and stop.

5.11 PROGRAMMING OF DDC SYSTEMS Evolution of DDC Programming The EMCSs having more complex functional controls are hea vily software-driven. From the 1960s to the 1970s, the software for HVAC&R control w as mostly performed in a w orkstation computer and programmed in the control manuf acturer’s f actory. Because of the trend to use distrib uted microprocessor-based DDC units in the late 1970s, many EMCS manufacturers provide some type of operator control language (OCL) for f eld-programmable line programming in the controllers using BASIC-type language. This meets the requirements of computer program de velopment, periodic updating, and necessary modif cations and improvements. Since the late 1980s, there is a trend to ward the function and object-orientated graphical programming to provide software for DDC units and the PC in the operator ’s workstation. Today, both traditional line programming using BASIC-like language (sometimes Pascal and C++ languages are also used) and graphical programming are used in DDC systems.

Graphical Programming Davison (1992) described that graphical programming is a schematic dra wing of a desirable control scheme (functional control), such as mechanical cooling or economizer control, using symbols called templates. The templates are displayed on the computer screen interconnected by lines that direct the f ow of data. The control scheme shows the inputs, through control operations, to outputs. After the diagram of the graphical programming is completed, it is con verted to a program usable by the DDC units and PC through a computer program. Compared with traditional line programming using a B ASIC-like language, graphical programming has the follo wing advantages: First, it is intuiti ve and easier to use and understand by f eld HVAC&R engineers and operators. Second, the user is more familiar with it because of meaningful visual symbols represent functions and de vices. It is simpler in documentation and easier in troubleshooting. Third, it pro vides a time-sa ving tool for the speci f cation and documentation of the HVAC&R energy management and control systems. Finally, graphical programming enables one to improve programming quality at a reduced cost. The limitations of graphical programming are that

5.54

CHAPTER FIVE

Zone schedule Zone temperature west

Time clock

74

Cool SP T 1

85

T 0

Cooling control

10K Zone temperature east



2



Average temperature

10K

SP Cool I DB

Economizer Min on Min off

T

0.5

Compressor Outdoor temperature 10K

Outdoor temperature lockout SO

SP Cool I DB

Min on Min off 180

3

Templates

DB

Dead band

SP

Set point

FIGURE 5.26 Graphical programming for mechanical cooling control in a small packaged unit.

it is bulky to display on screen and that it is ne w to us and needs to de velop required supports such as a portable interface that acts directly to the DDC units.

Templates In Davison’s (1992) paper , a template is a graphical symbol (icon) in graphical programming that describes a single or combined control scheme (speci f c functional control), as shown in Fig. 5.26. A template consists of input and output connections, a small section of computer code, and private data storage registers. A template is function- and object-oriented. Each template performs a small portion of a speci f c functional control. If se veral templates are connected by lines, they form a complete computer program for a speci f c functional control. The template symbol pro vides a visual reminder or memory of the function that the template accomplished. A template maintains its own set of private data and variables that cannot be altered by the action of any other template. The state of an instance of a template is contained in its private variables. They can be only modi f ed by the program code contained in that instance. The private variables of an instance of a template are hidden from the action from other templates. In a template, inputs are used to receive data from other templates or constants. Outputs are used to send data to other templates. Inputs and outputs between two or more templates are connected by lines that direct the data f ow between the templates. An input may recei ve a v alue from only one output in order to prevent overwriting of data at the input. On the other hand, an output data may be sent and used by any number of inputs. For a more comple x control function that needs se veral templates, a combination of templates called a macrotemplate is often used to simplify a graphical programming diagram.

ENERGY MANAGEMENT AND CONTROL SYSTEMS

5.55

Graphical Programming for Mechanical Cooling Control Figure 5.26 shows the graphical programming for mechanical cooling control for a small packaged unit. This packaged unit has a supply fan, a DX coil, a single-scroll compressor, and an economizer. The graphical programming for the control of this packaged unit includes mechanical cooling, economizer, heating, and fan control. During the scheduled occupied time, the zone schedule produces a 1.0 signal, and the cooling control template SP selects the input labeled 1.0 and passes the v alue 74°F (23.3°C) as the cooling set point. During unoccupied periods, SP passes the v alue 85°F (29.4°C). The average temperature of east and west zones is compared with the current set point of the cooling dead-band control template. If the average temperature exceeds the set point plus the dead-band v alue, 0.5°F (0.28°C), or 74.5°F (23.6°C), then the output of the cooling control template goes to 1.0, and the compressor is turned on. If the zone a verage temperature drops belo w the set point minus the dead-band v alue, 73.5°F (23.1°C), then the output of the cooling control template goes to 0.0 and the compressor is turned off. When the outdoor air temperature drops belo w 50 °F (10.0 °C) minus the dead-band v alue 3°F (1.7°C), or 47°F (8.3°C), then the outdoor temperature lock out template produces 1.0, and the compressor is lock ed out (turned of f). Minimum on and minimum of f timer templates pre vent the compressor from operating at short cycles. Graphical programming of DDC systems may become supplemental to the sequence of operation and control diagrams, and may become a part of the DDC design documentation.

5.12 TUNING DDC UNITS Most of the DDC units use PI or PID control modes. Tuning determines the gains and control parameters which have a direct impact on the steady-state error and transient characteristics of the DDC unit and the control system. A well-tuned DDC unit minimizes the steady-state error, or offset from the set point; sho ws a quick response to disturbance; and pro vides operating stability at all operating conditions. HV AC&R processes are nonlinear , and system characteristics change when seasons are varied. The DDC unit tuned at one condition may not be appropriate at other operating conditions.

Tuning PI Controllers For DDC units using PI control modes, proportional gain and inte gral gain should be properly selected. A high gain decreases the control stability and the of fset or error, whereas a low gain will produce a slo w response, which increases the control stability and of fset. In tuning PI controllers, trial and error is often used by the EMCS subcontractor. The trial-and-error method adjusts the gain until the desirable response to a set-point change is sho wn. This response should start with a small overshoot and rapidly damp to steady-state conditions. The trial-and-error method is time-consuming. Other tuning methods for PI controllers, such as closed- and open-loop process identi f cation, are also used. Refer to ASHRAE Handbook 1995, HVAC Applications, for details. Bekk er et al. (1991) recommended a root locus tuning method for f rst-order processes like most temperature and humidity control for PI controllers to achieve a critically damped response.

Self-Tuning PI and PID Controllers During the 1980s, different schemes were de veloped including Astrom and Hagglund ’s (1984) automatic tuning of PI and PID controllers. The tuning procedure suggested by Astrom and Hagglund is based on the identi f cation of one point on the Nyquist curv e of an open-loop system

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with relay feedback. In the earlier 1990s, commercial products of automatic tuning DDC units started to appear in the EMCS in buildings. Based on Astrom and Hagglund ’s principle, Wallenborg (1991) accomplished automatic tuning of PID controllers in supply air temperature and duct static pressure control e xperiments. After automatic tuning, its system performance was improved. The autotuner used in these experiments was different from a “true adaptive controller.” An autotuner is operated with f xed parameters during normal operation. The tuning e xperiment is initiated between the normal operation periods by the operator. The required parameters are calculated from the results of the tuning experiments.

Adaptive Control The controller has the ability to adapt to the control system by determining the optimum PID parameters and adjusting itself accordingly . An adaptive controller, or self-tuning controller , continuously updates its parameter during operation based on some on-line process identi f er and computer programs. Self-tuning adaptive controllers are also available as commercial products now.

5.13 FACTORS AFFECTING CONTROL PROCESSES As def ned in Sec. 5.1, the function of an air conditioning (HV AC&R) control system is to modulate the air conditioning system capacity to match the of f-design condition, load v ariation, and climate change, to maintain the indoor en vironment within desirable limits at optimum ener gy use.

Load The term load refers to the magnitude of the space load, coil load, refrigeration load, or boiler load that determines the amount of the supply air , chilled w ater, or hot w ater needed to control and maintain the controlled variable at the desirable value(s). Load variation and disturbances affect the controlled variable in three circumstances: part-load operation, intermittent operation, and disturbances. Part-Load Operation. The sizes of an air conditioning system and its components are al ways selected according to the magnitude of the load at the design condition, which is often called the design load or full load. In actual operation, air conditioning or HVAC&R systems operate at part load most of the time. Man y air systems e ven spend 85 to 90 percent of their annual operating hours at part-load operation. The load drops below the design load because of 1. Changes in the outdoor climate 2. Changes in the internal loads at the time of operation During part-load operation, the capacity of an air conditioning system must be appropriately reduced so that a desirable indoor environment can be maintained; at the same time, the energy use of the HVAC&R equipment can be saved.

Climate Change Outdoor climate change af fects not only the space load, but also the performance of HV AC&R systems. Outdoor climate parameters for HV AC&R include the dry-b ulb temperature, wet-bulb temperature, solar radiation, wind speed, and wind direction. They are usually speci f ed at design conditions for an HVAC&R system or at a rated condition for certain equipment. Outdoor climate

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changes during off-design conditions. The general trend in summer when the outdoor dry- and wetbulb temperatures fall is that the space cooling load drops accordingly . Because each system has its own characteristics, sometimes it is simpler and more con venient to consider the load ratio and climate parameter as tw o different factors during the calculation and analysis of system performance. For instance, the analysis of the energy use of a water-cooled chiller requires both the load ratio and the entering condenser water temperature (ECWT) to determine the po wer input to the compressor. Intermittent Operation. Many air systems do not operate continuously . Some operate only fe w hours within a diurnal c ycle. Others operate only in the daytime and shut do wn at night when the building is unoccupied. During the w arm-up and cool-do wn periods, space loads v ary a great deal from the design load as well as from that of a continuous part-load operation. The space loads represent transient loads of a dynamic model. Optimum starting and stopping of intermittently operated air systems is important for better indoor environmental control and energy savings. Disturbance. These can be sudden load changes or set-point changes within a short time, say, a fraction of an hour, that affect the controlled variable. The offset of the controlled variable resulting from a disturbance can be eliminated by a DDC system with a proportional-inte gral control mode. For disturbances resulting from a sudden climate change that af fects a 100 percent outdoor air handling unit, or from a sudden switch-on of the spotlights in a conditioned space, a DDC system that incorporates proportional-integral-derivative control mode may be more suitable.

System Capacity First, the capacity of an air conditioning (HV AC&R) system and its components should be adequate. An oversized electric heater produces a greater o vershoot of space temperature and w astes more energy than an appropriately sized heater. An undersized chiller always causes a higher space temperature during the cooling season and the EMCS can become out of control. Second, from the point of vie w of capacity control and ener gy conserv ation, a heating and cooling plant installed with multiple units is always better than a single-unit plant. A piece of equipment will operate more ef f ciently at a higher turndo wn ratio than at a lo wer turndown ratio. The turndown ratio RTD, often called part-load capability , is usually e xpressed as the percentage of design capacity: R TR 

minimum capacity

100 design capacity

(5.20)

Third, modulation of capacity continuously is always better for matching the load to outdoor climate change than tw o-position or step controls. Finally , capacity modulation of lar ge fans, pumps, and compressors using adjustable-frequency variable-speed drives is now often energy-eff cient and cost-effective in many circumstances.

Performance of Control Processes The performance of an EMCS is often represented by the performance of its control processes, or more directly its control loops that perform the control functions. HV AC&R control processes are mostly nonlinear processes whose characteristics cannot be e xpressed by f rst-order equations. Nordeen (1995) listed f ve criteria to use to e valuate the performance of an HV AC&R control process: ●

Stability. A control system is stable if the controlled v ariable does not sho w a continuing trend away from the set point following a disturbance, as shown in Fig. 5.9a. As described previously, a wider throttling range and a slo wer response both are bene f cial to system stability. The decay of

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●

●

●

●

the oscillation of the controlled variable is called damping. The more damped a control system is, the less it oscillates. Response time. The length of time needed for the controlled v ariable to reach the set point after a step change in set point or load is the response time. Overshoot. The difference between the maximum controlled variable and the set point following a change in set point or load is the overshoot. Settling time. The length of time needed for the controlled variable to reach steady-state following a disturbance. Offset. As def ned previously, this is the constant error that e xists between the controlled v ariable and the set point when a steady state is reached.

The problems associated with poor performance include the time lag or dead time, which is lar ge compared to the reaction rate or time constant of a control system; poor measurements of the controlled variables; inappropriate capacity control; and hysteresis within control and control components which create time lag. Time lag or dead time is the delay in time between the change in the controlled variable and when that change is sensed by the sensor , or when the controlled de vice is modulated, or when the capacity of the process plant varies. During the sizing of control v alves and air dampers, a linear relationship between the coil load and the valve stem travel was the primary design consideration during the past decades. Ho wever, many microprocessor-based DDC systems no w permit scaling f actors to be applied to the analog outputs of the DDC units, and thus an inherently nonlinear system will respond in a linear relationship.

Thermal Capacitance Thermal capacitance, which is sometimes called thermal inertia, is related to the mass of the b uilding envelope, equipment, and system components. Capacitance is usually calculated by the product of speci f c heat of the material and the mass of the material. The thermal capacitance of the HVAC&R process and the building envelope affects the performance of a control system as follows. First, the high thermal capacitance of the b uilding envelope, the equipment, chilled and hot w ater, etc., reduce the effect of a disturbance to v ary the controlled v ariable. For example, consider a sudden increase in the lighting load in a conditioned space. As the space air temperature increases because of heat released from the electric lights, a large portion of heat at the same time will be transferred to the b uilding envelope because of the additional temperature dif ference between the space air and the building structure. The increase in the space temperature resulting from the disturbance is thus signif cantly reduced. Second, during the warm-up and cool-down periods, because of the high thermal capacitance of the building envelope, heat exchanger, ducts, and pipes, a greater system capacity and longer period of time are needed to raise or cool down the space air temperature. Third, the thermal inf uence of the weather and of external load changes on the conditions of the indoor space air through the b uilding envelope is rather slo w, often needing se veral hours to reach full effect.

5.14 FUNCTIONAL CONTROLS The control functions of a microprocessor -based EMCS in a lar ge high-rise building become more and more comple x and demanding. To pro vide the required functional controls with satisf actory system performance is always the primary target of an EMCS. In an EMCS, there are four kinds of functional controls: generic, specif c, safety, and diagnostics.

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Generic Controls Generic controls are usually needed by most of the systems, units (equipment), and components. The software of generic controls is usually pro vided in the PC w orkstation and in the system controllers except system graphical displays. The editing of the generic softw are usually takes place in the PC workstation and can be do wnloaded to any one of the system controllers. In man y EMCSs, data and information can be monitored and collected from an y point in the system and analyzed in the w orking station PC and in the system controllers. Graphics displays, trending, totalization, scheduling, alarming, etc., are e xamples of the generic controls. The follo wing are features provided by various manufacturers (or vendors). Graphical Displays. The graphical displays pro vided in the PC w orkstation become the showpiece of an EMCS. Some EMCS manuf acturers demand that the softw are in the operator PC workstation be graphically oriented. The graphical displays include the following: ●

●

●

●

Building f oor plan graphics show the selected f oor plan and the space temperatures. Equipment graphics are pro vided for each major piece of equipment such as packaged unit, airhandling unit, and chiller with status of all points. Schematic graphics show the detailed system drawings which can be created, modif ed, and saved. High-resolution digitized photo-quality displays are also accommodated in some EMCSs.

The graphical display system can allo w a display of up to 10 graphical screens for comparison and monitoring of system status. Graphics combined with color coding, such as temperature that is assigned with different colors at v arious values, will increase viewer’s effectiveness. A graphical display system should provide navigation from a facility map down to a specif c f oor plan. Associated temperature and related parameters are pro vided on the f oor plan. An additional “click” of the mouse on an y questionable area will cause the serving HV AC&R system with its operating information to appear on the screen. Speed and convenience of editing are tw o key issues in graphical displays. If a 486 processor is used, graphics may take 5 to 7 s to paint. Including the na vigating through 5 to 10 screens in search of a problem, too much time will be needed. Because of the use of the Pentium chip, it is possible to reduce the graphical response to 2 to 3 s per screen paint. The operator must create, modify, or edit the graphics easily. Graphical displays play a small role in increasing the ef fectiveness of an y functional control in an EMCS. However, they enable the user to interact more ef f ciently with the data and information from the EMCS and to mak e proper decisions to operate the EMCS and the associated HV AC&R system effectively and eff ciently. Trending. Trending is the ability to pro vide continuous track of certain parameter(s), or operating status of a piece of equipment. The commonly used time interv al for the trend log is once for e very 30 min. Trending is mostly used for troubleshooting. For an example, if the output capacity of a heat exchanger is gradually reduced, scale may form on the heat-exchange surface. Totalization is another trend log which records the total accumulated operated time for each of the units in a plant emplo ying multiple units. Totalization is usually helpful for maintenance and troubleshooting purposes. The trend graph displays the trend data in graphical form. When the user asks for a trend graph, select the point(s). The Y axis of the graph is often automatically scaled; the X axis to indicate time is appropriately labeled. The user can also choose the capacity of trend points, time interv al between two trend data other than the standard 30 min, and selection of both changes of v alue (COV) that e xceed the de f ned v alue and the timed trend. The trend graph is a po werful tool. It is v ery helpful for tuning of PID control loops and analyzing operating problems. Scheduling. Prior to the DDC systems, the use of a time clock to schedule the start and stop of equipment automatically w as most widely adopted. A DDC system usually starts and stops the

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equipment according to the predetermined schedule in the system controllers. They can be in the form of day , week, month, and season schedules, and the users can de f ne the e xecuting periods. The graphical schedules are often used for clari f cation. Modif cations of schedules are easily made at the PC w orkstation and downloaded to the system controllers. Optimum start and stop are actually run optimizing of scheduling. Some manufacturers provide temporary schedules that will be used only once and supplement the existing schedule. Sometimes an o verride schedule which supersedes the e xisting schedule is also used only once. Alarming. This is the softw are for tracking and reporting the alarm conditions, i.e., abnormal conditions. The operator needs to determine the limits or the dif ference from the set points that causes the abnormal condition. The operator also should determine the reactions, logging, printing, displaying messages and graphics, and producing audible announcements that are tak en during an alarm. An alarm that is not ackno wledged by the operator within a speci f c time will mo ve to a higher level of priority. The system must automatically lock out alarms when an alarmed system is turned off and appropriate reactions have been taken. Discriminator Control. Discriminator control is a kind of optimum control which searches for the required lo west cooling supply air temperature and the highest heating supply temperature. This control also minimizes the amount of mixing of cold and w arm supply air, terminal reheat, and resets the cold deck supply temperature to reduce the mechanical cooling and the amount of reheat. Fault Detection and Diagnostics. Fault detection and diagnostics should monitor the operation of HVAC&R equipment, components, and control devices; analyze the data; and identify performance problems to be corrected. Fault detection and diagnostics are discussed in detail in the ne xt section.

Specific Controls Specif c controls including capacity controls are controls for a speci f c function in an HV AC&R system, a unit (or equipment), or a zone. The following are speci f c controls that are discussed in the corresponding sections later: ●

Zone temperature control Fan-coil unit VAV reheat

●

Fan-powered VAV box Air system control Economizer control Discharge air temperature control Minimum ventilation control Duct static pressure control Space humidity control Warm-up or cool-down control

●

Water system control Differential pressure control Chilled water temperature reset Condenser water temperature control Low-temperature heating system control

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Central plant control Multiple-chiller optimizing control Multiple-boiler optimizing control Condenser fan cycling Demand limit control

●

Safety controls

Commissioning and Maintenance A properly designed and installed EMCS needs commissioning to test and tune its controllers and system components according to the design speci f cation. A suff cient, clear, and well-follo wed operations manual and a well-implemented maintenance schedule are k ey factors for an ef fective and eff cient EMCS. Commissioning and maintenance are covered in detail in Chap. 32.

5.15 FAULT DETECTION AND DIAGNOSTICS Basics Modern air -handling units (AHUs), packaged units (PUs), and chillers become more and more complicated because of the IAQ, thermal comfort and energy eff ciency requirements, and the use of DDC. An HVAC&R operator is hardly able to monitor and detect the f ault operations of the AHUs, PUs, and chillers; f nd their causes; and correct them. An automatic fault detection and diagnostic system is important for the ef fective operation and control, for the monitoring and maintenance, and for the optimizing of utilization and continuous impro ving of the HVAC&R systems. As of the late 1990s, fault detection and diagnostics are already a standard component in man y large PUs. In HVAC&R operations, faults occur when the actual measured operating parameters de viate from the normal operating v alues. There are tw o types of detected f aults: complete f ailures and performance degradations. Complete f ailures are abrupt f aults that often cause discontinuation of the operation of a system or component. Symptoms of abrupt f aults can be easily observ ed. Peformance degradation is the result of an evolving fault accumulated during a certain time. The differences between the actually measured v alues of an operating parameter , such as temperature T, °F (°C); pressure p, in. WG (Pa); volume f ow rate V˙ , cfm (m 3 /s); or mass f ow rate m˙ , lb/s (kg /s) and the e xpected values (estimated, simulated, or set points) of temperature Texp, pressure pexp, volume f ow rate V˙exp , or mass f ow rate m˙exp under normal operating conditions are called residual. A fault can be detected by investigating and analyzing residuals. A temperature residual Tres, °F (°C), a pressure residual pres, in. WC (Pa), or a volume f ow rate residual ,V˙res cfm (m3 /s), can be calculated as Tres  T Texp pres  p pexp

(5.21)

V˙res  V˙ V˙exp V˙exp where the subscript exp indicates expected (predicted) values and the units of Texp, pexp, and are the same as those of T, p, and V˙ . Most of the measured operating parameters in a fault detection and diagnostic system are the same monitored parameters (sensed or measured) in an EMCS. Residuals are often normalized so that the dominant symptom may have approximately the same

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magnitude for different types of faults. The residual R can be normalized as R nor 

R R min R max R min

(5.22)

where Rnor  normalized residual Rmax, Rmin  maximum and minimum residuals In the late 1990s, a large rooftop packaged unit made by one U. S. manuf acturer offered a standard fault detection and diagnostic de vice to re view active and historical lists of diagnostic conditions. A total of 49 dif ferent diagnostics can be read at the human interf ace panel, and the last 20 diagnostics can be held in an acti ve history b uffer log at the panel. A human interf ace panel provides 2-line 40-character liquid-crystal display and a 16-b utton keypad for monitoring, setting, controlling, and diagnostics. In the late 1980s, the earlier de velopment of diagnostics of HV AC&R system operation w as mainly rule-based expert systems. During the late 1990s, the development of automating fault detection and diagnosis w as emphasized. Inputs and outputs of an HV AC&R operating process can be mathematically related by using autore gressive models with e xogenous inputs (ARX), artificial neural network (ANN) models, and many other developing models. Both ARX and ANN are called black-box because the y require less physical kno wledge of the operating process. These technologies are e xpected to be commercially a vailable after laboratory and f ield tests in the early 2000s.

Expert System Rule-Based Diagnostics Expert systems are discussed in Sec. 5.10, and diagnostics is one of the applications used in HVAC&R system operations. Bramble y et al. (1998) reported a tool — the outdoor air /economizer (OAE) diagnostician — which monitors the performance of an air -handling unit to pro vide outdoor air as a constant fraction of its supply f ow rate and detects problems of outdoor v entilation air control and economizer operation. An OAE diagnostician uses sensors that are commonly installed for control purposes and diagnoses the operating problems based on rules deri ved from engineering models of proper and improper performance of the AHUs. These rules are implemented in a decision tree structure in computer softw are. The diagnostician collects data periodically from the Building Automation System (BAS) to navigate the decision tree and produce conclusions. At each point of the tree, a rule is e valuted according to the collected data, and the results determine in which branch the diagnosis should be. At the end of the branch, a conclusion is reached corresponding to the current operating condition of the AHU. The operator or installer of the diagnostician enters data only once during setup. The OAE diagnostician w as installed and operated on three AHUs in a ne wly constructed and occupied 200,000-ft2 (18,600-m2) DOE William R. Wiley Environmental Molecular Sciences Laboratory and on four AHUs in a 72,700-ft 2 (6760-m2) Technical Management Center , both in Rishland, Washington. For each AHU, data were recorded hourly from sensors in the B AS for outside temperature, return air temperature, mixed air temperature, discharge air temperature, on/off status of the supply fan, and open/closed status of the chilled water and hot water valves. No sensors were installed specially for the diagnostician; all are used by the BAS for the control purpose. The data were automatically transferred hourly from the B AS to the diagnostician ’s database. The diagnostician then processed these data and produced the diagnostic results that can be viewed on the display. Shortly after initial processing of data, four of the se ven AHUs monitored were found to ha ve problems. The problems included a sensor malfunction, a return damper that was not closed fully, a mixed-air sensor problem, and a chilled w ater control problem. All problems were con f rmed after inspection.

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The diagnostician used color coding on a display to alert the HV AC&R system operator when the problem happened and pro vided assistance in modifying the causes as well as making the correction. The OAE diagnostician has proved itself effective in identifying installation and operating problems during the f eld testing of the outdoor ventilation control and economizer cycle. Dexter and Benouarets (1996) introduced a f ault diagnosis of HV AC&R plant items based on semiqualitative generic reference models. The scheme uses reference models describing f ault-free and fault operations that are collected from data producd by simulating a number of HV AC&R systems (or subsystems) of the same type. Results ha ve showed that a fuzzy-logic-based f ault diagnosis requiring no training on the actual plant can successfully identify faults in a simulated HVAC&R subsystem if the parameter denoting f aults is suff ciently large. The proposed fuzzy method of f ault diagnosis is computationally simple enough to be used for more comple x subsystems, such as AHUs. Stylianou and Nikanpour (1996) present a methodology that uses thermodynamic modeling, pattern recognition, and expert knowledge to determine whether a reciprocating chiller is f ault-free and to diagnose selected f aults. The status of this chiller consists of the of f c ycle, start-up, and steady-state operation. The off cycle deals with those f aults that are more easily detected when the chiller is off. Start-up deals with f aults which are related to refrigerant f ow characteristics and are generally more apparent during the transient period. During steady-state operation, faults of performance deterioration are detected. This methodology requires training data. Data can be collected during commissioning and online measurement of v ariables to b uild up normal steady-state linear re gression models and data collected from manufacturers. This methodology also needs additional treatment to impro ve the def nition of the thresholds to classify emer ging patterns and to establish the range of applicability of these patterns. Further study is required to produce a list of f aults and their associated transient characteristics and steady-state patterns.

ARX and ANN Model-Based Diagnosis Peitsman and Bakk er (1996), Peitsman and Soethout (1997), Lee et al. (1996), and Yoshida et al. (1996) discussed the application of ARX and ANN model-based f ault detection and diagnosis and the related laboratory testing results. Model-based diagnosis is a technique capable of f nding diagnoses based on the beha vior of the system and components. The beha vior is best understood as the interaction of observ ation and prediction e xpressed in the dif ference between the measured (observed) and predicted parameter values as previously covered, called residual. The behavior of the system and components is de f ned using ARX or ANN models. One advantage of a model-based diagnosis over rule-based diagnosis is that the model-based diagnosis does not rely on symptom-f ault patterns. Such patterns are incomplete, and it is diff cult for an expert to anticipate all possible f aults and to predict their symptoms. Another advantage is that a model-based diagnosis can be used for a new system even when there is no repair experience or only when a system model is available. ARX Model. Consider a process in an autoregressive model with exogenous inputs (ARX) having an input signal u(t) and an output signal y(t) and described by a simple input /output relationship as y(t) a 1y(t 1)    a ny(t n)  b1u(t 1)    bmu(t m) e(t)

(5.23)

where t  time e(t)  enter as a direct error a  autoregressive parameter (order n) b  exogenous parameter (order m) The model relating output and input and e xtra (exogenous) variables is linear . The actual performance of a system is nonlinear. However, such an approximation is allowable in fault detection and

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diagnosis. Extra inputs are added, and the use of historical measured inputs ( t 1, t 2, . . . ) will make it more nearly a nonlinear relationship. Procedure to Identify P arameters. After the ARX model has been selected, according to Peitsman and Bakker (1996), the procedure to identify parameters a and b in Eq. (5.23) is as follows: ●

●

●

●

Measure input /output data. Input and output v alues of the process are collected in a data set that consists of measured and design (predicted) data under “healthy” (fault-free) conditions. The measured data sets should contain suff cient dynamic conditions of the process. Model order (m and n) should be selected as the optimum values by trial and error. From the collected input/output data, estimate parameters a and b. Validate the selected model. The accuracy of the model is tested by comparing the predicted value with the measured values.

ANN Models. For an ANN model of three layers with an input layer , a hidden layer, and an output layer, as discussed in Sec. 5.10, the outputs can be calculated as a function of the weighted sum of the inputs, and an additional factor, the biases B as expressed in Eqs. (5.16) and (5.17): oon  V

1 1 e (V B)

i LnWLon

where iLn  input activation to a neuron in hidden layer WLon  weights of connection between hidden layer and outputs ANN models are nonlinear . A learning or training process in an ANN model is necessary. The initially assigned connection weights are adjusted repeatedly during the learning and training process until the error is within speci f ed values, as discussed in Sec. 5.10. After training, an ANN model can predict the outputs from given inputs. Reliability of ARX and ANN Models. Peitsman and Bakker (1996) recognized that the reliability of the ARX and ANN models (the use of mathematic equations to forecast the outputs) is good. One of the conditions needed to produce a reliable model is a healthy and dynamic data set that is measured under optimal and good operating conditions o ver a wide w orking range. Another necessary condition is the a vailability of a f xed time interv al (time step) because the re gression models use values of the previous time interval of the inputs and outputs to predict the output. System and Component Models. In Peitsman and Bakk er (1996), two types of models are produced with the measured and predicted learned data: ●

●

System models. A chiller or an AHU is considered a black box in which multiple independent inputs estimate outputs. Independent v ariables are necessary to form a system model. This model checks the whole system performance. In a system model if the possibility of a malfunction is detected, the next step is to use the component model to localize this malfunction. Component models. Major components are considered as black box es. The conf guration of component models is accomplished after studying the structure of the modeled HV AC&R systems. The model becomes a little bit gray (with some physical kno wledge) instead of black (without physical knowledge). After the system model has detected a malfunction, the component model can pinpoint the cause of the malfunction with a greater accuracy than the system model can.

Fault Detection and Diagnosis. After the parameters (variables) of a system model or component model are simulated using an ARX or ANN model, the predicted outputs of the model must be

ENERGY MANAGEMENT AND CONTROL SYSTEMS

70

20

60

15

Predicted 50

10

5

Detected fault

0

100

200

300

400 Time, s

500

Discharge air temperature Tdis,F

Measured

TresT Texp

Discharge air temperature Tdis,C

5.65

40

600

700

800

FIGURE 5.27 Comparison of the measured and predicted dischar ge air temperatures for a VAV system in an ARX system model. ( Source: Peitsman and Bakk er (1996) ASHRAE Transactions Part I 6637. Reprinted with permission. )

compared with the measured outputs of the system or the components (temperature, pressure, volume f ow) as shown in Fig. 5.27. Only when the measured output is not within the threshold of the predicted output may there be a detected fault or an incorrect model used. The threshold is usually determined from the statistical properties of the process. A residual (margin) of  x x exp   3 is used. Here x indicates the measured v alue, xexp is the assumed predicted or expected mean value, and  is the standard de viation of the predicted v alues. The probability that x falls within [x exp 3, x exp 3] is 0.9973. If  x x exp   3, a detected fault may occur, as shown in Fig. 5.27, in which the difference between the measured and predicted discharge air temperatures is greater than 3 . The possible causes of the f ault include block of chilled w ater f ow, high chilled water temperature, and a fouled cooling coil. Comparison of ARX and ANN Models. According to Peitsman and Bakk er (1996), the comparison of ARX and ANN models for fault detection and diagnosis is as follows: ARX

ANN

Prediction of process variables

Linear

Nonlinear

Results

ANN gives slightly better results than ARX models Shorter Longer

Training period

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5.16 CONTROLS IN ASHRAE/IESNA STANDARD 90.1-1999 ASHRAE/IESNA Standard 90.1-1999, Energy Standard for Buildings Except Low-Rise Residential Buildings, specif es the following mandatory provisions for HVAC&R controls. For exceptions and details refer to Standard 90.1-1999.

General ●

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Thermostatic controls. The supply of heating and cooling ener gy to each zone shall be controlled individually by thermostatic controls using the zone temperature as the response. A dwelling unit is permitted to be considered as a single zone. Dead band. When both heating and cooling are controlled, zone thermostatic controls shall be capable of providing a temperature range or dead band of at least 5 °F (2.8°C). Within that range the heating and cooling energy supplied to the zone is shut off. Set-point o verlap. When heating and cooling supply to a zone are controlled by separate zone thermostatic controls, means such as limit switches, mechanical stops, and software for DDC systems, shall be provided to prevent the heating set point e xceeding the cooling set point minus an y proportional band.

Off-Hour Controls HVAC systems with a design heating or cooling capacity greater than 65,000 Btu /h (19,050 W) and fan system power greater than 3 4 hp (0.56 kW) shall ha ve all the follo wing off-hour controls: automatic shutdown, setback controls, optimum start controls, shutoff damper controls, and zone isolation. ●

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Automatic shutdown. HVAC systems shall be equipped with at least one of the following: 1. Controls that can start and stop the system under dif ferent time schedules, are capable of retaining programming and time setting during loss of po wer for a period at least 10 hours, and include an accessible manual o verride or equi valent that allo ws temporary operation of the system for up to 2 hours. 2. An occupancy sensor that can shut the system of f when no occupant is sensed for a period of up to 30 minutes. 3. A manual operated timer that can adjust the system for two hours. 4. An interlock to a security system which shuts the system of f when the security system is activated. Setback controls. A heating system located where the heating outdoor design temperature is 40 °F (4.4°C) or less shall be equipped with controls that can automatically restart and temporarily operate the system as required to maintain zone temperature abo ve a heating set point adjustable down to 55°F (12.8°C) or lower. A cooling system located where the cooling outdoor design temperature is greater than 100 °F (37.8 °C) shall be equipped with controls that can automatically restart and temporarily operate the system as required to maintain zone temperature belo w a cooling set point adjustable up to 90°F (32.2°C) or higher or to prevent high space humidity. Optimum start contr ols. An air system with a supply v olume f ow rate e xceeding 10,000 cfm (4720 L /s), served by one or more supply f ans, shall have optimum controls. The optimum start shall be at least a function of the difference between space temperature and occupied setpoint, and the amount of time prior to scheduled occupancy. Shutoff damper controls. Both outdoor air supply and e xhaust system shall be installed with motorized dampers that will automatically shut off when the systems or the space are not in use. Out-

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door dampers for v entilation can be automatically shut of f during morning w arm-up and cooldown, or nighttime setback e xcept when v entilation reduces ener gy costs, or ventilation must be supplied. Both outdoor air or e xhaust dampers shall ha ve a maximum air leakage of 3 cfm /ft2 (15 L/sm2) at a pressure dif ference of 1.0 in WC (250 P a). As an e xception, gravity dampers (nonmotorized) are acceptable in buildings less than three stories in height or for b uildings of any height in climates with HDD65 less than 2700. Also excepted are systems with a design outside air intake of 300 cfm (142 L/s) or less that are equipped with motor operated dampers. Zone isolation. HVAC systems serving zones that are intended to operate or be occupied nonsimultaneously shall be di vided into isolation areas. Each isolation area shall be no lar ger than 25,000 ft 2 (2320 m 2) of f oor area nor include more than one f oor. Each isolation area shall be equipped wih isolation de vices that can automatically shut of f the supply of conditoned air , outside air, and exhaust air from the area. Each isolation area can be controlled independently.

REFERENCES Alexander, J., Aldridge, R., and O’Sullivan, D., Wireless Zone Sensors, Heating/Piping/Air Conditioning, no. 5, 1993, pp. 37 – 39. Alley, R. L., Selecting and Sizing outside and Return Air Dampers for VAV Economizer Systems, ASHRAE Transactions, 1988, Part I, pp. 1457 – 1466. Alpers, R., and Zaragoza, J., Air Quality Sensors for Demand Controlled Ventilation, Heating/Piping/Air Conditioning, no. 7, 1944, pp. 89 – 91. Anderson, R., Gems to Look for in EMCS, Heating/Piping/Air Conditioning, no. 11, 1991, pp. 47 – 52. ANSI/ASHRAE, Standard 135 – 1995, BACnet: A Data Communication Protocol for Building Automation and Control Networks, ASHRAE Inc., Atlanta, GA, 1995. Asbill, C. M., Direct Digital vs. Pneumatic Controls, Heating/Piping/Air Conditioning, November 1984, pp. 111 – 116. ASHRAE, ASHRAE Handbook 1995, HVAC Applications, Atlanta, GA, 1995. ASHRAE/IESNA, Standard 90.1 – 1999, Energy Standard for Buildings Except Low-Rise Residential Buildings, ASHRAE Inc., Atlanta, GA, 1999. Astrom, K. J., and Hagglund, T., A New Auto-Tuning Design, IFAC International Symposium on Adaptive Control of Chemical Processes, Copenhagen, 1984. Avery, G., Selecting and Controlling Economizer Dampers, Heating/Piping/Air Conditioning, no. 8, 1996, pp. 73 – 78. Becker, H. P., How Much Sense Do Room Occupancy Sensor Controls Make? ASHRAE Transactions, 1986, Part I, pp. 333 – 342. Bekker, J. E., Meckl, P. H., and Hittle, D. C., A Tuning Method for First Order Processes with PI Controllers, ASHRAE Transactions, 1991, Part II, pp. 19 – 23. Brambley, M., Pratt, R., Chassin, D., Katipamula, S., and Hatley, D., Diagnostics for Outdoor Air Ventilation and Economizers, ASHRAE Journal no. 10, 1998, pp. 49 – 55. Brothers, P., and Cooney, K., A Knowledge-Based System for Comfort Diagnostics, ASHRAE Journal, no. 9, 1989, pp. 60 – 67. Bushby, S. T., and Newman, M., BACnet: A Technical Update, ASHRAE Journal, no. 1, 1994, pp. S72 – S84. Curtiss, P. S., Shavit, G., and Kreider, J. F., Neural Networks Applied to Buildings — A Tutorial and Case Studies in Prediction and Adaptive Control, ASHRAE Transactions, 1996, Part I, pp. 1141 – 1146. Davison, F. G., Direct Digital Control Documentation Employing Graphical Programming, ASHRAE Journal, no. 9, 1992, pp. 46 – 52. Dexter, A. L., and Benouarets, M., A Generic Approach to Identifying Faults in HVAC Plants, ASHRAE Transactions, 1996, Part I, pp. 550 – 556. Elyashiv, T., Beneath the Surface: BACnet Data Link and Physical Layer Options, ASHRAE Journal, no. 11,

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1994, pp. 32 – 36. French, J. C., Object-Oriented Programming of HVAC Control Devices, ASHRAE Journal, no. 12, 1991, pp. 33 – 41. Gibson, G. L., and Kraft, T. T., Electric Demand Prediction Using Artif cial Neural Network Technology, ASHRAE Journal, no. 3, 1993, pp. 60 – 68. Glinke, T. J., The Open Protocol Choice: A Market’s Decision in Process, ACH&R News, Oct. 6, 1997, pp. 6 – 8. Grimm, N. R., and Rosaler, R. C., Handbook of HVAC Design, McGraw-Hill, New York, 1990. Haines, R. W., Proportional plus Integral Control, Heating/Piping/ Air Conditioning, January 1984, pp. 131 – 132. Haines, R. W., Reset Schedules, Heating/Piping/Air Conditioning, September 1985, pp. 142 – 146. Hall, J. D., and Deringer, J. J., Computer Software Invades the HVAC Market, ASHRAE Journal, no. 7, 1989, pp. 32 – 44. Hartman, T. B., Direct Digital Controls for HVAC Systems, McGraw-Hill, New York, 1993. Hayner, A. N., Engineering for Energy Management, Engineered Systems, no. 10, 1993, pp. 23 – 26. Honeywell, Engineering Manual of Automatic Control for Commercial Buildings, Honeywell Inc., Minneapolis, MN, 1988. Huang, S., and Nelson, R. M., Rule Development and Adjustment Strategies of a Fuzzy Logic Controller for an HVAC system, Part One — Analysis and Part Two — Experiment, ASHRAE Transactions, 1994, Part I, pp. 841 – 856. Jafar, M., Bahill, A. T., and Osborn, D. E., A Knowledge-Based System for Residential HVAC Applications, ASHRAE Journal, no. 1, 1991, pp. 20 – 26. Judson, K. W., An Argument for VOC Sensors, Engineered Systems, no. 4, 1995, pp. 34 – 36. Kammers, B. K., Have You Seen BACnet Yet? Engineered Systems, no. 6, 1996, pp. 24 – 44. Kawashima, M., Artif cial Neural Network Backpropagation Model with Three-Phase Annealing Developed for the Building Energy Predictor Shootout, ASHRAE Transactions, 1994, Part II, pp. 1096 – 1103. Kovacs, M., Intelligent Network Sensors, Engineered Systems, no. 9, 1996, pp. 28 – 34. Lee, J., and Russell, D., BACnet: Agent for Change, Engineered Systems, no. 6, 1996, pp. 24 – 40. Lee, Won-Yong, Park, C., and Kelly, G. E., Fault Detection in an Air-Handling Unit Using Residual and Recursive Parameter Identif cation Methods, ASHRAE Transactions, 1996, Part I, pp. 528 – 539. Lehr, V. A., Fuzzy Logic: A Technology and Design Philosophy, Heating/Piping/Air Conditioning, no. 5, 1996, pp. 41 – 46. Miller, R. C., and Seem, J. E., Comparison of Artif cial Neural Networks with Traditional Methods of Predicting Return Time from Night or Weekend Setback, ASHRAE Transactions, 1991, Part II, pp. 500 – 508. Nordeen, H., Fundamentals of Control From a System Perspective, Heating/Piping/Air Conditioning, no. 8, 1995, pp. 33 – 38. Peitsman, H. C., and Bakker, V. C., Application of Black-Box Models to HVAC Systems for Fault Detection, ASHRAE Transactions, 1996, Part I, pp. 628 – 640. Peitsman, H. C., and Soethout, L. L., ARX Models and Real-Time Model-Based Diagnosis, ASHRAE Transactions, 1997, Part I, pp. 657 – 671. Petze, J., Understanding Temperature Sensing Methods and Myths, Heating/Piping/Air Conditioning, November 1986, pp. 193 – 208. Petze, J., Modularity and the Design of Building Automation Systems, Heating/Piping/Air Conditioning, no. 8, 1995, pp. 43 – 46. Robertson, R., and Moult, R., Working Together, Engineered Systems, no.7, 1999, pp. 74 – 79. Schell, M., Making Sense Out of Sensors, Engineered Systems, no. 2, 1996, pp. 108 – 116. Scholten, A., Fuzzy Logic Control, Engineered Systems, no. 6, 1995, pp. 72 – 78. Shams, H., Nelson, R. M., Maxwell, G. M., and Leonard, C., Development of a Knowledge-Based System for HVAC Type Selection, ASHRAE Journal, no. 8, 1995, pp. 165 – 171. Shavit, G., The Evolution of Control during the Past 100 Years, ASHRAE Transactions, 1995, Part I, pp. 538 – 544. Shinn, K. E., A Specif er Guide to BACnet , ASHRAE Journal, no. 4, 1994, pp. 54 – 58.

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Skaer, M., Sensor Savvy, Engineered Systems, no. 7, 1996, pp. 24 – 32. Stylianou, M., and Nikanpour, D., Performance Monitoring, Fault Detection, and Diagnosis of Reciprocating Chillers, ASHRAE Transactions, 1996, Part I, pp. 615 – 627. Swan, B., The Language of BACnet, Engineered Systems, no. 7, 1996, pp. 24 – 32. The Trane Company, Building Control Unit Sizing for Tracer Summit Systems, Engineering Bulletin, American Standard Inc., 1993. The Trane Company, Guide Specif cations ICS-GS-3, American Standard Inc., 1993. Turpin, J. R., The Great Divide, The Experts Offer Their Opinions on Open Systems Standards, Engineered Systems, no. 5, 1996, pp. 35 – 36. Turpin, J. R., Interoperability, Where Art Thou? Engineered Systems, July 1999, pp. 56 – 70. Wallenborg, A. O., A New Self – Tuning Controller for HVAC Systems, ASHRAE Transactions, 1991, Part I, pp. 19 – 25. Yoshida, H., Iwami, T., Yuzawa, H., and Suzuki, M., Typical Faults of Air–Conditioning Systems and Fault Detection by ARX Model and Extended Kalman Filter, ASHRAE Transactions, 1996, Part I, pp. 557 – 564. Zhang, Z. J., Another Look at Traditional Control Valve Sizing Practice, ASHRAE Journal, no. 2, 1993, pp. 38 – 41. Zimmerman, A. J., Fundamentals of Direct Digital Control, Heating/Piping/Air Conditioning, no. 5, 1996, pp. 49 – 59.

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LOAD CALCULATIONS 6.1 SPACE LOAD CHARACTERISTICS 6.2 Space, Room, and Zone 6.2 Convective and Radiative Heat 6.2 Space and Equipment Loads 6.3 Night Shutdown Operating Mode 6.3 Influence of Stored Heat 6.6 6.2 COOLING LOAD AND COIL LOAD CALCULATIONS 6.6 Components of Cooling Load 6.6 Components of Cooling Coil Load 6.7 Difference between Cooling Load and Cooling Coil Load 6.8 Load Profile 6.9 Peak Load and Block Load 6.9 Characteristics of Night Shutdown Operating Mode 6.10 Moisture Transfer from the Building Structures 6.10 6.3 HISTORICAL DEVELOPMENT OF COOLING LOAD CALCULATIONS 6.11 6.4 METHODOLOGY — HEAT BALANCE 6.12 The Physical Model 6.12 Heat Balance Equations 6.12 Heat Balance of Space Air 6.13 Characteristics of Heat Balance Method 6.14 6.5 METHODOLOGY — TRANSFER FUNCTION 6.14 Basics 6.14 Transfer Function and Time Function 6.14 Calculation Procedure 6.15 CLTD/SCL/CLF Method 6.15 TETD/TA Method 6.15 6.6 DETAILED CALCULATION PROCEDURES FOR TFM 6.16 Conduction Heat Gain through Exterior Walls and Roofs 6.16 Heat Gain through Ceilings, Floors, and Interior Partition Walls 6.16 Solar Heat Gain and Conductive Heat Gain through Window Glass 6.17 Internal Heat Gain 6.17 Infiltration 6.24 Cooling Load Conversion Using Room Transfer Function 6.24 Space Cooling Load Calculation 6.25 Heat Extraction Rate and Space Air Transfer Function 6.25 Heat Loss to Surroundings 6.25 6.7 DETAILED CALCULATION PROCEDURE USING CLTD/SCL/CLF METHOD 6.26

Space Cooling Load due to Heat Gain through Exterior Walls and Roofs and Conductive Gain through Glass 6.26 Space Cooling Load due to Solar Heat Gain through Fenestration 6.28 Space Cooling Load due to Heat Gain through Wall Exposed to Unconditioned Space 6.28 Calculation of Internal Cooling Loads and Infiltration 6.29 Space Cooling Load of Night Shutdown Operating Mode 6.32 6.8 COOLING COIL LOAD 6.32 Basics 6.32 Fan Power 6.33 Duct Heat Gain 6.33 Temperature of Plenum Air and Ventilation Load 6.34 6.9 COOLING LOAD CALCULATION BY FINITE DIFFERENCE METHOD 6.34 Finite Difference Method 6.34 Simplifying Assumptions 6.36 Heat and Moisture Transfer at Interior Nodes 6.36 Heat and Moisture Transfer at Surface Nodes 6.37 Space Air Temperature and Cooling Loads 6.38 6.10 HEATING LOAD 6.39 Basic Principles 6.39 Transmission Loss 6.39 Adjacent Unheated Spaces 6.40 Latent Heat Loss and Heat Loss from Products 6.41 Infiltration 6.41 Setback of Night Shutdown Operation 6.41 6.11 LOAD CALCULATION SOFTWARE 6.42 Introduction 6.42 TRACE 600 — Structure and Basics 6.42 TRACE 600 Input — Load Methodology 6.43 TRACE 600 Input — Job 6.44 TRACE 600 Input — External Loads 6.45 TRACE 600 Input — Schedules 6.45 TRACE 600 Input — Internal Loads 6.46 TRACE 600 Input — Partition and Shading Devices 6.47 TRACE 600 — Minimum Input Requirements, Run, and Outputs 6.47 REFERENCES 6.49

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Space, Room, and Zone Space indicates either a v olume or a site without a partition or a partitioned room or group of rooms. A room is an enclosed or partitioned space that is usually treated as a single load. A conditioned room often has an indi vidual control system. A zone is a space, or several rooms, or units of space having some sort of coincident loads or similar operating characteristics. A zone may or may not be an enclosed space, or it may consist of man y partitioned rooms. It could be a conditioned space or a space that is not air conditioned. A conditioned zone is always equipped with an individual control system. A control zone is the basic unit of control.

Convective and Radiative Heat Whether heat enters the conditioned space from an e xternal source or is released to the space from an internal source, the instantaneous heat gains of the conditioned space can be classified into t o categories: convective heat and radiative heat, as shown in Fig. 6.1. When solar radiation strikes the outer surface of a concrete slab, most of its radiative heat is absorbed by the slab; only a small portion is reflected. After absorption, the outer surface temperature of the slab increases. If the slab and the conditioned space are in thermal equilibrium originally , there is then a con vective heat and radiative heat transfer from the surf ace of the slab to the space air and other surf aces. Meanwhile, heat transfer due to conduction tak es place from the surf ace to the inner part of the slab . Heat is then stored inside the slab. The stored heat is released to the space air when the surf ace temperature falls below the temperature of the inner part of the slab.

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FIGURE 6.1 Convective and radiant heat in a conditioned space and the temperatures of the interior surfaces.

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To maintain the preset space air temperature, the heat that has been con vected or released to the conditioned space should be removed from the space instantaneously.

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Space and Equipment Loads The sensible and latent heat transfer between the space air and the surroundings can be classified a follows: 1. Space heat gain qe, in Btu/h (W), represents the rate at which heat enters a conditioned space from an external source or is released to the space from an internal source during a given time interval. 2. Space cooling load, often simply called the cooling load Qrc , Btu / h (W), is the rate at which heat must be removed from a conditioned space so as to maintain a constant temperature and acceptable relative humidity. The sensible cooling load is equal to the sum of the con vective heat transfer from the surfaces of the building envelope, furnishings, occupants, appliances, and equipment. 3. Space heating load Qrh, Btu/h (W), is the rate at which heat must be added to the conditioned space to maintain a constant temperature and sometimes a specified relat ve humidity. 4. Space heat extraction rate Qex, Btu/h (W), is the rate at which heat is actually remo ved from the conditioned space by the air system. The sensible heat e xtraction rate is equal to the sensible cooling load only when the space air temperature remains constant. 5. Coil load Qc, Btu/h (W), is the rate of heat transfer at the coil. The cooling coil load Qcc, Btu/h (W), is the rate at which heat is remo ved by the chilled w ater fl wing through the coil or is absorbed by the refrigerant inside the coil. 6. The heating coil load Qch, Btu/h (W), is the rate at which heat is added to the conditioned air from the hot water, steam, or electric heating elements inside the coil. 7. Refrigerating load Qrl, Btu/h (W), is the rate at which heat is absorbed by the refrigerant at the evaporator. For central hydronic systems, the refrigerating load is the sum of the coil load plus the chilled w ater piping heat gain, pump po wer heat gain, and storage tank heat gain. F or most water systems in commercial b uildings, the water piping and pump po wer heat gain is only about 5 to 10 percent of the coil load. In an air conditioning system using DX coil(s), the refrigerating load is equal to the DX coil load. The instantaneous sensible heat gain of a conditioned space is not equal to the instantaneous sensible cooling load because of storage of part of the radiati ve heat inside the b uilding structures. Such phenomenon results in a smaller instantaneous cooling load than that of the heat gain when it is at its maximum value during a diurnal cycle, as shown in Fig. 6.2. If the space relati ve humidity is maintained at an approximately constant v alue, the storage effect of the moisture in the b uilding envelope and furnishings can be ignored. Then the instantaneous space latent heat gain will be the same as the instantaneous space latent cooling load.

Night Shutdown Operating Mode In many commercial buildings, air systems are often shut do wn during the nighttime or during unoccupied periods. The operating characteristics of the conditioned space in a 24-h diel c ycle can then be divided into three periods, as shown in Fig. 6.3. Night Shutdown Period. This period commences when the air system is switched of f and ends when the air system is switched on again. When the air -handling unit, packaged unit, or terminal unit is turned of f during summer in a hot and humid area, the infiltrated air (through the lift pipe shafts, and window cracks) and any heat transfer from the window glass, external wall, or roofs will cause a sudden increase in indoor temperature Tr of a few degrees Fahrenheit (Celsius). After that,

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FIGURE 6.2 Solar heat gain and heat gain from electric lights and the corresponding space cooling loads for a night shutdown air system.

Tr will rise f arther or drop slightly , depending on the dif ference between outdoor and indoor temperatures. At the same time, the indoor space relative humidity increases gradually because of the infiltra tion of hot and humid outdoor air and the moisture transfer from the wetted surf aces in the air system. Higher space relative humidity causes a moisture transfer from the space air to the b uilding envelopes and furnishings. During winter, after the air system is turned of f, the indoor temperature drops because of the heat loss through e xternal windows, walls, and roofs, as well as the infiltration of outdoor cold ai . Meanwhile, the space relative humidity increases mainly o wing to the fall of the indoor space temperature.

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Cool-down or Warm-up Period. This period commences when the air system be gins to operate and ends when the space temperature or other controlled v ariables ha ve attained predetermined limits. During summer , the supply of cold and dehumidified air after the air system is turned o causes a sudden drop in space air temperature and relati ve humidity. Both heat and moisture are transferred from the building envelope to the space air because of the comparati vely higher temperature and moisture content of the b uilding en velope. These heat and moisture transfers form the cool-down cooling load. Sometimes, the cool-down load can be the maximum summer design cooling load.

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FIGURE 6.3 Operating characteristics of an air conditioned space operated at night shutdo wn mode: (a) summer cooling mode (hot and humid area) and (b) winter mode.

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If refrigeration is used for cooling during the cool-do wn period, then it usually lasts less than 1 h, depending mainly on the tightness of the building and the differences of the space temperatures and humidity ratios between the shutdown and cool-down periods. If the cool-down of the space air and the building envelope is by means of the free cooling of outdoor air , the cool-down period may last for several hours. During winter, the supply of w arm air during the w arm-up period raises the space temperature and lowers the space relative humidity. Because of the higher space temperature, heat transfer from the space air to the colder b uilding envelope and furnishings, and the heat ener gy required to raise the temperature of the building envelope, forms the warm-up heating load. Conditioning Period. This period commences when the space air temperature has f allen or risen to a v alue within the predetermined limits. It ends when the air system is shut do wn. In summer , cold and dehumidified air is usually supplied to the space to o fset the space cooling load and to maintain a required temperature and relati ve humidity . During winter , warm air is supplied to compensate for heat losses from the conditioned space. Space temperature is often controlled and maintained within predetermined limits by control systems. In commercial buildings located in areas with a cold winter , the air system may operate in dutycycling mode during nighttime, unoccupied periods in winter, or intermediate seasons to maintain a night setback temperature. In duty-c ycling mode, the fan and heater turn on and of f to maintain a desired temperature. Space temperature is set back at night, e.g., to 55 or 60°F (12.8 or 15.6°C), to prevent freezing of w ater pipes and to produce a comparati vely smaller temperature lift during a warm-up period. The operation of the air system in a 24-h diel c ycle in winter is then di vided into night setback period, warm-up period, and conditioning period. The required load on the heating coil during a morning warm-up period at winter design conditions is usually the winter design heating load.

Influence of Stored Heat The curve of solar heat gain from a west windo w is shown in Fig. 6.2, as well as the cooling load curve due to this solar heat gain for a conditioned space operated at night shutdo wn mode in summer. The difference between the maximum solar heat gain qhgm and the maximum cooling load Qclm during the conditioning period indicates the amount of heat stored inside the b uilding structures. This dif ference significantly a fects the size of air conditioning equipment required. The amount of heat stored depends mainly on the mass of the b uilding envelope (whether it is hea vy, medium, or light), the duration of the operating period of the air system within a 24-h c ycle, and the characteristics of heat gain, whether radiant heat or con vective heat predominates. F or solar radiation transmitted through a west window, Qclm may have a magnitude of only 40 to 60 percent of qhgm. ASHRAE Handbook divides the mass of the b uilding construction into the follo wing three groups: Heavy construction: approximately 130 lb/ft2 (634 kg/m2) floor are Medium construction: approximately 70 lb/ft2 (342 kg/m2) floor are Light construction: approximately 30 lb/ft2 (146 kg/m2) floor are

6.2 COOLING LOAD AND COIL LOAD CALCULATIONS Components of Cooling Load SH__ ST__ LG__ DF

Cooling load calculations for air conditioning system design are mainly used to determine the v olume fl w rate of the air system as well as the coil and refrigeration load of the equipment — to size

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the HVAC&R equipment and to pro vide the inputs to the system for ener gy use calculations in order to select optimal design alternati ves. Cooling load usually can be classified into t o categories: external and internal.

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External Cooling Loads. These loads are formed because of heat gains in the conditioned space from e xternal sources through the b uilding en velope or b uilding shell and the partition w alls. Sources of external loads include the following cooling loads: 1. 2. 3. 4. 5.

Heat gain entering from the exterior walls and roofs Solar heat gain transmitted through the fenestrations Conductive heat gain coming through the fenestrations Heat gain entering from the partition walls and interior doors Infiltration of outdoor air into the conditioned spac

Internal Cooling Loads. These loads are formed by the release of sensible and latent heat from the heat sources inside the conditioned space. These sources contribute internal cooling loads: 1. People 2. Electric lights 3. Equipment and appliances If moisture transfers from the b uilding structures and the furnishings are e xcluded, only infiltrate air, occupants, equipment, and appliances have both sensible and latent cooling loads. The remaining components have only sensible cooling loads. All sensible heat gains entering the conditioned space represent radiati ve heat and con vective heat e xcept the infiltrated ai . As in Sec. 6.1, radiative heat causes heat storage in the b uilding structures, converts part of the heat gain into cooling load, and makes the cooling load calculations more complicated. Latent heat gains are heat gains from moisture transfer from the occupants, equipment, appliances, or infiltrated ai . If the storage ef fect of the moisture is ignored, all release heat to the space air instantaneously and, therefore, they are instantaneous cooling loads.

Components of Cooling Coil Load If the conducti ve heat gain from the coil’ s framework and the support is ignored, the cooling coil load consists of the follo wing components, as shown in Fig. 6.4 by the summer air conditioning cycle O-m-cc-s-r-rf-m of a constant v olume of supply air , single supply duct, and serving a single zone. 1. Space cooling load Qrc, including sensible and latent load 2. Supply system heat gain qss, because of the supply f an heat gain qsf , and supply duct heat gain qsd 3. Return system heat gain qrs because of heat gains of recessed electric lights and ceiling plenum qrp, of return duct qrd , and return fan qrf , if any 4. Sensible and latent load because of the outdoor v entilation rates Qo to meet the requirements of the occupants and others In Fig. 6.4, the summer air conditioning c ycle O-m-cc-s-r-rf-m consists of an adiabatic mixing process O-m-rf, a cooling and dehumidifying process m-cc, a supply system heat gain process cc-s, a space conditioning process r-s, and a return system heat gain process r-rf. Here, O indicates the status of outdoor air , m the mixture of outdoor air and recirculating air , cc the conditioned air

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FIGURE 6.4 Difference between cooling load and cooling coil load on psychrometric chart.

leaving the coil, s the supply air, r the conditioned space, and rf the recirculating air. All the cooling loads and heat gains are in Btu /h (W). Usually , both the supply and return system heat gains are sensible loads. These components are absorbed by the supply air and return air and appear as cooling and dehumidifying loads at the cooling coil during summer.

Difference between Cooling Load and Cooling Coil Load For the same air conditioning cycle shown in Fig. 6.4, note the following: 1. The space cooling load is represented by Qrc, and the cooling coil load is represented by Qc. Since supply system heat gain qss and return system heat gain qrs are both instantaneous cooling loads, then Qc  Qrc  qss  qrs  Qo

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(6.1)

where Qo  load due to the outdoor ventilation air intake, Btu/h (W). 2. The space cooling load is used to determine the supply v olume fl w rate V˙s, whereas the coil load is used to determine the size of the cooling coil in an air -handling unit or DX coil in a packaged unit. 3. A cooling load component influences both V˙s and the size of the cooling coil, whereas a cooling coil load component may not affect V˙s. 4. Infiltration heat gain is an instantaneous cooling load. From Fig. 6.4 it is apparent that the load due to the outdoor ventilaion air Qo, sometimes called the ventilation load, is a coil load. If Qo is considered a cooling load, the volume fl w rate of the air system will be oversized.

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Load Profile A load prof le shows the variation of space load within a certain time period, such as a 24-h operating c ycle or an annual operating c ycle, as shown in Fig. 6.5. In a load pro f le, the space load is always plotted against time. F or a space cooling load, the magnitude of the curv e is positive; for a space heating load, the magnitude is negative. A load prof le may be used to illustrate the load v ariation of an air conditioned space — a room, a zone, a f oor, a building, or a project. The shape of a load prof le depends on the outdoor climate and, therefore, the latitude, orientation, and structure of the b uilding. It is also af fected by the operating characteristics and the v ariation of the internal loads. The load duration curv e is the plot of number of hours v ersus the load ratio. The load ratio is def ned as the ratio of cooling or heating load to the design full load, both in Btu /h, over a certain period. The period may be a day , a week, a month, or a year. The load duration curv e is helpful in many operating and energy consumption analyses.

Peak Load and Block Load The zone peak load is the maximum space cooling load in a load pro f le of a control zone of the same orientation and similar internal load characteristics calculated according to the daily outdoor dry-bulb and coincident wet-b ulb temperature curv es containing summer or winter outdoor design conditions. For a zone cooling load with se veral components, such as solar load through windo w glass, heat transfer through roofs, or internal load from electric lights, the zone peak load is al ways

FIGURE 6.5 Load prof le, peak load, and block load.

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the maximum sum of these zone cooling load components at a gi ven time. The block load is the maximum sum of the zone cooling loads of v arious load prof les of the control zones of a b uilding f oor or a building at the same time. The block load of a space, room, f oor, or building is the maximum sum of cooling load components in that space, room, f oor, or building at a given time. For air systems, the supply v olume flow rate required for a control zone is calculated based on the zone peak load; and the supply volume flow rate for a specific area, room, floor, or building should be calculated based on the block load (cooling) of this specif ic area, room, floor, or building. F or conditioned space using v ariable-air-volume systems and space air conditioning systems, the required cooling coil load or refrigeration load can be calculated based on block load of the corresponding specific area that air system serves. The heating load is usually the peak heating load of a space, room, zone, f oor, or building in a load prof le. Block load is not needed in heating load calculations because the solar radiation and internal loads are not tak en into account. Design heating load depends entirely on indoor -outdoor temperature difference. Suppose a typical f oor of a multistory b uilding has f ve zones: a north, a south, an east, and a west perimeter zone and an interior zone. For each zone, there is a corresponding daily cooling load prof le curve and a zone peak load QN,13, Qs,12, QE,8, QW,16, and QI,14. Here, subscripts N, S, E, W, and I indicate north, south, east, west, and interior, respectively; 8, 12, 13, 14, and 16 represent time at 8, 12, 13, 14, and 16 h, respectively. The block load of this typical f oor Qmax,t is not calculated as the sum of the zone peak loads QN,13, Qs,12, QE,8, QW,16, and QI,14. Rather the block load is calculated for a speci f c time. For example, block load Qmax.14, Btu/h (W), is the block load at 14 h, i.e., Qmax.14  QN,14  QS,14  QE,14  QW,14  QI,14

(6.2)

Characteristics of Night Shutdown Operating Mode Compared with the continuous operating mode, night shutdown operating mode has the follo wing characteristics: 1. 2. 3. 4.

A greater peak load, therefore a higher air system and initial cost A higher maximum heat extraction rate A smaller accumulated heat extraction rate over the 24-h operating cycle A lower power consumption for the fans, compressors, and pumps

Moisture Transfer from the Building Structures

SH__ ST__ LG__ DF

In hot and humid climates, the frequenc y of outdoor wet-b ulb temperatures higher than 73 °F (22.8°C) exceeds 1750 h during the six warmest consecutive months annually. At the same time, the air inf ltrated through the elevator and pipe shafts during shutdo wn periods often causes a temperature increase of more than 7°F (3.9°C) and a relative humidity increase exceeding 10 percent. These result in a signi f cant increase of the latent load because of the moisture transfer from the b uilding envelope and furnishings to the space air during the cool-do wn period in summer. Such an increase in the coil ’s latent load not only necessitates a greater refrigerating capacity , but also lo wers the sensible heat ratio of the space conditioning process. A space conditioning process remo ves heat and moisture from the space by the supply air , or sometimes adds moisture and heat to compensate for space heat losses. In industrial applications where a v ery low relative humidity is

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maintained in the conditioned space, the latent load due to the moisture transfer from the b uilding envelope should be added to the space cooling load calculations during night shutdo wn operation mode.

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6.3 HISTORICAL DEVELOPMENT OF COOLING LOAD CALCULATIONS In the 1930s, Houghton et al. introduced the analysis of heat transmission through the b uilding envelope and discussed the periodic heat f ow characteristics of the b uilding en velope. In 1937, ASHVE Guide introduced a systematic method of cooling load calculation in volving the division of various load components. In the ASHVE Guide , solar radiation f actors were introduced and their inf uence on e xternal w alls and roofs w as tak en into consideration. Both the windo w crack and number-of-air-changes methods were used to calculate inf ltration. Mackey and Wright first introduced the concept of sol-air temperature in 1944. In the same paper, they recommended a method of approximating the changes in inside surf ace temperature of w alls and roofs due to periodic heat flo w caused by solar radiation and outside temperature with a ne w decrement f actor. In 1952, Mackey and Gay analyzed the dif ference between the instantaneous cooling load and the heat gain o wing to radiant heat incident on the surf ace of the building envelope. In 1964, Palmatier introduced the term thermal storage factor to indicate the ratio between the rate of instantaneous cooling load in the space and rate of heat gain. One year after , Carrier Corporation published a design handbook in which the heat storage f actor and equi valent temperature difference (ETD) were used to indicate the ratio of instantaneous cooling load and heat gain because of the heat storage ef fect of the b uilding structure. This cooling load calculation method w as widely used by many designers until the current ASHRAE methods were adopted. In 1967, ASHRAE suggested a time-a veraging (TA) method to allocate the radiant heat o ver successive periods of 1 to 3 h or 6 to 8 h, depending on the construction of the b uilding structure. Heat gains through w alls and roofs are tab ulated in total equi valent temperature dif ferentials (TETDs). In the same year , Stephenson and Mitalas recommended the thermal response f actor, which includes the heat storage ef fect for the calculation of cooling load. The thermal response factor evaluates the system response on one side of the structure according to random temperature excitations on the other side of the structure. This concept had been de veloped and forms the basis of the weighting f actor method (WFM) or transfer function method (TFM) in the 1970s. In 1977, ASHRAE introduced a single-step cooling load calculation procedure that uses the cooling load factor (CLF) and cooling load temperature dif ference (CLTD); these are produced from the simplif ed TFM. In 1963, Kusuda and Achenbach used a digital computer to analyze the thermal en vironment of occupied underground space. The use of computers to design building mechanical systems was f rst accepted in 1965 when a group of mechanical engineering consultants or ganized Automated Procedures for Engineers Consultants Inc. (APEC) because of sharing of softw are and de velopment costs. Because of the need for computerized load and ener gy calculations, ASHRAE established a Committee on Energy Consumption in 1965 and renamed the Task Group on Energy Requirements (TGER) for Heating and Cooling in 1967. In the mid-1970s, ASHRAE and the National Bureau of Standards (NBS) published the computerized calculation of heating and cooling loads in ener gyestimating programs. In 1980, the U.S. Department of Ener gy sponsored a computer program for energy estimation and load calculation through hour -by-hour detailed system simulation, called DOE-2, which w as published through Los Alamos National Laboratory and La wrence Berk eley Laboratory. In this program, a custom weighting f actor method for v arious room con f gurations is used for heating and cooling load calculations. Man y computerized thermal load and ener gy calculating software programs had been de veloped in the 1980s. Since the 1980s, because of the wide

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adoption of personal computers, the use of computer-aided HVAC&R design was rapidly increased and many thermal load and energy analysis programs were developed in this period.

6.4 METHODOLOGY — HEAT BALANCE The Physical Model The exact method to calculate the space cooling load is to use heat balance equations to determine the temperature of the interior surfaces of the building structure at time t simultaneously and then to calculate the space sensible cooling load, which is equal to the sum of the con vective heat transfer from these surfaces, latent cooling loads, and the cooling load due to inf ltrated air at time t. Consider a typical air conditioned room, as shown in Fig. 6.1. In this room, the building envelope consists of mainly walls, window, ceiling, and f oor. There is also heat transfer from occupants, electric lights, and equipment. The heat transfer between v arious surf aces tak es place under the following assumptions: ●

●

●

●

Only one-dimensional transient heat f ow through the building envelope is considered. The room air is perfectly mix ed with the supply air so that the resulting room air temperature is uniform. The materials of the b uilding envelope are homogeneous. The surf ace temperature, the surf ace heat-transfer coeff cient, and the absorpti vity for each surf ace are uniform v alues. Ref ectivity is very small and can be ignored. The cooling load is calculated based on the mean value of a f xed time interval, such as 1 h.

Heat Balance Equations Any interior surf ace of this air conditioned room may recei ve conduction heat transfer at time t, denoted by qi,t, Btu/h (W), from the adjacent building material. Each interior surface receives shortwave solar radiation from the window glass and longwave radiative heat transfer from other interior surfaces and from the surf aces of the lighting f xtures, appliances, equipment, and occupants. Convective heat transfer is also present between these interior surfaces and room air. The sensible heat balance at the ith surface (for i  1, 2, . . ., m) at time t can be expressed as follows: qi, t  [h ci (Tr, t  Ti, t) 

SH__ ST__ LG__ DF

k

 h ij (Tj, t  Ti, t)] Ai  Sir, t  L ir, t  E ir, t  Oir, t j1

(6.3)

where hci  convective heat-transfer coeff cient of ith surface, Btu/h  ft2 °F (W/m2 °C) hij  radiative heat-transfer coeff cient between interior surface i and j, i  j, Btu/hft2 °F (W/m2 °C) Tr,t  room air temperature at time t, °F (°C) Ti,t  temperature of ith surface at time t, °F (°C) Tj,t  temperature of jth surface at time t, °F (°C) Ai  area of ith surface, ft2 (m2) Sir, t  solar radiation transmitted through window glass and absorbed by ith surface at time t, Btu/h (W) Lir, t  radiative energy from electric lights and absorbed by ith surface at time t, Btu/h (W) Eir, t  radiative energy from equipment and absorbed by ith surface at time t, Btu/h (W) Oir, t  radiative energy from occupants and absorbed by ith surface at time t, Btu/h (W)

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Because the conductive heat and radiative heat in Eq. (6.3) are all continuous functions of time f (t), the transient conductive heat transfer qi,t either can be solved by a partial differential equation using numerical solutions (f nite difference method) or can be solved by means of the conduction transfer function [Eq. (6.10)]. Usually, the room air temperature Tr,t is considered as constant, and Eq. (6.3) and the conduction transfer function equation can be solv ed simultaneously to determine the interior surf ace temperatures Ti, t. Then the sensible cooling load at time t, denoted by Qrs,t, Btu / h (W), can be calculated as Q rs, t 

m

 h ci (Ti, t  Tr, t) Ai  60V˙if,tcpa(To, t  Tr, t)  Sc, t  L c, t  E c, i  Oc, t i1

__RH TX

(6.4)

where V˙if, t  volume f ow rate of inf ltrated air at time t, cfm [m3/(60 s)]   air density, lb/ft3 (kg/m3) cpa  specif c heat of moist air, Btu/lb °F (J/kg°C) To,t  temperature of outdoor air at time t, °F(°C) Sc,t  solar heat coming through windows and convected into room air at time t, Btu/h (W) Lc,t  sensible heat from electric lights and convected into room air at time t, Btu/h (W) Ec,t  sensible heat from equipment and convected into room air at time t, Btu/h (W) Oc,t  sensible heat from occupants and convected into room air at time t, Btu/h (W) Since the latent heat gains convert to latent cooling loads instantaneously, therefore the space latent cooling load at time t, or Qrl,t, Btu/h (W), can be calculated as Q rl, t  qil, t  60V˙if  (wo, t  wr, t)h fg  E l, t  Ol, t

(6.5)

where qil, t  latent heat from ith interior surface and convected into room air at time t, Btu/h (W) wo, t  humidity ratio of inf ltrated air at time t, lb/lb (kg/kg) wr, t  humidity ratio of room air at time t, lb/lb (kg/kg) hfg  latent heat of condensation, Btu/lb (J/kg) El, t  latent heat from equipment at time t, Btu/h (W) Ol, t  latent heat from occupants at time t, Btu/h (W) Then the space cooling load at time t, denoted by Qrc, t, Btu/h (W), is Qrc,t  qrs,t  qrl,t

(6.6)

Heat Balance of Space Air The conditioned air for this room or space is supplied from the ceiling dif fuser at a mass f ow rate at time t of m˙s,t  V˙s, ts. Here, V˙s,t indicates the v olume f ow rate of supply air at time t, cfm (m3/min), and s is the density of supply air , lb/ft3 (kg/m3). The supply air is then mix ed with the space air. The resulting mixture absorbs con vective sensible heat and latent heat from v arious surfaces. At time t, the outdoor air inf ltrates into the space at a volume f ow rate V˙if, t , cfm [m3 /(60 s)]. The sensible heat balance of the space air can be expressed as 60 V˙s, t scpa (Tr, t  Ts,t) 

m

 h ci Ai(Ti, t  Tr, t)  60V˙if, t cpa (To, t  Tr, t)  Sc, t  L c, t  E c, t  Oc, t

i1

(6.7)

where Ts,t  supply air temperature at time t, °F ( °C). If the space air temperature is allo wed to f oat, Eq. (6.3), the transient conductive equation, and Eq. (6.7) must be solved simultaneously.

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The heat balance method is more direct and clear in load calculation methodology . Using the heat balance method, the assumption of linear supposition is not required, and the changing of certain parameters, such as the surf ace convective heat-transfer coeff cient, can be modeled as required. If moisture transfer should be included in the cool-do wn period in nighttime shutdo wn mode in a location where the outdoor climate is hot and humid in summer , then the heat balance method will give comparatively more accurate results. However, the heat balance method demands laborious w ork, more computing time, complicated computer programs, and e xperienced users. Only e xpensive mainframe computers could run computer programs adopting the heat balance methodology in the 1970s and early 1980s. The heat balance method is generally used for research and analytical purposes.

6.5 METHODOLOGY — TRANSFER FUNCTION Basics The transfer function method or weighting f actor method is a simpli f cation of the laborious heat balance method. The wide application of the TFM is due to the user -friendliness of the inputs and outputs of the TFM softw are and the sa ving of computing time. In the transfer function method, interior surf ace temperatures and the space cooling load were f rst calculated by the e xact heat balance method for many representative constructions. The transfer function coeff cients (weighting factors) were then calculated which con vert the heat gains to cooling loads. Sometimes, transfer function coeff cients were also developed through test and experiments. The room transfer function coef f cients (weighting f actors ) were originally generated based on a typical room con f guration of 15 ft  15 ft with 10-ft (4.5 m  4.5 m with 3-m) ceiling and one exposure of 30 percent glass in the early 1970s . In the late 1980s, the introduction of 14 inf uential parameters of zone characteristics by So well (1988) enabled the adopted room transfer function coeff cients to more closely match the room type to be used. Today, TFM is the most widely adopted computer-aided load calculation method in HVAC&R consulting f rms.

Transfer Function and Time Function The transfer function K of an element or a system is the ratio of the Laplace transform of the output Y to the Laplace transform of the input or driving force G, or Y  KG

(6.8)

When a continuous function of time f(t) is represented at regular intervals t and its magnitudes are f (0), f (), f (2), . . ., f (n), the Laplace transform is given as

(z)  f (0)  f ()z 1  f (2)z 2      f (n)z n

(6.9a)

where   time interval, h z  es

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The preceding polynomial in z1 is called the z transform of the continuous function f (t). In Eq. (6.8), Y, K, and G can all be represented in the form of a z transform. Because of the radiative component and the associated heat storage ef fect, the space sensible cooling load at time t can be related to the sensible heat gains and pre vious cooling loads in the form of a continuous function of time f(t), which can be expressed as a z transform.

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Weighting factors are transfer function coef f cients presented in the form of z transform functions. Weighting factors are so called because they are used to weight the importance of current and historical values of heat gain and cooling load on currently calculated cooling loads.

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Calculation Procedure The calculation of space cooling load using the transfer function method consists of tw o steps. First, heat gains or heat loss from e xterior w alls, roofs, and f oors is calculated using response factors or conduction transfer function coef f cients; and the solar and internal heat gains are calculated directly for the scheduled hour. Second, room transfer function coeff cients or room weighting factors are used to convert the heat gains to cooling loads, or the heat losses to heating loads. As described in Sec. 6.2, the sensible inf ltration heat gain is the instantaneous sensible cooling load. All latent heat gains are instantaneous latent cooling loads. The TFM is limited because the cooling loads thus calculated depend on the v alue of transfer function coeff cients as well as the characteristics of the space and ho w they are varied from those used to generate the transfer function coef f cients. In addition, TFM assumed that the total cooling load can be calculated by simply adding the indi vidual components — the superposition principle. However, this assumption can cause some errors.

CLTD/SCL/CLF Method The cooling load temperature dif ference (CL TD)/solar cooling load (SCL) /cooling load f actor (CLF) method f rst calculates the sensible cooling load based on the TFM. The result is divided by the U value, shading coeff cient, or sensible heat gain to generate the CL TD, SCL, or CLF. Thus, it provides a direct, one-step space cooling load calculation instead of a heat gain – cooling load conversion, a tw o-step calculation in TFM. Cooling load calculation using the CL TD/SCL/CLF method can be either computer -aided or performed manually for a check or rough estimate. The CLTD/SCL/CLF method is one of the members of the TFM family. In the CLTD/SCL/CLF method, the CLTD is used to calculate the sensible cooling load for the exterior wall and roofs. Recently, an SCL factor has been added which represents the product of the solar heat gain at that hour and the fraction of heat storage ef fect due to v arious types of room construction and f oor coverings. CLF is used to calculate internal sensible cooling loads. The limitations of the TFM are also carried through to the CL TD/SCL/CLF results. Furthermore, the grouping of CLTD/SCL/CLF may cause additional errors. TETD/TA Method In the total eqi valent temperature dif ference (TETD) /time-averaging (TA) method, heat gains of a number of representative exterior wall and roof assemblies qw, Btu/h (W), are calculated as qw  AU(TETD)

(6.9b)

where A  area of wall or roof, ft2 (m2) U  overall heat-transfer coeff cient of wall or roof, Btu/h ft2 °F (W/m2 °C) In Eq. (6.9b), TETD, in °F (°C), can be evaluated by: ●

●

Using the conduction transfer function as in TFM to determine qw. Then it is di vided by the U value to generate TETD values. Using the following relationship: TETD  Tsol, a  Tr  (Tsol,  Tsol, a )

(6.9c)

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where Tsol, a  daily average sol-air temperature, °F (°C) Tsol,  sol-air temperature at time lag  h, °F (°C)

 effective decrement factor

The internal heat gains and conducti ve heat gain are calculated in the same manner as in the TFM. The radiant fraction of each of the sensible heat gains is then allocated to a period including the current and successive hours, a total of 1 to 3 h for light construction and 6 to 8 h for heavy construction. The TETD/ TA method is also a member of the TFM f amily and is de veloped primarily for manual calculation. TETD/TA is simpler in the conversion of heat gains to cooling loads. However, the time-averaging calculation procedure is subjecti ve — it is more an art than a rigorous scienti f c method. Also the TETD/TA method inherits the limitations that a TFM possesses if the TFM is used to calculate the TETD.

6.6 DETAILED CALCULATION PROCEDURES FOR TFM Most of the widely adopted computer softw are programs for space load calculations are based on the transfer function method, i.e., the weighting f actors method. The follo wing are the detailed calculation procedures for TFM. Conduction Heat Gain through Exterior Walls and Roofs The sensible heat gain through an e xterior wall or a roof using TFM can be calculated by the conduction transfer function. It uses a sol-air temperature at time t, denoted by Tsol, t, to represent the combined temperature excitation of the outdoor temperature and solar heat on the e xterior surface of an e xterior wall or roof, and a constant indoor temperature T r . Conduction-transfer function coeff cients are calculated based on an outdoor surf ace heat-transfer coeff cient ho  3.0 Btu /hft2 °F (17 W/m2 °C), and an indoor surface heat-transfer coeff cient hi  1.46 Btu/hft2 °F (8.3 W/m2 °C). The external heat gain through an e xterior wall or roof at time t, denoted by qe,t, Btu/h (W), can be calculated as qe,t 

bT n0

n sol,t – n



 dn n1

qe,t – n A



 Tr  cn A n0

(6.10)

t  time, h   time interval, h n  summation index of number of terms Tsol,t – n  sol-air temperature at time t n, °F (°C) qe,t – n  conduction heat gain at time tn, Btu/h (W) bn, cn, dn  conduction-transfer function coeff cients; refer to ASHRAE Handbook for details A  interior surface area of wall or roof, ft2 (m2)

where

Harris and McQuiston (1988) provide the conduction-transfer function coeff cients for 41 representive wall assemblies and 42 roof assemblies with variations in components, insulation, and mass. Heat Gain through Ceilings, Floors, and Interior Partition Walls

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If the temperature of the adjacent space Tad, °F (°C), is constant, or the variation of Tad is small, the sensible heat gain due to the heat transfer from the adjacent space through the ceiling, f oor, or interior partition wall at time t, denoted by qp,t, Btu/h (W), can be calculated as qp,t  UA(Tad  Tr)

(6.11)

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where U  overall heat-transfer coeff cient of ceiling, f oor, or partition wall, Btu/h ft2 °F (W/m2 °C) A  surface area of ceiling, f oor, or partition wall, ft2 (m2) If the temperature of the adjacent space Tad,t varies with time, then the sensible heat gain transferring through the ceiling, f oor, or partition w all at time qp, t, Btu/h (W), can be calculated from Eq. (6.10), except qe,t should be replaced by qp,t, qe,t – n replaced by qp,t – n, and Tsol,t replaced by Tad,t. Here qp,t – n indicates the sensible heat gain of the ceiling, f oor, or partition w all at time t  n, Btu/h (W); and Tad,t indicates the adjacent space temperature, °F (°C).

Solar Heat Gain and Conductive Heat Gain through Window Glass Although the inward heat f ow from the solar radiation absorbed by the windo w glass and the heat f ow due to the outdoor and indoor temperature dif ference are actually combined, it is simpler and acceptably accurate to separate this composite heat gain into solar heat gain and conducti ve heat gain; both are sensible heat gains. Solar heat gain qso,t, Btu/h (W), is given as qso,t  As, t (SC)(SHGFt)  Ash, t (SC)(SHGFsh, t)

(6.12)

where As,t  sunlit area of window glass at time t, ft2 (m2) Ash,t  shaded area of window glass at time t, ft2 (m2) SC  shading coeff cient SHGFt  solar heat gain factor at time t considering orientation, latitude, month, and hour, Btu/h  ft2 (W/m2) SHGFsh,t  solar heat gain factor for shaded area at time t, considering latitude, month, and hour, Btu/h  ft2 (W/m2) Usually, the SHGF t of the vertical surface facing north orientation is tak en as SHGF sh,t. Conductive heat gain due to the outdoor-indoor temperature difference qwin,t, Btu/h (W), is given as qwin,t  Uwin Awin (To,t  Tr)

(6.13)

where Uwin  overall heat-transfer coeff cient of window including glass and frame (Table 3.7), Btu/h ft2 °F (W/m2 °C) Awin  gross area of window including glass and frame, ft2 (m2) To,t  outdoor temperature at time t considering month, hour, and location, °F (°C)

Internal Heat Gain The internal sensible heat gain consists of the sensible heat gain from people, and from equipment and appliances.

from electric lights,

People. Human beings release both sensible heat and latent heat to the conditioned space. The radiative portion of the sensible heat gain is about 70 percent when the indoor en vironment of the conditioned space is maintained within the comfort zone. The space sensible heat gain for occupants staying in a conditioned space at time t, denoted by qsp,t, Btu/h (W), can be calculated as qsp,t  Np,t (SHGp) where Np,t  number of occupants in conditioned space at time t SHGp  sensible heat gain of each person, Btu/h (W)

(6.14) __SH __ST __LG DF

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Space latent heat gain for occupants staying in a conditioned space at time t, denoted by qlp,t, Btu /h (W), is given as qlp,t  Np,t (LHGp)

(6.15)

where LHGp  latent heat gain of each person, Btu/h (W). Table 6.1 lists the heat gain from occupants in conditioned space, as abridged from ASHRAE Handbook 1989 , Fundamentals. In Table 6.1, total heat is the sum of sensible and latent heat. The adjusted heat is based on a normally distributed percentage of men, women, and children among the occupants. Lighting. The sensible heat gains from the electric lights depend on the types of installation, as follows: Inside Conditioned Space . For electric lights installed inside the conditioned space, such as light f xtures hung belo w the ceiling, the sensible heat gain released from the electric lights, the emitting element, and light f xtures qs, l is equal to the sensible heat released to the conditioned space qes, l, Btu/h (W); both depend mainly on the criteria of illumination and the type and ef fciency of electric lights and can be calculated as qs.l  3.413 Wlamp Fusl Fal  3.413 WA Afl

(6.16)

where Wlamp  rated input of electric lights, W WA  wattage per ft2 of f oor area, W/ft2 (W/m2) Fusl  ratio of wattage in use to installation wattage In Eq. (6.16), Fal indicates an allo wance factor for light f xtures, such as Ballast losses. F or rapidstart 40-W f uorescent f xtures, Fal varies from 1.18 to 1.3 with a recommended v alue of 1.2 (ASHRAE Handbook 1993, Fundamentals). Recess-Mounted F ixtures Using Return Plenum. For situations in which electric lights are recess-mounted on the ceiling and the ceiling plenum is used as a return plenum, the fraction of the

TABLE 6.1 Rates of Heat Gain from Occupants of Conditioned Spaces*

SH__ ST__ LG__ DF

Degree of activity

Typical application

Total heat of adults, male, Btu/h

Seated at theater Seated at theater Seated, very light work Moderately active off ce work Standing, light work; walking Walking; standing Light bench work Moderate dancing Walking 3 m/h; light machine work Heavy work Heavy machine work; lifting Athletics

Theater — matinee Theater — evening Off ces, hotels, apartments Off ces, hotels, apartments Department store, retail store Drugstore, bank Factory Dance hall Factory Factory Factory Gymnasium

390 390 450 475 550 550 800 900 1000 1500 1600 2000

Total heat adjusted,† Btu/h 330 350 400 450 450 500 750 850 1000 1450 1600 1800

Sensible heat, Btu/h 225 245 245 250 250 250 275 305 375 580 635 710

Latent heat, Btu/h 105 105 155 200 200 250 475 545 625 870 965 1090

*Tabulated values are based on 75°F room dry-bulb temperature. For 80°F room dry-bulb temperature, the total heat remains the same, but the sensible heat values should be decreased by approximately 20 percent and the latent heat values increased accordingly. All values are rounded to nearest 5 Btu/h. † Adjusted heat gain is based on normal percentage of men, women, and children for the application listed, with the postulate that the gain from an adult female is 85 percent of that for an adult male, and that the gain from a child is 75 percent of that for an adult male. Sources: Adapted with permission from ASHRAE Handbook 1989, Fundamentals.

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__RH TX

FIGURE 6.6 Heat released from a recess-mounted ventilated lighting f xture.

sensible heat gain of electric lights that enters the conditioned space qes,l, Btu/h (W), is closely related to the type of lighting f xture, the ceiling plenum, and the return system. If the ceiling plenum is used as a return plenum in a multistory b uilding, as shown in Fig. 6.6, the heat released from recessed f uorescent light f xtures and the heat transfer between the outdoor air, plenum air, return air, and space air are as follows: ●

Radiative and convective heat transfer from the lighting f xture downward directly into the conditioned space qld, Btu/h (W), which is calculated as qld  qs,l  qlp  (1  Flp) qs,l

(6.17)

where qlp  heat released by electric lights to return air, Btu/h (W) Flp  fraction of heat released from light f xture to plenum air The fraction that enters conditioned space qld depends on the v olume f ow rate of the return air f owing through the lighting f xture and the type of f xture. Its magnitude should be obtained from the lighting f xture manufacturer. In Fig. 6.7 is sho wn the relationship between Flp and the intensity of volume f ow rate of return air V˙r / Afl if the lighting f xture is ventilated. Usually, Flp varies between 0.4 and 0.6 for a v entilated f xture in a return air plenum. F or un ventilated lighting f xtures recess-mounted in a return air plenum, Flp varies between 0.15 and 0.5.

__SH __ST __LG DF

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CHAPTER SIX

TX

FIGURE 6.7 Relationship between Flp and the intensity of return air V˙r / A fl.

●

Heat carried away by return air from the ceiling plenum qret, Btu/h (W), which is calculated as qret  60 V˙r r cpa (Tp  Tr)

(6.18)

where vV˙r  olume f ow rate of return air, cfm [m3 /(60 s)] r  density of return air, lb/ft3 (kg/m3) cpa  specif c heat of return air, Btu/lb °F (J/kg°C) Tp, Tr  temperatures of plenum air and space air, °F (°C) ●

Heat transfer from the plenum air to the conditioned space through the suspended ceiling qcl and heat transfer from the plenum air to the conditioned space through the composite f oor qf , both in Btu/h (W), can be calculated as qcl  Ucl Acl (Tp  Tr)

(6.19a)

qf  Ufl Af (Tp  Tr)

(6.19b)

where Ucl, Uf  overall heat-transfer coeff cient of ceiling and composite f oor, Btu/h  ft2 °F (W/m2 °C) Acl, Af  area of ceiling and composite f oor, ft2 (m2) ●

●

Heat transfer between the outdoor air and the plenum air through the e xterior wall of the ceiling plenum qwp,t, Btu / h (W), can be calculated by conductive transfer function as shown in Eq. (6.10). Because qwp,t is comparatively small, it is often ignored. For a multistory building where all the f oors are air conditioned, heat gain from the electric lights that enters the conditioned space, heat to space, qes,l is given as qes,l  qld  qcl  qf

SH__ ST__ LG__ DF

(6.20)

Of this, about 50 percent is radiati ve, and the rest is con vective. Heat released from the electric lights to the return air including radiati ve transfer upw ard, heat to plenum, qlp, Btu/h (W), is

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6.21

__RH TX

given as qlp  qret  qcl  qf

(6.21)

In a return air plenum, precise calculation of the plenum air temperature Tp is rather complicated. A simplif ed method is to assume a steady-state heat transfer between the plenum air and the conditioned space air and to use a cooling load temperature dif ference CLTDwp (CLTDwp can be found from Table 6.2) to calculate the heat transfer through the exterior wall of the plenum. Then, based on the heat balance at the plenum air, qlp  Uwp Awp (CLTDwp)  (60 V˙rrcpa  Ucl Acl  Ufl Afl) (Tp  Tr)

(6.22)

where Uwp  overall heat-transfer coeff cient of exterior wall, Btu/h ft2 °F (W/m2 °C) Awp  area of exterior wall, ft2 (m2) CLTDwp  cooling load temperature difference of exterior wall of ceiling plenum, °F (°C) Plenum air temperature Tp can thus be determined. Surface-Mounted Fixtures under Ceiling. If the lighting f xtures are surface-mounted under the suspended ceiling, then the fraction of heat gain do wnward that enters the conditioned space qes,l, Btu/h (W), can be calculated as qes,l  Fes,l qs,l

(6.23 )

where Fes,l  fraction of sensible heat gain entering the conditioned space for surf lighting f xtures, which varies from approximately 0.8 to 0.95.

ace-mounted

Equipment and Appliances. With equipment and appliances, all energy inputs are con verted to heat ener gy, e.g., in the motor windings, combustion chamber, rubbing surf aces, even at components where mechanical work is performed. A portion of heat released may be e xhausted locally by a mechanical ventilation system. In man y industrial applications, the space sensible heat gain due to the machine load when a motor is located inside the conditioned space qs,e, Btu/h (W), can be calculated as qs,e  2546Php Fload Fuse

1  Fexh

mo

(6.24)

where Php  rated horsepower of machine, hp Fload  load factor indicating ratio of actual power required to rated power Fuse  use factor indicating ratio of actually used equipment and appliance to total installed Fexh  heat removal factor due to mechanical exhaust system mo  motor eff ciency In Table 6.3 are listed the ef f ciencies of motors from a fraction of a horsepo wer to 250 hp (187 kW). The bigger the motor , the higher the motor ef f ciency. A high-eff ciency motor is often cost-effective. The Energy Policy Act of 1992 (EP ACT), after a phase-in period of 5 years which ended on October 24 1997, requires that all co vered motors meet increased minimum ef f ciency levels. If the motor is located outside the conditioned space, then in Eq. (6.24), mo  1. In man y types of equipment and appliances installed with e xhaust hoods in the conditioned space, energy input and Fexh are better determined according to the actual performance of similar projects. For example, for food preparation appliances equipped with e xhaust hoods, only radiation from the surf ace of the appliances, which as a fraction of the total heat input is between 0.1 and 0.5, should be counted; or Fexh is between 0.9 and 0.5. If there is no mechanical e xhaust system, Fexh  0.

__SH __ST __LG DF

RH__

TX

SH__ ST__ LG__

DF 6.22

TABLE 6.2 CLTD for Calculating Sensible Cooling Loads from Sunlit Walls of North Latitude, °F

1

2

3

4

5

6

7

8

9

10

11

12

13

14

15

16

17

18

19

20

21

22

23

24

17 23 30 29 26 33 35 27

10 13 18 17 17 22 23 17

Group C walls: typical, outside 1-in. stucco, 2-in. insulation (5.7lb/ft ), 4-in. concrete, 0.75-in. plaster or gypsum, inside U  0.119 Btu/hft2°F; mass, 63 lb/ft2 3

N NE E SE S SW W NW

15 19 22 22 21 29 31 25

14 17 21 21 19 17 29 23

13 16 19 19 18 25 27 21

12 14 17 17 16 22 25 20

11 13 15 15 15 20 22 18

10 11 14 14 13 18 20 16

9 10 12 12 12 16 18 14

8 10 12 12 10 15 16 13

8 11 14 12 9 13 14 11

7 13 16 13 9 12 13 10

7 15 19 16 9 11 12 10

8 17 22 19 10 11 12 10

8 19 25 22 11 11 12 10

9 20 27 24 14 13 13 11

10 21 29 26 17 15 14 12

12 22 29 28 20 18 16 13

13 22 30 29 22 22 20 15

14 23 30 29 24 26 24 18

15 23 30 29 25 29 29 22

16 23 29 29 26 32 32 25

17 23 28 28 25 33 35 27

17 22 27 27 25 33 35 27

17 21 26 26 24 32 35 27

16 20 24 24 22 31 33 26

22 20 18 19 20 22 22 22

7 10 12 12 9 11 12 10

CH06

Difference in CLTD

Wang (MCGHP)

Facing

Maximum CLTD

39445

Solar time, h

Hours of maxiMinimum mum CLTD CLTD

Group D walls: typical, outside 1-in. stucco, 4-in. concrete, 1- or 2-in. insulation (2 lb/ft3), 0.75-in. plaster or gypsum, inside U  0.119 – 0.20 Btu/hft2°F; mass, 63 lb/ft2 15 17 19 20 19 28 31 25

13 15 17 17 17 25 27 22

12 13 15 15 15 22 24 19

10 11 13 13 13 19 21 17

9 10 11 11 11 16 18 14

7 8 9 10 9 14 15 12

6 7 8 8 8 12 13 10

6 8 9 8 7 10 11 9

6 10 12 10 6 9 10 8

6 14 17 13 6 8 9 7

6 17 22 17 7 8 9 7

7 20 27 22 9 8 9 8

8 22 30 26 12 10 10 9

10 23 32 29 16 12 11 10

12 23 33 31 20 16 14 12

13 24 33 32 24 21 18 14

15 24 32 32 27 27 24 18

17 25 32 32 29 32 30 22

18 25 31 31 29 36 36 27

19 24 30 30 29 38 40 31

19 23 28 28 27 38 41 32

19 22 26 26 26 37 40 32

18 20 24 24 24 34 38 30

16 18 22 22 22 31 34 27

21 19 16 17 19 21 21 22

6 7 8 8 6 8 9 7

19 25 33 32 29 38 41 32

13 18 25 24 23 30 32 25

3 3 4 4 4 5 6 5

2 2 2 2 2 4 5 3

1 1 1 1 1 3 3 2

0 0 0 0 0 1 2 1

1 1 1 1 1 0 1 0

2 9 11 5 0 0 1 0

7 27 31 18 1 2 2 2

8 36 47 32 5 5 5 5

9 39 54 42 12 8 8 8

12 35 55 49 22 12 11 11

15 30 50 51 31 16 15 15

18 26 40 48 39 26 19 18

21 26 33 42 45 38 27 21

Direct applications and adjustments are stated in the text. Source: Abridged with permission from ASHRAE Handbook 1989, Fundamentals.

23 27 31 36 46 50 41 27

24 27 30 32 43 59 56 37

24 26 29 30 37 63 67 47

25 25 27 27 31 61 72 55

26 22 24 24 25 52 67 55

22 18 19 19 20 37 48 41

15 14 15 15 15 24 29 25

11 11 12 12 12 17 20 17

9 9 10 10 10 13 15 13

7 7 8 8 8 10 11 10

5 5 6 6 5 8 8 7

18 9 10 11 14 16 17 18

1 1 1 1 1 0 1 0

26 39 55 51 46 63 72 55

27 40 56 52 47 63 71 55

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N NE E SE S SW W NW

pg 6.22

Group G walls: typical, outside 1-in. stucco, airspace; 1-, 2-, or 3-in. insulation (2 lb/ft3); 0.75-in. plaster or gypsum, inside U  0.081 – 0.78 Btu/hft2°F; mass, 16 lb/ft2

SECOND PASS

N NE E SE S SW W NW

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TABLE 6.3 Heat Gain from Typical Electric Motors Location of motor and driven equipment with respect to conditioned space or airstream Motor nameplate motor or rated horsepower

Full-load Nominal rpm

effciency, percent

0.05 0.125

1,500 1,500

35 35

0.25 0.33 0.50

1,750 1,750 1,750

54 56 60

0.75 1 1 2 3 5 7.5 10 15 20 25 30 40 50 60 75 100 150 200 250

1,750 1,750 1,750 1,750 1,750 1,750 1,750 1,750 1,750 1,750 1,750 1,750 1,750 1,750 1,750 1,750 1,750 1,750 1,750 1,750

72 75 77 79 81 82 84 85 86 87 88 89 89 89 89 90 90 91 91 91

Motor in, driven equipment in, Btu/h

Motor out, driven equipment in, Btu/h

Motor in, driven equipment out, Btu/h

Motor type: shaded pole 360 900

130 320

240 590

640 840 1,270

540 660 850

1,900 2,550 3,820 5,090 7,640 12,700 19,100 24,500 38,200 50,900 63,600 76,300 102,000 127,000 153,000 191,000 255,000 382,000 509,000 636,000

740 850 1,140 1,350 1,790 2,790 3,640 4,490 6,210 7,610 8,680 9,440 12,600 15,700 18,900 21,200 28,300 37,800 50,300 62,900

Motor type: split-phase 1,180 1,500 2,120 Motor type: three-phase 2,650 3,390 4,960 6,440 9,430 15,500 22,700 29,900 44,400 58,500 72,300 85,700 114,000 143,000 172,000 212,000 283,000 420,000 569,000 699,000

For motors operating more than 750 h /year, it is usually cost-effective to use a high-eff ciency motor. Typical eff ciency ratings are as follows: 5 hp, 89.5%; 10 hp, 91.0%; 50 hp, 94.1%; 100 hp, 95.1%; 200 hp, 96.2%. Source: Abridged with permission from ASHRAE Handbook 1989, Fundamentals.

Because of the widespread installation of microcomputers, display terminals, printers, copiers, calculators, and facsimile machines, the heat released from machines in of fice buildings has increased considerably in recent years. K omor (1997) listed the measured of fice plug loads from 17 U.S. of fice buildings in the 1990s which v aried from 0.44 to 1.11 W / ft2. Precise of fice equipment heat gain can be calculated from manuf acturers’ data using a load factor Fload between 0.3 and 0.5. The latent heat gain from the equipment and appliances ql,e, Btu/h (W), can be calculated from the mass f ow rate of water vapor evaporated m˙ w, lb / h (kg / h), as ql,e  1075 m˙ w

(6.25)

__SH __ST __LG DF

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TX

Infiltration

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CHAPTER SIX

Inf ltration is the uncontrolled inw ard f ow of outdoor air through cracks and openings in the b uilding envelope due to the pressure dif ference across the en velope. The pressure dif ference may be caused by any of the following: 1. Wind pressure 2. Stack effect due to the outdoor and indoor temperature difference 3. Mechanical ventilation In summer, for low-rise commercial buildings that have their exterior windows well sealed, and if a positive pressure is maintained in the conditioned space when the air system is operating, normally the inf ltration can be considered zero. For high-rise b uildings, inf ltration should be considered and calculated in both summer and winter. Inf ltration is discussed again in Sec. 20.4. As soon as the v olume f ow rate of in f ltrated air V˙inf , cfm (m 3 /min), is determined, the space sensible heat gain from inf ltration qinf, Btu/h (W), can be calculated as qs, inf  60V˙inf ocpa (To  Tr)

(6.26)

where o  density of outdoor air, lb/ft3 (kg/m3). The space latent heat gain from in f ltration ql, inf, Btu/h (W), can be calculated as ql,inf  60V˙inf o (wo  wr)h fg, 32

(6.27)

where wo, wr  humidity ratio of outdoor and space air, respectively, lb/lb (kg/kg) hfg, 32  latent heat of vaporization at 32°F, Btu/lb (J/kg)

Cooling Load Conversion Using Room Transfer Function The conversion of space sensible heat gains qrs,t, Btu/h (W), having radiative only or radiati ve and convective components to space sensible cooling loads Qrs,t, Btu/h (W), using the transfer function method and room transfer function coeff cients can be expressed as follows: Q rs,t 

 (v0 qs,t  v1qs,t –   v2qs,t – 2    )  (w1 Q rs,t –   w2 Q rs,t – 2    )

i1

(6.28)

where i  number of heat gain components in same group   time interval tn  time at tn Here v0, v1, v2, . . ., w1, w2, . . . are the coeff cients of the room transfer function; refer to the ASHRAE Handbook for details. Their relationship can be expressed from Eqs. (6.8) and (6.9) as K(z) 

SH__ ST__ LG__ DF

v0  v1z 1  v2z 2     1  w1z 1  w2z 2    

(6.29)

The magnitude of the coef ficients depends on the duration of the time interv al, fraction of the radiati ve component, and heat storage capacity because of the 14 influential parameters of zone characteristics, such as zone geometry , height, exterior wall construction, interior shade, furniture, zone location, glass percentage, and type of partition, midfloor, slab, ceiling, roof, and floor.

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Space Cooling Load Calculation The types of conversion of space heat gains to space cooling loads can be grouped into the follo wing two categories: 1. Space sensible cooling loads Qrs,t that are only a fraction of the space sensible heat gain qrs,t. These kinds of sensible heat gains have both radiative and convective components, and it is diff cult to separate the con vective component from the radiati ve component, such as sensible heat gains from exterior walls and roofs, and solar heat gains through windo ws. Equation (6.28) will be used to convert these types of sensible heat gains qrs,t to cooling loads Qrs,t. 2. Space heat gains qin,t are instantaneous space cooling loads Qin,t, both in Btu/h (W), or Qin, t  qin, t

(6.30)

This category includes all latent heat gains qrl, convective sensible heat gains, inf ltration sensible heat gain qinf, s, and sensible heat gains whose con vective component can be separated from the radiative component, such as air-to-air heat gain from windows, lights, and people. The space cooling load Qrc, t, Btu/h (W), is their sum, or Qrc,t  Qrs,t  Qin, t  Ql,t  Qs,t  Ql,t

(6.31)

where Qs,t, Ql,t  space sensible cooling load and latent load, respectively, Btu/h (W). As mandated in ASHRAE/IESNA Standard 90.1-1999, Optimum Start Contr ols, pickup loads either for cooling or heating depend on the dif ference between space temperature and occupied set point and the amount of time prior to scheduled occupanc y, and is often determined by computer software.

Heat Extraction Rate and Space Air Transfer Function In the calculation of the conduction heat gain through the e xterior w all and roofs by using the conduction transfer function [Eq. (6.10)] and then con verting to sensible cooling load from the room transfer function, the space temperature Tr is found to be a constant. Most direct-digital control (DDC) zone control systems are no w adopting proportional-inte gral control mode. When the air system and the DDC system are in operation, Tr is a constant once it has achie ved a stable condition. However, for many air systems operated at nighttime shutdo wn mode, Tr will drift away from the set point during the shutdown period. The space air transfer function relates the heat e xtraction rate at time t, denoted by Qex, t, Btu/h (W), to the space air temperature at time t, denoted by Tr, t, °F (°C ), as 1

2

 pi (Q ex, t –   Q rc,t – i)  i0  gi (Tr,con  Tr, t – i) i0

(6.32)

where pi, gi  space air transfer function coeff cients, refer to ASHRAE Handbook for details Qex,t  heat extraction rate at time t, Btu/h (W) Qrc,t i  calculated space cooling load at time ti, Btu/h (W) Tr, con  assumed constant space air temperature, °F (°C) Heat Loss to Surroundings There is al ways a radiant heat loss from the outer surf ace of the b uilding to the sk y vault without clouds because atmospheric temperature is lo wer at high altitudes, as described in Sec. 3.12. In many locations in the United States, there are also radiant heat losses to the surroundings at

__SH __ST __LG DF

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CHAPTER SIX

nighttime in summer due to the lo wer ground temperature. ASHRAE Handbook 1993 , Fundamentals, recommends a simplif ed calculation procedure so that the sum of sensible cooling loads from heat gain by conduction through e xterior w alls and roofs, from conduction and solar heat gain through windows, and from heat gain through interior partitions, ceilings, and f oors at time t, denoted by Qex, t, Btu/h (W), plus the sensible cooling loads from the radiant component of internal heat gain at time t, denoted by Qin, t, Btu/h (W), will be multiplied by a f actor Fsur to tak e into account these heat losses. The corrected sensible cooling load at time t, denoted by Qrs, cor, t, Btu/h (W), is Qrs, cor, t  Fsur (Qex, t  Qin, t)

(6.33)

The factor of heat loss to surroundings Fsur can be calculated as Fsur  1  0.02 

1  (Uroof Aroof  Uwall Awall  Uwin Awin  Upart Apart) L ex

(6.34)

where Lex  length of space exterior wall, ft U  overall heat-transfer coeff cient (subscript win for window and part for interior partitions), Btu/hft2 °F (W/m2 °C) A  area of component of building envelope, ft2 (m2)

6.7 DETAILED CALCULATION PROCEDURE USING CLTD/SCL/CLF METHOD The follo wing sections describe in greater detail the principles of the CL TD/SCL/CLF method. They also pro vide a simple, manual cooling load calculation procedure in case an estimate or a rough check of computer-aided cooling load calculation is required. Space Cooling Load due to Heat Gain through Exterior Walls and Roofs and Conductive Gain through Glass If the ratio of sensible cooling load to sensible heat gain through the e xterior w all or roof Qrs,w /qrs,w  CLTD/ T, for an exterior sunlit wall or roof under the combined effect of solar radiation and the outdoor-indoor temperature difference, the one-step calculation of space sensible cooling load Qrs,w, Btu/h (W), can be performed as Qrs,w  UA (CLTD)

(6.35)

where U  overall heat-transfer coeff cient of exterior wall or roof, Btu/h ft2 °F (W/m2 °C) A  area of exterior wall, roof, or window including frame or sash, ft2 (m2) CLTD  cooling load temperature difference, °F (°C) The CLTD values recommended by ASHRAE for calculating the space sensible cooling load through f at roofs and sunlit w alls of v arious constructions are listed in Tables 6.2 and 6.4, respectively. The values in both Table 6.2 and Table 6.4 were calculated under the following conditions. In other words, these are the conditions under which the listed data can be applied directly without adjustments: ●

●

SH__ ST__ LG__ DF

●

●

Indoor temperature of 78°F (25.6°C) Outdoor maximum temperature of 95 °F (35°C) with an outdoor daily mean of 85 °F (29.4°C) and an outdoor daily range of 21°F (11.7°C) Solar radiation of 40° north latitude on July 21 Roof with dark, f at surface

39445

TABLE 6.4 CLTD for Calculating Sensible Cooling Loads from Flat Roofs, °F Solar time, h 2

3

4

5

6

7

8

9

10

11

12 13 14 15 16

17 18

19

20

21

22

23

24

Hours of maximum Minimum CLTD CLTD

Maximum Difference CLTD in CLTD

Without suspended ceiling 13

0.093

30 26

23

19 16 13

10

9

8

9

13

17 23 29 36 41

46 49

51

50

47

43

39

35

19

8

51

43

18

0.078

38 36

33

30 28 25

22

20 18

17

16

17 18 21 24 28

32 36

39

41

43

43

42

40

22

16

43

27

1-in. wood with 2-in. insulation 2.5-in. wood with 1-in. insulation 8-in. lightweight concrete 4-in. heavyweight concrete with 2-in. insulation 2.5-in. wood with 2-in. insulation Roof terrace system 6-in. heavyweight concrete with 2-in. insulation 4-in. wood with 2-in. insulation

10

0.083

25 20

16

13 10

7

5

5

7

12

18

25 33 41 48 53

57 57

56

52

46

40

34

29

18

5

57

52

15

0.096

34 31

29

26 23 21

18

16 15

15

16

18 21 25 30 34

38 41

43

44

44

42

40

37

21

15

44

29

33

0.093

39 36

33

29 26 23

20

18 15

14

14

15 17 20 25 29

34 38

42

45

46

45

44

42

21

14

46

32

54

0.090

30 29

27

26 24 22

21

20 20

21

22

24 27 29 32 34

36 38

38

38

37

36

34

33

19

20

38

18

15

0.072

35 33

30

28 26 24

22

20 18

18

18

20 22 25 28 32

35 38

40

41

41

40

39

37

21

18

41

23

77

0.082

30 29

28

27 26 25

24

23 22

22

22

23 23 25 26 28

29 31

32

33

33

33

33

32

22

33

22

11

77

0.088

29 28

27

26 25 24

23

22 21

21

22

23 25 26 28 30

32 33

34

34

34

33

32

31

20

21

34

13

19 20

0.082 0.064

35 34

33

32 31 29

27

26 24

23

22

21 22 22 24 25

27 30

32

34

35

36

37

36

23

21

37

16

CH06

2.5-in. wood with 2-in. insulation 4-in. wood with 2-in. insulation

Wang (MCGHP)

Description of Weight, U value, construction lb/ft2 Btu/hft2 °F 1

With suspended ceiling

SECOND PASS pg 6.27 bcj 5/19/2K

Conditions of direct application and adjustments are stated in the text. Source: Abridged with permission from ASHRAE Handbook 1989, Fundamentals.

6.27

__RH

TX

__SH __ST __LG

DF

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pg 6.28

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CHAPTER SIX

●

●

Outer surf ace R value Ro  0.333 h ft2 °F/Btu (0.06 m 2 °C/W) and inner surf ace Ri  0.685 h ft2 °F/Btu (0.123 m2 °C/W) No attic fans or return air ducts in suspended ceiling space

The following formula can be used for adjustments when the conditions are dif ferent from those mentioned: CLTDcorr  CLTD  78  Tr  Tom  85

(6.36)

where 78  Tr  indoor temperature correction; Tr is indoor temperature, °F (°C) Tom  85  outdoor temperature correction; Tom is outdoor mean daily temperature, °F (°C) In Table 6.4, the roof terrace system includes the following: ●

●

●

●

●

2-in. (50-mm) lightweight concrete Airspace 2-in. (50-mm) insulation [5.7 lb/ft3 (91.2 kg/m3)] 0.5-in. (13-mm) slag or stone 0.375-in. (9.5-mm) felt and membrane

For a pitched roof with a suspended ceiling, the area A in Eq. (6.35) should be the area of the suspended ceiling. If a pitched roof has no suspended ceiling under it, then the actual CLTD is slightly higher than the value listed in Table 6.4 because a greater area is exposed to the outdoor air.

Space Cooling Load due to Solar Heat Gain through Fenestration The space sensible cooling load from solar heat gain transmitted through the windo w facing a specif c direction Qrs, s, Btu/h (W), can be calculated as follows: Qrs, s  Qsun  Qsh  As  SCLs  SC  Ash  SCLsh  SC

(6.37)

where Qsun  space cooling load from solar heat gain through sunlit area of window glass, Btu/h (W) Qsh  space cooling load from solar heat gain through shaded area of window glass, Btu/h (W) As, Ash  sunlit and shaded area, ft2 (m2) SC  shading coeff cient SCLs  solar cooling load for sunlit glass facing specif c direction, Btu/h ft2 (W/m2) SCLsh  solar cooling load for shaded area as if glass is facing north, Btu/h ft2 (W/m2) Zone types for use with SCL s and SCL sh tables, single-story b uilding, are listed in Table 6.5. July solar cooling loads for sunlit glass 40 ° north latitude are listed in Table 6.6. Refer to ASHRAE Handbook 1993, Fundamentals, for zone types for multistory buildings and other details. In Eq. (6.37), at a gi ven time, As  Ash  Aglass. Here Aglass indicates the glass area of the window, in ft 2 (m2). In the northern hemisphere for a conditioned space with southern orientation, the maximum SCLs may occur in December instead of June. Space Cooling Load due to Heat Gain through Wall Exposed to Unconditioned Space SH__ ST__ LG__ DF

When a conditioned space is adjacent to an area that is unconditioned, and if the temperature f uctuation in this area is ignored, then the sensible heat gain qrs transferred through the partitioned w alls and interior windows and doors, in Btu/h (W), can be calculated as

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6.29

TX

TABLE 6.5 Zone Types for Use with SCL and CLF Tables, Single-Story Building Zone parameters

Zone type

No. walls

Floor covering

Partition type

Inside shade

1 or 2 1 or 2 1 or 2 1 or 2 1 or 2 1 or 2 3 3 3 3 3 3 3 4 4 4

Carpet Carpet Vinyl Vinyl Vinyl Vinyl Carpet Carpet Carpet Vinyl Vinyl Vinyl Vinyl Carpet Vinyl Vinyl

Gypsum Concrete block Gypsum Gypsum Concrete block Concrete block Gypsum Concrete block Concrete block Gypsum Gypsum Concrete block Concrete block Gypsum Gypsum Gypsum

* * Full Half to none Full Half to none * Full Half to none Full Half to none Full Half to none * Full Half to none

__RH

Error band

Glass solar

People and equipment

Lights

Plus

A B B C C D A A B B C B C A B C

B C C C D D B B B C C C C B C C

B C C C D D B B B C C C C B C C

9 9 9 16 8 10 9 9 9 9 16 9 16 6 11 19

Minus 2 0 0 0 0 6 2 2 0 0 0 0 0 3 6 1

A total of 14 zone parameters are def ned. Those not shown in this were selected to achieve the minimum error band shown in the right-hand column for solar cooling load. *The effect of inside shade is negligible in this case. Source: Adapted from ASHRAE Handbook 1997, Fundamentals. Reprinted with permission.

qrs  AU(Tun  Tr)

(6.38)

where Tun, Tr  daily mean air temperature of adjacent area that is unconditioned and space temperature, respectively, °F ( °C). F or an adjacent area that is not air conditioned and has heat sources inside, such as a kitchen or boiler room, Tun may be 15°F (8.3°C) higher than the outdoor temperature. F or an adjacent area without an y heat source other than electric lights, Tun  Tr may be between 3 and 8 °F (1.7 and 4.4 °C). For f oors built directly on the ground or located abo ve a basement that is neither v entilated nor conditioned, the space sensible cooling load from the heat gain through the f oor can often be ignored.

Calculation of Internal Cooling Loads and Infiltration The calculation of the internal heat gains of people, lights, equipment, and appliances qint, s and heat gains of in f ltration qinf, all in Btu /h (W), using the CL TD/SCL/CLF method is the same as that which uses TFM. The sensible internal heat gain that contains the radiati ve component is then multiplied by a cooling load f actor CLF int to convert to space sensible cooling load Qint, s, Btu/h (W), and can be calculated as Qint, s  qint, s (CLFint)

(6.39)

In conditioned spaces in which an air system is operated at nighttime shutdo wn mode, CLFint is equal to 1 during the occupied period when the air system is operating. Refer to ASHRAE Handbook 1993, Fundamentals, for CLFint when the air system is operated 24 h continuously. Since internal latent heat gains qint, l are instantaneous internal latent cooling loads Qint, l both in Btu/h (W), Qint,l  qint,l

(6.40)

__SH __ST __LG DF

TX

DF TABLE 6.6 July Solar Cooling Load for Sunlit Glass, 40° North Latitude 6.30

Solar time, h

Glass face

1

2

3

4

5

6

7

8

9

10

11

N NE E SE S SW W NW Horiz.

0 0 0 0 0 0 1 1 0

0 0 0 0 0 0 0 0 0

0 0 0 0 0 0 0 0 0

0 0 0 0 0 0 0 0 0

1 2 2 1 0 0 0 0 0

25 85 93 47 9 9 9 9 24

27 129 157 95 17 17 17 17 69

28 134 185 131 25 24 24 24 120

32 112 183 150 41 30 30 30 169

35 75 154 150 64 35 35 35 211

38 55 106 131 85 39 38 38 241

12

13

14

15

16

17

18

19

20

21

22

23

24

39 40 45 49 84 133 114 50 245

36 37 39 41 63 151 158 84 217

31 32 33 34 42 152 187 121 176

31 26 26 27 31 133 192 143 125

36 18 18 18 20 93 156 130 70

12 7 7 7 8 35 57 46 29

6 3 3 3 4 17 27 22 14

3 2 2 2 2 8 13 11 7

1 1 1 1 1 4 6 5 3

1 0 0 0 0 2 3 3 2

0 0 0 0 0 1 2 1 1

37 45 55 58 79 117 101 46 232

35 41 48 49 63 135 139 76 212

32 36 41 42 46 138 166 108 180

31 30 34 35 37 126 173 128 137

35 23 25 26 27 94 147 119 90

16 13 15 15 16 46 66 51 53

10 9 10 10 11 31 43 33 37

7 6 7 8 8 21 30 22 27

5 5 5 6 6 15 21 16 19

4 3 4 4 4 11 15 11 14

3 3 3 3 3 8 11 8 11

34 43 52 51 70 110 98 44 207

32 40 47 47 54 124 132 73 189

29 36 43 42 40 125 153 102 160

29 31 37 36 33 111 156 118 123

34 25 30 29 26 80 128 107 83

14 16 20 19 16 37 50 39 53

10 13 17 16 13 28 35 26 44

8 11 15 14 12 23 28 21 38

7 10 13 13 10 20 24 17 34

6 9 12 11 9 17 21 15 30

6 8 11 10 8 15 19 13 27

31 44 57 55 63 94 84 41 188

30 42 53 51 52 106 112 64 176

28 39 48 47 41 109 130 87 156

29 35 43 42 36 100 135 101 128

32 29 37 35 30 78 116 94 96

17 22 29 27 22 45 57 42 72

14 19 25 24 19 37 46 33 63

12 17 22 21 17 33 39 29 56

11 15 20 19 15 29 35 25 50

10 14 18 17 14 26 31 22 45

9 12 16 16 12 23 28 20 41

Zone type A

Zone type B 1 1 1 1 1 3 4 3 4

1 2 2 1 1 2 3 2 3

22 73 80 40 8 9 9 9 22

23 109 133 81 15 16 16 16 60

24 116 159 112 21 22 22 22 104

28 101 162 131 36 27 27 27 147

32 73 143 134 56 31 31 31 185

35 58 105 122 74 36 35 34 214

N NE E SE S SW W NW Horiz.

5 7 9 9 7 14 17 12 24

5 6 8 8 7 12 15 11 21

4 6 8 7 6 11 13 10 19

4 5 7 6 5 10 12 9 17

4 6 8 6 5 9 11 8 16

24 75 83 45 12 15 17 14 34

23 106 130 82 18 21 22 20 68

24 107 148 107 23 26 27 25 107

27 88 145 121 36 29 31 29 144

30 61 124 121 54 33 34 32 175

33 49 89 107 70 36 36 34 199

N NE E SE S SW W NW Horiz.

8 11 15 14 11 21 25 18 37

7 10 13 13 10 19 23 16 33

6 9 12 11 9 17 20 15 30

6 8 11 10 8 15 18 13 27

6 9 11 10 7 14 17 12 24

21 63 70 39 12 18 21 17 38

21 87 107 68 17 22 24 21 64

21 90 123 90 21 25 28 24 95

24 77 124 102 32 28 30 27 124

27 58 110 104 46 31 33 30 150

29 49 85 95 59 34 34 32 171

37 52 74 96 86 58 37 37 233

38 48 63 69 87 89 59 37 239

Zone type C 34 47 62 82 79 57 37 36 212

35 45 56 59 79 86 59 36 215

Zone type D 31 48 65 78 67 51 35 33 185

32 46 60 60 69 74 53 34 191

Notes: 1. Values are in Btu/hft2. 2. Apply data directly to standard double-strength glass with no inside shade. 3. Data apply to 21st day of July. 4. For other types of glass and internal shade, use shading coeff cients as multiplier. For externally shaded glass, use north orientation. Source: ASHRAE Handbook 1997, Fundamentals. Reprinted with permission.

bcj 5/19/2K

1 1 1 1 1 4 5 4 5

pg 6.30

2 1 2 2 2 5 6 5 6

SECOND PASS

2 2 2 2 2 6 8 6 8

CH06

N NE E SE S SW W NW Horiz.

Wang (MCGHP)

40 44 53 63 96 101 65 40 259

39445

40 48 67 97 97 64 40 40 257

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pg 6.31

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LOAD CALCULATIONS

6.31

Internal load density (ILD), W/ft2 (W/m2), indicates the total internal heat gains of people, lights, and equipment, and it can be calculated as ILD 

SHGp  LHGp 3.413 Afl

 WA,l  WA,e

__RH TX

(6.41)

where SHGp, LHGp  sensible and latent heat gains for occupants, respectively, Btu/h (W) Af  f oor area, ft2 (m2) WA,l, WA,e  lighting and equipment power density, respectively, W/ft2 (W/m2) Both inf ltration sensible heat gain qinf, s and inf ltration latent heat gain qinf, l, in Btu /h (W), are instantaneous space cooling loads; they also can be expressed as Qinf, s  qinf, s

Qinf, l  qinf, l

(6.42)

where Qinf, s, Qinf, l  inf ltration sensible and latent cooling loads, respectively, Btu/h (W). Example 6.1. A return air plenum in a typical construction and operating characteristics: Wattage of electric lights Return air volume f ow rate Density of return air r Fraction of heat to plenum Flp U values: exterior wall of plenum suspended ceiling f oor Area of ceiling and f oor Area of exterior wall of plenum CLF of electric lights CLTD of exterior wall of plenum Space temperature

f oor of a multistory b uilding has the follo wing

1.5 W/ft2 (16.1 W/m2) 11,800 cfm (334 m3 /min) 0.073 lb/ft3 (1.168 kg/m3) 0.5 0.2 Btu/h ft2 °F (1.136 W/m2 °C) 0.32 Btu/h  ft2 °F (1.817 W/m2 °C) 0.21 Btu/h  ft2 °F (1.192 W/m2 °C) 11,900 ft2 (1106 m2) 1920 ft2 (178 m2) 1.0 24°F (13.3°C) 75°F (23.9°C)

Determine the return air plenum temperature and the space cooling load from the electric lights when they are recess-mounted on the ceiling and the ceiling plenum is used as a return plenum. Solution. From Eq. (6.17), qlp  Flpqs, l  0.5  1.5  11,900  3.413  30,461 Btu/h (W) Since 60V˙r r cpa  60  11,800  0.073  0.243  12,559 Btu / h F, and gi ven Eq. (6.22), the temperature of return air in the return plenum is Tp  Tr 

qlp  Uwp AwpCLTDwp 60V˙r r cpa  Ucl Acl  UflAfl

 75 

30,717  0.2  1920  24 12,559  0.32  11,900  0.21  11,900

 75 

39,933  77.12 F (25.07 C) 18,866

__SH __ST __LG DF

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pg 6.32

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CHAPTER SIX

From Eqs. (6.19 a) and (6.19 b), heat transfer from the plenum air to conditioned space through the ceiling and f oor is calculated as

TX

qcl  Ucl Acl (Tp  Tr)  0.32  11,900(77.1  75)  7997 Btu / h (W) qfl  Ufl Afl (Tp  Tr)  0.21  11,900(77.1  75)  5248 Btu / h (W) Then, from Eq. (6.20), heat to space is calculated as qes, l  qld  qcl  qf  30,461  7997  5248  43,706 Btu/h (W) From Eq. (6.39), the space sensible cooling load from electric lights is Qrs, l  CLFint qes, l  1.0  43,692  43,692 Btu/h (12,802 W) Space Cooling Load of Night Shutdown Operating Mode In commercial buildings, air systems are often operated in night shutdown mode during unoccupied hours in summer . The accumulated stored heat because of the e xternal heat gains increases the space cooling load during cool-down and conditioned periods. On the other hand, heat losses to the surroundings due to the radiant heat exchange between the outer surface of the building and the sky vault and surroundings decrease the accumulated stored heat as well as the space cooling load, although the radiative heat losses to the sky vault and surroundings partly compensate the stored heat released to the space. Ho wever, overlooking the remaining stored heat released to the space during cool-down and conditioned periods in summer is the limitation of the CL TD/SCL/CLF method compared to TFM, especially when peak load occurs during the cool-do wn period. As in TFM, an increase of up to 10 percent is recommended by ASHRAE/IES Standard 90.1-1989 for pickup load during the cool-down period for air systems operated at nighttime shutdown mode.

6.8 COOLING COIL LOAD Basics Based on the principle of heat balance, the cooling coil load is given as Total enthalpy of entering air  total enthalpy of leaving air  cooling coil load (or cooling capacity)  heat energy of condensate Since the heat ener gy of the condensate is small and can be ignored, Btu/h (W), can be calculated as Q cc  60V˙s s (h aeh cc)

the cooling coil load Qcc, (6.43)

where vV˙s  olume f ow rate of supply air, cfm [m3 /(60 s)] s  density of supply air, lb/ft3 (kg/m3) hae, hcc  enthalpy of entering air and conditioned air leaving coil, respectively, Btu/lb (J/kg) SH__ ST__ LG__ DF

Of this, the sensible cooling coil load Qcs, Btu/h (W), is Q cs  60V˙s s cpa(Tae  Tcc)

(6.44)

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LOAD CALCULATIONS

6.33

where Tae, Tcc  temperature of entering air and conditioned air lea ving coil, respectively, °F (°C). And the latent coil load Qcl, Btu/h (W), is Q cl  60V˙s s (wae  wcc)h fg, 32

__RH TX

(6.45)

where wae, wcc  humidity ratio of entering air and conditioned air lea ving coil, respectively, lb/lb (kg/kg). Also, Qcc  Qcs  Qcl

(6.46)

From Fig. 6.4, alternatively, the sensible cooling coil load can be calculated as Qcs  Qrs  qs, s  qr, s  Qo, s

(6.47)

where qs, s, qr, s  supply and return system heat gain (as mentioned in preceding section, both are instantaneous cooling loads), Btu/h (W) Qo, s  sensible load from outdoor air intake, Btu/h (W) And the latent coil load can be calculated as Qcl  Qrl  Qo, l

(6.48)

where Qo, l  latent load from outdoor air intak e, Btu/h (W). The supply system heat gain consists of mainly the supply f an power heat gain qsf and supply duct heat gain qsd; and the return system heat gain comprises the return f an power heat gain qrf, return duct heat gain qrd, and ceiling plenum heat gain qrp, all in Btu/h (W). Fan Power In the air duct system, the temperature increase from the heat released to the airstream because of frictional and dynamic losses is nearly compensated by the e xpansion of air from the pressure drop of the airstream. Therefore, it is usually assumed that there is no signi f cant temperature increase from frictional and dynamic losses when air f ows through an air duct system. Fan power input is almost entirely con verted to heat ener gy within the f an. If the f an motor is located in the supply or return airstream, the temperature increase across the supply (or return f an) Tf, °F (°C), can be calculated as Tf 

0.37pt f m

(6.49)

where pt  fan total pressure, in. WC f , m  total eff ciency of fan and eff ciency of motor If the motor is located outside the airstream, then, in Eq. (6.49), m  1. The pt of the return fan for a central hydronic air conditioning system in commercial b uildings is usually 0.25 to 0.5 of the pt of the supply fan. Therefore, the temperature increase of the return fan is f ar smaller than that of the supply f an. The temperature increase of the relief f an or e xhaust fan affects only the relief or e xhaust airstream. It is not a part of the supply and return system heat gain. A relief f an is used to relie ve excess space pressure when 100 percent outdoor air is f owing through the supply fan for free cooling. Duct Heat Gain Duct heat gain is the heat transfer caused by the temperature difference between the ambient air and the air f owing inside the air duct. Duct heat gain is af fected by this temperature dif ference, the

__SH __ST __LG DF

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pg 6.34

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3/7/2k

CHAPTER SIX

thickness of the duct insulation layer , air v olume f ow rate, and the size and shape of the duct. Detailed calculations are presented in Secs. 17.4 and 20.16. A rough estimate of the temperature increase of the supply air for an insulated duct is as follows: Supply air velocity

Air temperature rise

 2000 fpm (10 m/s)  2000 fpm (10 m/s)

1°F/100 ft (0.6°C/30 m) duct length 0.75°F/100 ft (0.45°C/30 m) duct length

Temperature of Plenum Air and Ventilation Load For a ceiling plenum using a return plenum, the plenum air temperature can be calculated from Eq. (6.22). The temperature increase of the plenum air, caused by the heat released from the electric lights ( Tp  Tr), is affected by their po wer input, type of lighting f xture, return air v olume f ow rate, and construction of the ceiling plenum. The temperature increase of plenum air Tp  Tr is usually between 1 and 3°F (0.6 and 1.7°C). From Eqs. (6.26) and (6.27), the sensible and latent loads Qo, s and Qo, l, Btu/h (W), which are attributable to the outdoor air intak e, can be similarly calculated, except V˙inf in Eqs. (6.26) and (6.27) should be replaced by the volume f ow rate of outdoor air V˙o, cfm (m3 /min). System heat gains are mainly due to con vective heat transfer from the surf aces. For simplif cation, they are considered instantaneous cooling coil loads.

6.9 COOLING LOAD CALCULATION BY FINITE DIFFERENCE METHOD When both heat and moisture transfer from the surf ace of the w alls, ceiling and carpet or f oors should be considered in the space cooling load calculation during the cool-do wn period in summer in a location where the outdoor climate is hot and humid, the f nite difference method might be the best choice.

Finite Difference Method Because of the rapid increase in the use of microcomputers in the HV AC&R calculations, it is now possible to use a f nite dif ference method, a numerical approach, to solv e transient simultaneous heat- and moisture-transfer problems in heating and cooling load calculations and ener gy estimations. The f nite difference method divides the building structures into a number of sections. A f ctitious node i is located at the center of each section or on the surf ace, as shown in Fig. 6.8. An energy balance or a mass balance at each node at selected time interv als results in a set of algebraic equations that can be emplo yed to determine the temperature and moisture for each node in terms of neighboring nodal temperatures or moisture contents, nodal geometry, and the thermal and moisture properties of the b uilding structure. The stored heat ener gy and moisture are e xpressed as an increase of internal energy and moisture content at the nodes. Heat conduction can be approximated by using the f nite difference form of the F ourier law, as qi1:i  SH__ ST__ LG__ DF

kAi (T ti1  T t1) x

where k  thermal conductivity, Btu/ h ft °F (W/m °C) A i  area of building structure perpendicular to direction of heat f ow, ft2 (m2)

(6.50)

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6.35

__RH TX

FIGURE 6.8 Building structures and nodes for a typical room.

__SH __ST __LG DF

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pg 6.36

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TX

T  temperature, °F (°C) x  spacing between the nodal points, ft (m) In Eq. (6.50), superscript t denotes at time t. Each nodal equation is solved explicitly in terms of the future temperature of that node. The explicit method is simpler and clearer than the more comple x implicit method. The time derivative is then approximated by a forw ard f nite difference in time, or  T ti T tt Ti i  t t Compared with the transfer function method to calculate the cooling load, method has the following benef ts: 1. 2. 3. 4.

(6.51) the f nite dif ference

It solves heat- and moisture-transfer load calculations simultaneously. The concept and approach are easily understood. It permits custom-made solutions for special problems. It allows direct calculation of cooling loads and energy estimates.

Its drawbacks are mainly due to the great number of computerized calculations and comparati vely fewer computer programs and less information and experience are available.

Simplifying Assumptions When the f nite difference method is used to calculate space cooling loads, simplif cations are often required to reduce the number of computer calculations and to solv e the problem more easily . The errors due to simplif cation should be within acceptable limits. For a typical room in the b uilding, as sho wn in Fig. 6.8, the follo wing are the simplifying assumptions: ●

●

●

●

●

●

●

●

Heat and moisture f ow are one-dimensional. Thermal properties of the building materials are homogeneous. The properties of the airstream f owing over the surface of the building structures are homogeneous. The surf ace temperature dif ferences between the partition w alls, ceiling, and f oors are small; therefore, the radiative exchange between these surfaces can be ignored. The radiative energy received on the inner surf ace of the b uilding structures can be estimated as the product of the shape f actor and the radiati ve portion of the heat gains, and the shape f actor is approximated by the ratio of the receiving area to the total zone area. During the operating period, the heat capacity of the space air is small compared with other heat gains; therefore, it can be ignored. When the air system is not operating during the night shutdo wn period, the heat capacity of the space air has a signif cant inf uence on the space air temperature; therefore, it should be taken into account. Different heat- and mass-transfer coef f cients and analyses are used for the operating period and shutdown periods.

Heat and Moisture Transfer at Interior Nodes SH__ ST__ LG__ DF

Consider an interior node i as shown in the upper part of Fig. 6.8. F or a one-dimensional heat f ow, if there is no internal energy generation, then according to the principle of heat balance,

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6.37

TX

Conduction heat from node i  1  conduction heat from node i  1  rate of change of internal energy of node i qi1:i  qi1:i  

__RH

Ui t

bcb Ai x (T tt  T ti) i t

(6.52)

where Ui  internal energy of node i, Btu/lb (J/kg)  b , cb  density and specif c heat of building material, lb/ft 3 (kg/m 3 ) and Btu/lb °F (J/kg°C) t  selected time interval, s or min Substituting Eq. (6.50) into (6.52) and solving for Tit  t, we have T tt  Fo(T ti1  T ti1)  (1  2Fo)T ti i

(6.53)

In Eq. (6.53), Fo is the Fourier number and is def ned as Fo 

kb t bcb (x)2

(6.54)

Subscript b indicates the building material. The choice of spacing x and the time interv al t must meet some criterion to ensure con vergence in the calculations. The criterion is the stability limit, or Fo  12

(6.55)

Similarly, for moisture transfer at the interior nodes, X tt  Fo mass (X ti1  X ti1)  (1  2 Fo mass) X ti i

(6.56)

and Fo mass 

Dlv t  12 (x)2

(6.57)

where X  moisture content, dimensionless Dlv  mass diffusivity of liquid and vapor, ft2 /s (m 2 /s) Heat and Moisture Transfer at Surface Nodes For a one-dimensional heat f ow, the energy balance at surface node i of the partition wall as shown in Fig. 6.8 is Conductive heat from node i  1  convective heat transfer from space air  latent heat of moisture transfer from space air  radiative heat from internal loads  rate of change of internal energy of node i kAi (T ti1  T ti)  h ci Ai(T tr  T ti)  ah mi Ai X ti (w tr  w tis) h fg x  F1:i L r  Fp:iOr  Fm:i M r  bcb Ai

 T ti x T tt i 2 t

__SH __ST __LG DF

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pg 6.38

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CHAPTER SIX

hci  convective heat-transfer coeff cient of surface i, Btu/hft2 °F (W/m 2 °C) a  density of space air, lb/ft 3 (kg/m 3) hmi  convective mass-transfer coeff cient of surface i, ft/s (m/s) wtis  humidity ratio corresponding to surface i at time t, lb/lb (kg/kg) h f g  latent heat of vaporization at surface temperature, Btu/lb (J/kg) F1:i, Fp:i, Fm:i  shape factor between surface of lights, occupants, appliances and surface i L r ,Or, Mr  radiative portion of heat energy from lights, occupants, and appliances and machines, Btu/ h (W)

where

Solving for T itt gives





Titt  2Fo T ti1  Bi T tr  

a h mi h fgX ti (w tr  w tis ) h ci

Fp:i Or Fm:i M r F1:i L r   h ci Ai h ci Ai h ci Ai

  [1  2Fo(1  Bi)]T

t i

(6.58)

In Eq. (6.58) Bi is the Biot number and can be expressed as Bi 

h c x k

(6.59)

The stability limit of the surface nodes requires that (6.60)

Fo(1  Bi)  1/2

Similarly, according to the principle of conservation of mass, the moisture content at surface node i is given as X tt  (1  2Fo mass)X ti  2Fo mass X ti1  i

2ah mi t Xti (w tr  w tis)

b x

(6.61)

The temperature and moisture content at other surf ace nodes such as the f oor, ceiling, exterior walls, glass, and Plexiglas of the lighting f xture can be found in the same manner. According to ASHRAE Handbook 1993, Fundamentals, the radiative and convective portions of the heat gains are as follows:

Fluorescent lights People External walls and roofs Appliance and machines

Radiative, percent

Convective, percent

50 33 60 20 – 80

50 67 40 80 – 20

Space Air Temperature and Cooling Loads If the inf ltrated air is ignored, the heat balance on the space air or the plenum air can be described by the following relationship: Internal energy of supply air  convective heat transfer from building structures  convective heat transfer from internal loads  internal energy of space air SH__ ST__ LG__ DF

60V˙sscpaT tt  s

n

 T tt )  L c  Oc  M c  60V˙ssc T tt i r r  h ck Ak(T tt k1 pa

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Solving for Trt  t, we have  60V˙sscpaT tt s T tt  r

n

 L c  Oc  M c i  h ck AkT tt k1

60V˙sscpa 

n



k1

(6.62)

h ck Ak

where ,V˙s Ts  volume f ow and temperature of supply air, cfm (m3 /(60 s)) and °F (°C) Lc, Oc, Mc  convective heat from lights, occupants, and appliances, Btu/h (W) The temperature of the return plenum air can be similarly calculated. loads, therefore, can be calculated as Q tt  rs

The space sensible cooling

n

 T tt )  L c  Oc  M c i r  h ck Ak(T tt k1

(6.63)

The latent heat gains are instantaneous latent cooling loads.

6.10 HEATING LOAD Basic Principles The design heating load, or simply the heating load, is always the maximum heat energy that might possibly be required to supply to the conditioned space at winter design conditions to maintain the winter indoor design temperature. The maximum heating load usually occurs before sunrise on the coldest days. The following are the basic principles of heating load calculation that are dif ferent from those for the cooling load calculation: ●

●

●

All heating losses are instantaneous heating loads. The heat storage ef fect of the b uilding structure is ignored. Solar heat gains and the internal loads are usually not tak en into account except for those internal loads Qin, Btu/h (W), that continuously release heat to the conditioned space during the operating period of the whole heating season. Only that latent heat Ql , Btu/h (W), required to evaporate liquid water for maintaining necessary space humidity is considered as heating load.

For a continuously operated heating system, the heating load Qrh, Btu/h (W), can be calculated as Q rh  Q tran  Q if,.s  Q l  Q mat  Q in

(6.64)

where Qtran  transmission loss, Btu/h (W) Qinf, s  sensible heat loss from inf ltrated air, Btu/h (W) Qmat  heat added to entering colder product or material, Btu/ h (W) Transmission Loss Transmission loss Qtran, Btu/h (W), is the sum of heat losses from the conditioned space through the e xternal w alls, roof, ceiling, f oor, and glass. If the calculation is simpli f ed to a steady-state heat f ow, then Q tran   AU(Tr  To)

(6.65)

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where A  area of walls, roof, ceiling, f oor, or glass, ft2 (m2) U  overall heat-transfer coeff cient of walls, roof, ceiling, f oor, or glass, Btu/h  ft2 °F (W/m2 °C) When the winter outdoor design temperature is used for To, the heat loss calculated by transient heat transfer will be less than that from Eq. (6.65) because of the c yclic f uctuations of the outdoor temperature and the heat storage in the building structures. For concrete slab f oors on a grade, heat loss Qf , Btu/h (W), is mostly through the perimeter instead of through the f oor and the ground. It can be estimated as Q fl  PCfl(Tr  To)

(6.66)

where P  length of perimeter, ft (m) Cf  heat loss coeff cient per foot (meter) of perimeter length, Btu/h ft °F (W/m °C) For areas ha ving an annual total of heating de gree-days HDD65  5350 and for a concrete w all with interior insulation in the perimeter ha ving an R value of 5.4 h  ft2 °F/Btu (0.97 m 2 °C/W), Cf  0.72 Btu / h ft °F (1.25 W/m °C). Refer to ASHRAE Handbook 1989, Fundamentals, chapter 25, for more details. For basement walls, the paths of the heat f ow below the grade line are approximately concentric circular patterns centered at the intersection of the grade line and the basement w all. The thermal resistance of the soil and the w all depends on the path length through the soil and the construction of the basement wall. A simplif ed calculation of the heat loss through the basement w alls and f oor Qb,g, Btu/h (W), is as follows: Q b, g  Ab, gUb, g(Tbase  To)

(6.67)

where Ab,g  area of basement wall or f oor below grade, ft2 (m2) Ub,g  overall heat-transfer coeff cient of wall or f oor and soil path, Btu/h  ft2 °F (W/m 2 °C) The values of Ub,g are roughly given as follows:

Uninsulated wall Insulated wall Basement f oor

0 to 2 ft below grade

Lower than 2 ft

0.35 0.14 0.03

0.15 0.09 0.03

Refer to ASHRAE Handbook 1989, Fundamentals, for details. The space heating load in the perimeter zone, including mainly transmission and in f ltration losses, is sometimes expressed in a linear density qh, ft, in Btu /h per linear foot of e xternal wall, or Btu/hft (W/m).

Adjacent Unheated Spaces Heat loss from the heated space to the adjacent unheated space Qun, Btu/h (W), is usually assumed to be balanced by the heat transfer from the unheated space to the outdoor air, and this can be calculated approximately by the following formula: SH__ ST__ LG__ DF

Q un 

n

m

 AiUi(Tr  Tun)  (j1  AjUj  V˙inf o cpa)(Tun  To) i1

(6.68)

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where Ai, Ui  area and overall heat-transfer coeff cient of partitions between heated space and unheated space, ft2 (m2) and Btu/h  ft2 °F (W/m2 °C) Aj, Uj  area and overall heat-transfer coeff cient of building structures exposed to outdoor air in unheated space, ft2 (m2) and Btu/h  ft2 °F (W/m2 °C)

__RH TX

The temperature of the unheated space Tun, °F (°C), can be calculated as

Tun 

(60V˙inf ocpa 

n

 AiUiTr  i1

n

m

 AjUj)To j1 m

 AiUi  60V˙inf ocpa  j1  AjUj i1

(6.69)

Latent Heat Loss and Heat Loss from Products In Eq. (6.64), Ql, Btu/h (W), represents the heat required to e vaporate the liquid w ater to raise the relative humidity of the space air or to maintain a specif c space relative humidity, i.e., Q l  m˙ w h fg,57  [60V˙oinf o(wr  wo)  m˙ p]h fg,57

(6.70)

where m˙ w  mass f ow of water evaporated, lb/h (kg/s) V˙oinf  volume f ow rate of outdoor ventilation air and inf ltrated air, cfm [m3 /(60 s)] m˙ p  mass f ow of water evaporated from minimum number of occupants that always stay in conditioned space when heating system is operated, lb/h (kg/s) In Eq. (6.70), hfg,57 indicates the latent heat of v aporization at a wet-b ulb temperature of 57 °F (13.9°C), that is, 72°F (22.2 °C) dry-b ulb temperature and a relati ve humidity of 40 percent. Its value can be taken as 1061 Btu/lb (2.47  106 J/kg). For factories, heat added to the products or materials that enter the heated space within the occupied period Qmat, Btu/h (W), should be considered part of the heating load and can be calculated as Q mat  m˙ matcpm(Tr  To)

(6.71)

where m˙ mat  mass f ow rate of cold products and cold material entering heated space, lb/h (kg/s) cpm  specif c heat of product or material, Btu/lb  °F (J/kg°C) Infiltration Inf ltration can be considered to be 0.15 to 0.4 air changes per hour (ach) at winter design conditions only when (1) the exterior window is not well sealed and (2) there is a high wind velocity. The more sides that have windows in a room, the greater will be the in f ltration. For hotels, motels, and high-rise domicile b uildings, an inf ltration rate of 0.038 cfm /ft2 (0.193 L /sm2) of gross area of exterior windows is often used for computations for the perimeter zone. As soon as the v olume f ow rate of inf ltrated air V˙inf , cfm (m 3 /min), is determined, the sensible heat loss from inf ltration Qinf, s, Btu/h (W), can be calculated as Q inf, s  V˙inf ocpa(Tr  To)

(6.72)

where o  density of outdoor air, lb/ft3 (kg/m3). Setback of Night Shutdown Operation During a nighttime or unoccupied period, when the space temperature is set back lo wer than the indoor temperature during the operating period, it is necessary to w arm up the conditioned space

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the next morning before the arri val of the occupants in of f ces or other b uildings. The warm-up or pickup load depends on the pickup temperature dif ference, the outdoor-indoor temperature dif ference, the construction of the building envelope or building shell, and the time required for warm-up. There are insuff cient data to determine the oversizing factor of the capacity of the heating plant for morning warm-up. According to tests by Trehan et al. (1989), one can make a rough estimate of the energy required to warm up or raise a space temperature by 12 °F (6.7°C) for a single- or tw o-story building with a roof-ceiling U value of 0.03 Btu /h ft2 °F (0.17 W/m2 °C) and a U value for external walls of 0.08 Btu/h ft2 °F (0.45 W/m2 °C) as follows: Outdoor-indoor temperature difference, °F 35 55 55

Warm-up period, h

Oversizing factor, percent

1 2 1

40 40 100

As discussed in Sec. 5.16, heating pickup load during warm-up period depends on the difference between space temperature and occupied set point, and the amount of time prior to scheduled occupancy, and is often determined by computer software.

6.11 LOAD CALCULATION SOFTWARE Introduction Today, most of the load calculations are performed by personal computers. Among the widely adopted load calculation and ener gy analysis softw are, only Building Load Analysis and System Thermodynamics (BLAST) developed by the University of Illinois adopts the heat balance method. All the others are based on the transfer function or weighting factors method. Load calculation softw are can also be di vided into tw o cate gories. The f irst includes those developed by go vernment or public institutions, such as the Department of Ener gy (DOE-2.0), National Bureau of Standards Load Program (NBSLP), and BLAST, which are “white box,” or transparent to the user , and called public domain softw are. The second cate gory consists of those programs developed by the pri vate sector, such as TRACE 600, developed by The Trane Company; HAP E20-II, developed by Carrier Corporation; and HCC (loads) and ESP (ener gy), developed by Automatic Procedures for Engineering Consultants Inc. (APEC). The pri vatesector-developed softw are programs were based on the published literature of go vernment, research, and public institutions such as ASHRAE. The most widely used, reliable, user-friendly, and continuously supported load and ener gy calculation programs in design are TRACE 600, HAP E20-II, and DOE-2.1E. BLAST is the most elaborate load calculation program developed in the United States and is usually considered a research tool. Most HV AC&R designers do care about the accuracy of the computational results of the softw are; however, the priority is the user friendly inputs and outputs.

Trace 600 — Structure and Basics

SH__ ST__ LG__ DF

In this section we include the basics and inputs of the most widely used softw are program in load calculations, TRACE 600 written for DOS, in the United States. Recently , Windows 95-based TRACE 700 Load Design has become a vailable. TRACE 600 Load and TRACE 700 Load Design are similar in structure, basics, and engineering capabilities. Most of the load and ener gy calculation softw are consists of four principal programs: loads, systems, plant, and economics (LSPE). TRACE 600 divides into f ve phases:

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●

●

●

●

6.43

Load. In the load phase, all external and internal heat gains are calculated. The heat gain prof le is then converted to a cooling load prof le. Design. In the design phase, based on the maximum block space sensible load, the volume f ow and the size of the air system are determined from the psychrometric analysis. In design phase it also calculates the coil’s load and selects the size of the coils. System. In the system simulation, the program predicts the load that is imposed on the equipment according to the space load prof le and the type of air system selected. Equipment. In the equipment (plant) phase, the energy consumption of the f ans, furnaces or boilers, and refrigeration systems is determined based on the hourly coil loads. Economic. During the economic analysis, utility cost is calculated from the ener gy use. When it is combined with installation cost and maintenance cost, a life c ycle cost is a vailable for comparison.

__RH TX

Only the load calculation program is co vered in this section. The a vailable documentation for TRACE 600 includes User’s Manual, Engineering Manual, Quick Reference, Cook Book, Software Bulletin — Getting Started, etc. The hardware requirements for TRACE 600 are as follows: Personal computer Hard disk

Ram requirement DOS Version 3.1 or higher

IBM AT 286 processor or higher Program: 7.5 Mbytes hard disk space (3 M bytes for load calculation) Job: 3 to 10 Mbytes depending on the size of job 640 kbytes

The TRACE 600 data sheet can be created by using the input editor . The input f le is organized in card format. In each input card, every input is shown by a given f eld. Sometimes, the value and unit of input are separated into tw o f elds. Entering the inputs can be done either by using f eld mode or by using full-screen mode. In f eld mode, the user can access help and selection to determine and select a speci f ed input. Field mode input is preferable. In full-screen mode, the user can view a full screen of ra w input. Each line sho ws the inputs of a card and is pre f xed by a tw ocharacter code. Each f eld in a card is separated by a forward slash. Input data are subdi vided into f ve groups: job, load, system, equipment, and economic. Man y input cards and f elds inside the input cards are optional, i.e., may be left blank.

Trace 600 Input — Load Methodology In TRACE 600, card 10 lists various methods of cooling and heating load calculations for the user’s selection. There are f ve cooling load calculating methods: ●

●

CEC-DOE2. This method adopts the transfer function method for both heat gain and space cooling load calculations. The space load calculations adopt the precalculated weighting f actors listed in DOE 2.1 Engineering Manual. This is a comparati vely exact cooling load and ener gy calculation method, especially for air systems operated at nighttime shutdo wn mode. CEC-DOE2 needs more computational calculations. Ho wever, it is often not a primary problem when the computations are performed by a powerful PC. CLTD-CLF. This method also adopts the transfer function method for both heat gain and space cooling load calculations. Cooling load temperature dif ference (CLTD) and cooling load f actor (CLF) tables are prepared based on the e xact transfer function coef f cients or weighting f actors

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from the TFM method. For many commercial buildings, when the air system is shut do wn during an unoccupied period, the external heat gains entering the space cannot be reasonably allocated over the cool-down period and the successive operating hours. ●

●

TETD-TA1 and TETD-TA2. The TETD-TA1 method adopts the transfer function method to calculate heat gain, and TETD-TA2 uses an approximate TETD to calculate heat gain which adopts a decrement factor and a time lag  to describe the amplitude and time-delay characteristics of the heat wave inside an exterior wall or a concrete roof slab. Both TETD-TA1 and TETD-TA2 use the time-averaging (TA) method to con vert the heat gain to space cooling load. The TA technique lacks scientif c support. TETD-PO. This method also adopts an approximate TETD to calculate the heat gain. It uses Post Off ce RMRG weighting f actors to con vert a heat gain of 100 percent radiati ve to space cooling load, and Post Of f ce RMRX weighting f actors to con vert a heat gain which is not 100 percent radiative, such as heat gains from people, lights, and equipment.

There are six heating load calculation methods. Fi ve of them — CEC-DOE2, CLTD-CLF, TETD-TA1, TETD-TA2, and TETD-PO — have already been described in the cooling load calculation methodology. The sixth heating load calculation method is called the U ATD method. In the UATD method, heat losses are calculated based on the U value  area  temperature difference, which is the temperature dif ference of the outdoor and indoor design temperatures. Heat losses are also considered as instantaneous heating load. For peak load calculation at winter design conditions, internal heat gains are not tak en into account in the U ATD method. Ho wever, for energy use calculation, internal heat gains need to be taken into account according to the load schedules. Otherwise the heating ener gy use becomes too conservative.

Trace 600 Input — Job

SH__ ST__ LG__ DF

There are 11 job cards: Cards 01 to 05 are used to describe the name of the project, its location, the client of the project, the program user, and comments. Card 08, climatic information, lists the name of weather f le, summer and winter clearness numbers, outdoor dry- and wet-b ulb temperatures, ground ref ectance, as well as the b uilding orientation. Card 09, load simulation periods, covers the f rst and last month of cooling design, summer period, and daylight savings time. It also co vers the peak cooling load hour. In card 10, load simulation parameters (optional), the selection of cooling and heating load methodology has been described in preceding paragraphs. In addition, the input data co ver airf ow input and output units, percentage of wall load included in the return air , and room circulation rate when the transfer function method is used in the conversion of heat gain to space cooling load. Outdoor air dry-bulb temperature and humidity ratio are also required to determine the state of the supply air during psychrometric analysis. The sensible and latent loads due to the v entilation outdoor air intak e are a component of cooling coil load. They do not af fect the supply air f ow and the size of the supply f an if the mixture of outdoor and recirculating air is extracted by the supply fan. Card 11, energy simulation parameters (optional), describes the f rst and last months of ener gy simulation, holiday and calendar type, and the conditioned f oor area. It also determines the input data of calculation level whether it is at room, zone, or air system level. Card 12, resource utilization f actors (optional), covers the input data of ener gy utilization factors which indicate the inef ficiency of producing and transfering ener gy, of electricity, gas, oil, steam, hot w ater, chilled w ater, and coal. Card 13, daylighting parameters (optional), describes the atmospheric moisture, atmospheric turbidity , or a measure of aerosols that affects daylighting, inside visible reflecti vity, and geometry method, such as glass percentage (GLAS-PCT).

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Trace 600 Input — External Loads There are seven cards for external load input. Card 19, load alternative description, will save separate f les for each of the Load, Design, System, and Equipment alternatives. Card 20, general room parameters, determines the room number , zone reference number, room description, f oor length and width, construction type (a 2- to 12-in. (50- to 300-mm), light- or heavyweight concrete construction), plenum height, R value of the acoustic ceiling, f oor-tof oor height, perimeter length, and duplicate f oor and duplicate room /zone multipliers to sa ve input data for a similar f oor or room/zone. Card 21, thermostat parameters, lists input data of room design cooling and heating dry-b ulb temperatures and relati ve humidities; the highest and lo west room temperatures allo wable to drift up during lo w occupanc y or nonoccupanc y; cooling and heating thermostat schedules; location of the thermostat; light, medium, or hea vy room construction and the corresponding 2 to 8 h of time averaging; and carpet covering. Card 22, roof parameters, describes the room number, alternate roof numbers, whether roof area is equal to f oor area, roof length and width, roof U value, roof construction type, roof direction, roof tilting angle, and roof absorptivity . Card 23, skylight parameters (optional), covers the room number, roof number, length and width of skylight, number of skylights or percentage of glass of roof area, skylight U value, skylight shading coeff cient, external and internal shading types, percentage of solar load to be pick ed up by return plenum air because a portion of sk ylight is e xposed to the plenum air , visible light transmissi vity, and the fraction of inside visible light being ref ected on the inside surface of the glazing. Card 24, wall parameters, lists the input data of room number; w all number; the length, height, U value, and construction type of the w all; wall direction; w all tilting angle; w all absorptivity; and ground ref ectance. Card 25, wall glass parameters, describes the room number, wall number, glass length and height, number of windows or percentage of glass (of gross wall area), glass U value, shading coeff cient, external and internal shading type, percentage of solar load to return plenum air , visible light transmissivity, and the inside visible light being ref ected on the inside surface of the glazing. Trace 600 Input — Schedules In TRACE 600, card 26 speci f es the load and operating schedules of the internal loads, fans, reheating, and daylighting (optional). The operating schedule of internal loads and equipment af fects the design load (maximum block load), especially the annual energy use. 1. For the design load calculation of internal loads include people, equipment: ●

●

●

lighting, and miscellaneous

TRACE 600 assumes that the internal loads are scheduled at 100 percent, 24 h /day during cooling design months operation and at 0 percent, 24 h /day during heating design months, a schedule of cooling only (CLGONLY). If the pickup, cool-down, or w arm-up load is greater than the design load, the user should consider oversizing the cooling capacity and heating capacity to handle the pickup loads. For the energy use calulation of internal loads, choose or create a schedule at which the internal load varies 24 h/day.

2. For the outdoor ventilation air and inf ltration loads: ●

TRACE 600 assumes that the design outdoor v entilation air and in f ltration are scheduled at 100 percent, 24 h/day for both cooling and heating design months.

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●

●

For the calculation of the ener gy use, only when the outdoor v entilation air system is separated from the main supply system is it then possible to choose and to create a schedule in which the airf ow of outdoor ventilation air varies according to people’s occupancy. During the calculation of ener gy use for a v ariable-air-volume (VAV) system, the airf ow o f the outdoor v entilation air may be v aried based on its control systems into the follo wing schedule: 100 percent, 24 h /day; varying according to supply v olume f ow of the air system; varying according to people’s occupancy. For the calculation of ener gy use due to in f ltration, choose and create a schedule in which the in f ltration v aries in versely with supply f an schedule, such as for hours f an schedule is 100 percent, inf ltration is zero; then when f an schedule is 0 percent, the in f ltration is 100 percent.

3. Operating schedules of f ans that af fect the design load and ener gy use calculations are as follows: At design load, TRACE 600 assumes that the main supply f an and auxiliary f an, if any, are scheduled at 100 percent, 24 h /day as the def ault value during cooling mode operation, and fan heat will be included in the cooling coil load. Similarly , at design load, main supply and auxiliary fans are scheduled at 0 percent, 24 h /day during heating mode operation, and no fan heat will be taken into account. In the energy use calculation, main supply and auxiliary fans in a constant-volume air system are scheduled at 100 percent, 24 h /day. Or choose or create a schedule at which f ans will be cycling (on and off) depending on the space cooling and heating load. In the energy use calculation, main supply and auxiliary f ans in a VAV system are scheduled depending on the space cooling and heating load. For a room e xhaust f an operated at cooling design load, it is scheduled at 100 percent, 24 h/day as the default value. For the energy use calculation, choose or create an appropriate operating schedule. ●

●

●

●

4. The reheat minimum percentage determines the ratio of reheat minimum airf ow for that room to the design supply airf ow. 5. Daylighting schedule describes the a vailability of daylighting each hour . Choose an a vailable schedule or create an appropriate daylighting schedule. Trace 600 Input — Internal Loads Four cards cover the input data of internal loads and airf ows.

SH__ ST__ LG__ DF

Card 27, people and lights, describes the room number, people’s density and corresponding unit, people’s sensible and latent loads, lighting heat gain and corresponding unit, type of lighting f xture, ballast f actor, and daylighting reference points 1 and 2, to sho w whether there is daylighting control or one or tw o control areas. According to the types of lighting f xtures and the return air f ow per unit f oor area, the percentage of lighting load to return air should be determined more precisely. Refer to Sec. 6.6 for details. Card 28, miscellaneous equipment (optional), covers the room number , miscellaneous equipment reference number , equipment description, energy consumption v alue and corresponding units, schedule and ener gy metering code, percentage of sensible load, percentage of miscellaneous equipment load to space and to air path, radiant fraction, and whether the plenum air , exhaust air, or return air path picks up the miscellaneous load. Card 29, room airf ows, covers the f ow rates of outdoor ventilation air during cooling and heating operation and their corresponding units, the f ow rates of in f ltration air for cooling and heating and their corresponding units, and the amount of minimum reheat and its units. Card 30, fan f ow rates, describes the room number, airf ow of the main and auxiliary cooling and heating supply fans and their corresponding units, and airf ow of room exhaust fan and its unit.

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Trace 600 Input — Partition and Shading Devices

__RH TX

Five cards cover the input data of partition and shading devices. Card 31, partition parameters (optional), describes the room number , adjacent room number , partition number, and the length, height, U value, and construction type of the partition w all. The adjacent space temperature during cooling and heating periods can follo w these pro f les: constant, sine f t, prorate the adjacent space temperature against the outdoor air temperature, vary as the outdoor air temperature, or there is no heat transfer from the adjacent space through the partition. Card 32, exposed f oor parameters (optional), lists the room number, adjacent room number, exposed f oor number, the perimeter length and loss coef f cient of the slab on grade; the area, U value, and construction type of the exposed f oor; and adjacent space temperature during cooling and heating periods whether it is constant, sine curve f t, or prorated curve f t. Card 33, external shading devices (optional), describes its constructional type, glass height and glass width, the height abo ve glass and the horizontal projection (projection out) of the o verhang, the projection left and right, the left and right projection out, and the shading due to adjacent buildings. TRACE 600 does not allo w for shading because of both adjacent b uilding and overhang. Card 34, internal shading de vices, covers the types, overall U value, overall shading coef ficient, location, operating schedule, and overall visible light transmittance of internal shading device; requires that belo w the minimum outdoor air dry-b ulb temperature and abo ve the maximum solar heat gain, the internal shading de vice is deplo yed; and sun control and glare control probability. Card 35, daylight sensor (optional), describes the daylight sensor reference number , percentage of space af fected, lighting set point in footcandles (lux es), type of control (continuous or stepped), minimum power or minimum light percentage, light control steps, manually operated control probability, height of the reference point and its distance from the glass, window sill height, ratio of the glass length seen by a sensor to the distance between the reference point and the glass, and skylight length and distance from the reference point.

Trace 600 — Minimum Input Requirements, Run, and Outputs For load calculation and energy analysis, the following are the minimum input requirements: Card 08: climatic information, f eld 2: weather f lename. Card 20: general room parameters, f eld 1: room number. Card 40: system type, f eld 1: system number, and f eld 2: system type. Card 41: zone assignment, f eld 1: system number, f eld 2: system serving these zones begins at, and f eld 3 : end at. After the user has completed all the input required for the job, he or she must edit for errors, check for possible data loss, and run and sa ve. Choose output from the TRACE 600 menu. Select the desired section of the output. Print the compressed report in a desired order. Example 6.2. The plan of a typical f oor in a multistory of f ce b uilding in Ne w York City is shown in Fig. 6.9. The tenants and the partition w alls of this f oor are unknown during the design stage. The core part contains the restrooms, stairwells, and mechanical and electrical service rooms, and it has mechanical e xhaust systems only . This building has the follo wing construction characteristics:

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Area of perimeter zone Area of interior zone Floor-to-f oor height Floor-to-ceiling height Window: Percentage of glass/exterior wall U value Interior shading, SC Exterior wall: Length Height U value Partition wall: Length Height U value

6300 ft2 (586 m2) 5600 ft2 (520 m2) 13 ft (4 m) 9 ft (2.74 m) 0.345 0.67 Btu/h ft2 °F (3.8 W/m 2 °C) 0.36 120 ft (36.6 m) 13 ft (2.74 m) 0.12 Btu/ h ft 2 °F (0.68 W/(m2 °C) 200 ft (61 m) 9 ft (2.74 m) 0.12 Btu/h ft2 °F (0.68 W/m2 °C)

This typical f oor has the follo wing internal loads: The occupant density is 7 persons per 1000 ft2 (93 m 2). For each person, there is 250 Btu /h (73 W) sensible heat gain, and 250 Btu /h (73 W) latent heat gain. The lighting load, including ballast allo wance, is 1 W/ft2 (10.7 W /m2). The miscellaneous equipment load due to personal computers and appliances is 1 W/ft2 (10.7 W/m2).

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FIGURE 6.9 Typical plan of a high-rise off ce building.

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The summer indoor design temperature is 75 °F (23.9°C). If the inf ltration load and the shading due to the adjacent building are ignored, calculate the maximum space cooling load during summer outdoor design conditions of this typical f oor by using TRACE 600 Load program with the CLTDCLD method. Solution. The maximum space cooling loads for this typical f oor, calculated by using TRACE 600 Load, with CLTD-CLF method on July 7, 5 P.M., are as follows:

TRACE 600, Btu/h Envelope loads Glass solar Glass conduction Wall conduction Partition wall Ceiling load Internal loads Lights People 31,446 Miscellaneous Grand total

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Manually calculated in Example 7.2 of f rst edition, Btu/h

36,846 11,095 9,927 1,296 4,226 63,390 36,147 37,994 105,587 168,977

198,770

Compare the TRACE 600 calculated space cooling load of 168,977 Btu /h (49,520 W) with the manually calculated space cooling load of the same typical f oor with the same building and operating characteristics (in Example 7.2 of Handbook of Air Conditioning and Refrig eration, 1st ed.) of 198,770 Btu / h (58,240 W). The TRACE 600 calculated space cooling load is only 85 percent of the manually calculated value. This is due to a lower solar CLF and therefore lower solar cooling loads in the TRACE 600 calculation.

REFERENCES Amistadi, H., Energy Analysis Software Review, Engineered Systems, no. 10, 1993, pp. 34 – 45. ASHRAE, ASHRAE/IESNA Standard 90.1-1999, Energy Standard for Buildings Except Low-Rise Residential Buildings, ASHRAE Inc., Atlanta, GA, 1989. ASHRAE, ASHRAE Handbook 1997, Fundamentals, Atlanta, GA, 1997. Ayres, J. M., and Stamper, E., Historical Development of Building Energy Calculations, ASHRAE Journal, no. 2, 1995, pp. 47 – 53. Carrier Air Conditioning Co., Handbook of Air Conditioning System Design, 1st ed., McGraw-Hill, New York, 1965. Deringer, J. J., An Overview of Standard 90.1: Building Envelope, ASHRAE Journal, February 1990, pp. 30 – 34. Harris, S. M., and McQuiston, F. C., A Study to Categorize Walls and Roofs on the Basis of Thermal Response, ASHRAE Transactions, 1988, Part II, pp. 688 – 715. Johnson, C. A., Besent, R.W., and Schoenau, G. J., An Economic Parametric Analysis of the Thermal Design of a Large Off ce Building under Different Climatic Zones and Different Billing Schedules, ASHRAE Transactions, 1989, Part I, pp. 355 – 369. Kerrisk, J. F., Schnurr, N. M., Moore, J. E., and Hunn, B. D., The Custom Weighting-Factor Method for Thermal Load Calculations in the DOE-2 Computer Program, ASHRAE Transactions, 1981, Part II, pp. 569 – 584. Kimura, K. I., and Stephenson, D. G., Theoretical Study of Cooling Loads Caused by Lights, ASHRAE Transactions, 1968, Part II, pp. 189 – 197.

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Komor, P., Space Cooling Demands from Off ce Plug Loads, ASHRAE Journal, no. 12, 1997, pp. 41 – 44. Kreith, F., and Black, W. Z., Basic Heat Transfer, Harper & Row, New York, 1980. Mackey, C. O., and Gay, N. R., Cooling Load from Sunlit Glass, ASHVE Transactions, 1952, pp. 321 – 330. Mackey, C. O., and Wright, L. T., Periodic Heat Flow — Homogeneous Walls or Roofs, ASHVE Transactions, 1944, pp. 293 – 312. McQuiston, F. C., and Spitler, J. D., Cooling and Heating Load Calculation Manual, 2d ed., ASHRAE Inc., Atlanta, GA, 1992. Mitalas, G. P., Transfer Function Method of Calculating Cooling Loads, Heat Extraction Rate and Space Temperature, ASHRAE Journal, no. 12, 1972, p. 52. Palmatier, E. P., Thermal Characteristics of Structures, ASHRAE Transactions, 1964, pp. 44 – 53. Persily, A. K., and Norford, L. K., Simultaneous Measurements of Inf ltration and Intake in an Off ce Building, ASHRAE Transactions, 1987, Part II, pp. 42 – 56. Romine, T. B., Cooling Load Calculation: Art or Science? ASHRAE Journal, no. 1, 1992, pp. 14 – 24. Rudoy, W., and Duran, F., Development of an Improved Cooling Load Calculation Method, ASHRAE Transactions, 1975, Part II, pp. 19 – 69. Rudoy, W., and Robins, L. M., Pulldown Load Calculations and Thermal Storage during Temperature Drift, ASHRAE Transactions, 1977, Part I, pp. 51 – 63. Snelling, H. J., Duration Study for Heating and Air Conditioning Design Temperature, ASHRAE Transactions, 1985, Part II B, p. 242. Sowell, E. F., Classif cation of 200,640 Parametric Zones for Cooling Load Calculations, ASHRAE Transactions, 1988, Part II, pp. 754 – 777. Sowell, E. F., and Chiles, D. C., Zone Descriptions and Response Characterization for CLF /CLTD Calculations, ASHRAE Transactions, 1985, Part II A, pp. 179 – 200. Sowell, E. F., and Hittle, D. C., Evolution of Building Energy Simulation Methodology, ASHRAE Transactions, 1995, Part I, pp. 851 – 855. Stephenson, D. G., and Mitalas, G. P., Cooling Load Calculations by Thermal Response Factor Method, ASHRAE Transactions, 1967, Part III, pp. 1.1 – 1.7. Sun, T. Y., Air Conditioning Load Calculation, Heating/Piping/Air Conditioning, January 1986, pp. 103 – 113. The Trane Company, TRACE 600: Engineering Manual, The Trane Company, LaCrosse, WI, 1992. The Trane Company, TRACE 600: User’s Manual, LaCrosse, WI, 1992. Trehan, A. K., Fortmann, R. C., Koontz, M. D., and Nagda, N. L., Effect of Furnace Size on Morning Pickup Time, ASHRAE Transactions, 1989, Part I, pp. 1125 – 1129. Wang, S. K., Air Conditioning, vol. 1, Hong Kong Polytechnic, Hong Kong, 1987. Williams, G. J., Fan Heat: Its Source and Signif cance, Heating/Piping/Air Conditioning, January 1989, pp. 101 – 112.

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7.1 FUNDAMENTALS 7.2 Types of Water System 7.2 Volume Flow and Temperature Difference 7.4 Water Velocity and Pressure Drop 7.5 7.2 WATER PIPING 7.7 Piping Material 7.7 Piping Dimensions 7.7 Pipe Joints 7.13 Working Pressure and Temperature 7.13 Expansion and Contraction 7.14 Piping Supports 7.15 Piping Insulation 7.15 7.3 VALVES, PIPE FITTINGS, AND ACCESSORIES 7.16 Types of Valves 7.16 Valve Connections and Ratings 7.17 Valve Materials 7.18 Piping Fittings and Water System Accessories 7.18 7.4 WATER SYSTEM PRESSURIZATION AND THE PRESENCE OF AIR 7.19 Water System Pressurization Control 7.19 Open Expansion Tank 7.20 Closed Expansion Tank 7.21 Size of Diaphragm Expansion Tank 7.21 Pump Location 7.23 Air in Water Systems 7.23 Penalties due to Presence of Air and Gas 7.24 Oxidation and Waterlogging 7.24 7.5 CORROSION AND DEPOSITS IN WATER SYSTEM 7.25 Corrosion 7.25 Water Impurities 7.25 Water Treatments 7.26 7.6 CLOSED WATER SYSTEM CHARACTERISTICS 7.27 System Characteristics 7.27 Changeover 7.28 7.7 CENTRIFUGAL PUMPS 7.30 Basic Terminology 7.30 Performance Curves 7.32 Net Positive Suction Head 7.33 Pump Selection 7.33 7.8 PUMP-PIPING SYSTEMS 7.34 System Curve 7.34 System Operating Point 7.34

Combination of Pump-Piping Systems 7.35 Modulation of Pump-Piping Systems 7.36 Pump Laws 7.37 Wire-to-Water Efficiency 7.37 7.9 OPERATING CHARACTERISTICS OF CHILLED WATER SYSTEM 7.38 Coil Load and Chilled Water Volume Flow 7.38 Chiller Plant 7.39 Variable Flow for Saving Energy 7.40 Water Systems in Commercial Buildings 7.40 7.10 PLANT-THROUGH-BUILDING LOOP 7.40 Bypass Throttling Flow 7.40 Distributed Pumping 7.41 Variable Flow 7.41 7.11 PLANT-BUILDING LOOP 7.43 System Description 7.43 Control Systems 7.43 System Characteristics 7.45 Sequence of Operations 7.46 Low T between Chilled Water Supply and Return Temperatures 7.49 Variable-Speed Pumps Connected in Parallel 7.49 Use of Balancing Valves 7.49 Common Pipe and Thermal Contamination 7.51 7.12 PLANT-DISTRIBUTED PUMPING 7.52 7.13 CAMPUS-TYPE WATER SYSTEMS 7.53 Plant-Distribution-Building Loop 7.54 Plant-Distributed Building Loop 7.56 Multiple Sources-Distributed Building Loop 7.57 Chilled and Hot Water Distribution Pipes 7.58 7.14 COMPUTER-AIDED PIPING DESIGN AND DRAFTING 7.58 General Information 7.58 Computer-Aided Drafting Capabilities 7.58 Computer-Aided Design Capabilities 7.59 REFERENCES 7.60

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Types of Water System Water systems that are part of an air conditioning system and that link the central plant, chiller/boiler, air-handling units (AHUs), and terminals may be classified into the foll wing categories according to their use: Chilled Water System. In a chilled water system, water is first cooled in the ater chiller — the evaporator of a reciprocating, screw, or centrifugal refrigeration system located in a centralized plant — to a temperature of 40 to 50°F (4.4 to 10.0°C). It is then pumped to the w ater cooling coils in AHUs and terminals in which air is cooled and dehumidified. After fl wing through the coils, the chilled w ater increases in temperature up to 60 to 65°F (15.6 to 18.3°C) and then returns to the chiller. Chilled water is widely used as a cooling medium in central hydronic air conditioning systems. When the operating temperature is belo w 38°F (3.3°C), inhibited glycols, such as ethylene glycol or propylene glycol, may be added to w ater to create an aqueous solution with a lo wer freezing point. Evaporative-Cooled Water System. In arid southwestern parts of the United States, cooled water is often produced by an evaporative cooler to cool the air.

evaporative-

Hot Water Systems. These systems use hot w ater at temperatures between 450 and 150°F (232 and 66°C) for space and process heating purposes. Hot w ater systems are co vered in greater detail in Chap. 8. Dual-Temperature Water System. In a dual-temperature w ater system, chilled water or hot w ater is supplied to the coils in AHUs and terminals and is returned to the w ater chiller or boiler mainly through the following two distribution systems: ●

●

Use supply and return main and branch pipes separately. Use the common supply and return mains, branch pipe, and coil for hot and chilled w ater supply and return.

The changeover from chilled w ater to hot w ater and vice v ersa in a b uilding or a system depends mainly on the space requirements and the temperature of outdoor air . Hot w ater is often produced by a boiler; sometimes it comes from a heat reco very system, which is discussed in later chapters. Condenser Water System. In a condenser w ater or cooling w ater system, the latent heat of condensation is removed from the refrigerant in the condenser by the condenser w ater. This condenser water either is from the cooling to wer or is surf ace water taken from a lak e, river, sea, or well. For an absorption refrigeration system, heat is also remo ved from the solution by cooling w ater in the absorber. The temperature of the condenser water depends mainly on the local climate. Water systems also can be classified according to their operating characteristics into the foll wing categories: Closed System. In a closed system, chilled or hot w ater fl wing through the coils, heaters, chillers, boilers, or other heat e xchangers forms a closed recirculating loop, as shown in Fig. 7.1 a. In a closed system, water is not exposed to the atmosphere during its fl wing process. The purpose of recirculation is to save water and energy.

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Open System. In an open system, the water is exposed to the atmosphere, as shown in Fig. 7.1 b. For example, chilled water comes directly into contact with the cooled and dehumidified air in th air washer, and condenser water is exposed to atmosphere air in the cooling to wer. Recirculation of water is used to save water and energy.

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FIGURE 7.1 Types of water systems. (a) Closed system; (b) open system; (c) once-through system.

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Open systems need more w ater treatments than closed systems because dust and impurities in the air may be transmitted to the w ater in open systems. A greater quantity of mak eup water is required in open systems to compensate for evaporation, drift carryover, or blow-down operation. Once-Through System. In a once-through system, water fl ws through the heat e xchanger only once and does not recirculate, as shown in Fig. 7.1 c. Lak e, river, well, or sea water used as condenser cooling water represents a once-through system. Although the water cannot recirculate to the condenser because of its rise in temperature after absorbing the heat of condensation, it can still be used for other purposes, such as flushing ater in a plumbing system after the necessary w ater treatments, to conserve water. In many locations, the law requires that well w ater be pumped back into the ground.

Volume Flow and Temperature Difference The heating and cooling capacity of w ater when it fl ws through a heat e xchanger Qw, Btu/h (W), can be calculated as Q w  V˙wwcpw (Twe  Twl)  500V˙gal (Twe  Twl)  500V˙gal T

(7.1)

where V˙w  volume fl w of water, ft3 /h (m3/s) V˙gal  volume fl w rate of water, gpm (L/s) w  density of water, lb/ft3 (kg/m3) cpw  specific heat of ater, Btu/lb  °F (J / k g  °C) T we ,Twl  temperature of water entering and leaving heat exchanger,°F (°C) Tw  temperature drop or rise of water after fl wing through heat exchanger,°F (°C) Here, the equivalent of pounds per hour is gpm  60 min /h  0.1337 ft3/gal  62.32 lb /ft3  500. Equation 7.1 also gi ves the relationship between Tw and Vgal during the heat-transfer process. The temperature of w ater lea ving the w ater chiller should be no lo wer than 37°F (2.8°C) to prevent freezing. If the chilled water temperature is lower than 37°F (2.8°C), brine, ethylene glycol, or prop ylene glycol should be used. Brine is discussed in a later chapter . F or a dual-temperature water system, the hot water temperature leaving the boiler often ranges from 100 to 150°F (37.8 to 65.6°C), and returns at a T between 20 and 40°F (11.1 and 22.2°C). For most dual-temperature w ater systems, the v alue of V˙gal and the pipe size are determined based on the cooling capacity requirement for the coils and w ater coolers. This is because chilled water has a smaller Tw than hot water does. Furthermore, the system cooling load is often higher than the system heating load. F or a chilled water system to transport each refrigeration ton of cooling capacity, a Tw of 8°F (4.4°C) requires a V˙gal of 3 gpm (0.19 L /s), whereas for a Tw of 24°F (13.3°C), V˙gal is only 1.0 gpm (0.063 L/s). The temperature of water entering the coil Twe, the temperature of water leaving the coil Twl, and the difference between them Tw  Twl  Twe are closely related to the performance of a chilled water system, air system, and refrigeration system: ●

●

●

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Temperature Twe directly affects the power consumption in the compressor. The temperature dif ferential Tw is closely related to the v olume fl w of chilled w ater V˙gal and thus the size of the water pipes and pumping power. Both Twe and Tw influence the temperature and humidity ratio of air le ving the coil.

If the chilled w ater temperature lea ving the w ater chiller and entering the coil is between 44 and 45°F (6.7 and 7.2°C), the off-coil temperature in the air system is usually around 55°F (12.8°C) for conventional comfort air conditioning systems. In lo w-temperature cold air distrib ution systems,

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chilled w ater lea ving the chiller may be as lo w as 34 °F (1.1 °C), and the of f-coil temperature is often between 42 and 47°F, typically 44°F (5.6 and 8.3°C, typically 6.7°C). The greater the v alue of Tw for chilled w ater, the lower the amount of w ater f owing through the coil. Current practice is usually to use a value of Tw between 10 and 18°F (5.6 and 10.0°C) for chilled water systems in buildings. Kelly and Chan (1999) noted a greater Tw results at lower total power consumption in w ater pumps, cooling to wer f ans, and chillers. Ho wever, a greater Tw means a lar ger coil and air -side pressure drop. F or chilled w ater systems in a campus-type central plant, a value of Tw between 16 and 24°F (8.9 and 13.3°C) is often used.

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Water Velocity and Pressure Drop The maximum w ater velocity in pipes is go verned mainly by pipe erosion, noise, and water hammer. Erosion of water pipes is the result of the impingement of rapidly mo ving water containing air bubbles and impurities on the inner surf ace of the pipes and f ttings. Solden and Sie gel (1964) increased their feedw ater velocity gradually from 8 ft /s (2.4 m /s) to an a verage of 35.6 ft /s (10.8 m/s). After 3 years, they found no evidence of erosion in the pipe or a connected check v alve. Erosion occurs only if solid matter is contained in w ater f owing at high v elocity. Velocity-dependent noise in pipes results from f ow turbulence, cavitation, release of entrained air , and water hammer that results from the transient pressure impact on a sudden closed v alve. Ball and Webster (1976) performed a series of tests on 83 -in. copper tubes with elbo ws. At a w ater velocity of 16.4 ft /s (5.0 m/s), the noise le vel was less than 53 dB A. Tests also sho wed that cold w ater at a speed up to 21 ft/s (6.4 m /s) did not cause cavitation. In copper and steel pipes, water hammer at a w ater velocity of 15 ft /s (4.6 m /s) exerted a pressure on 2-in.- (50-mm-) diameter pipes that w as less than 50 percent of their design pressure. Given the abo ve results, excluding the ener gy cost for the pump po wer, the maximum w ater velocity in certain short sections of a water system may be raised to an upper limit of 11 ft /s (3.35 m /s) for a special purpose, such as enhancing the heat-transfer coeff cients. Normally, water f ow in coils and heat e xchangers becomes laminar and seriously impairs the heat-transfer characteristics only when its v elocity drops to a v alue less than 2 ft /s (0.61 m /s) and its corresponding Re ynolds number is reduced to about 10,000 (within the transition re gion). In evaporators, Redden (1996) found that at lo w tube w ater velocity at 1.15 ft /s (0.35 m /s) and lo w heat f ux, instability of heat transfer occurred and caused the chilled w ater leaving temperature to f uctuate by 4°F (2.2°C). In condensers, condensing operation is not af fected even if the condenser water velocity in the tube w as about 1 ft /s (0.31 m /s). Water velocity should also be maintained at not less than 2 ft/s (0.61 m/s) in order to transport the entrained air to air vents. When pipes are being sized, the optimum pressure drop Hf , commonly e xpressed in feet (meters) of head loss of w ater per 100 ft of pipe length ( p in pascals of pressure drop per meter length), is a compromise between ener gy costs and in vestments. At the same time, the agecorrosion of pipes should be considered. Generally, the pressure drop for w ater pipes inside b uildings Hf is in a range of 1 ft /100 ft to 4 ft / 100 ft (100 to 400 Pa /m), with a mean of 2.5 ft /100 ft (250 Pa/m) used most often. Because of a lo wer increase in installation cost for smaller -diameter pipes, it may be best to use a pressure drop lower than 2.5 ft/100 ft (250 Pa/m) when the pipe diameter is 2 in. or less. Age corrosion results in an increase in the friction f actor and a decrease in the ef fective diameter. The factors that contrib ute to age corrosion are sliming, caking of calcareous salts, and corrosion. Many scientists recommended an increase in friction loss of 15 to 20 percent, resulting in a design pressure drop of 2 ft /100 ft (200 Pa/m), for closed water systems; and a 75 to 90 percent increase in friction loss, or a design pressure drop of 1.35 ft/100 ft (135 Pa/m), for open water systems. Figures 7.2, 7.3, and 7.4 sho w the pressure drop charts for steel, copper, and plastic pipes, respectively, for closed w ater systems. Each chart sho ws the v olume f ow (gpm), pressure drop V˙gal Hf (ft/100 ft), water velocity vw (ft/s), and water pipe diameter D (in.). Given any two of these parameters, the other tw o can be determined. F or instance, for a steel w ater pipe that has a w ater volume f ow of 1000 gpm, if the pressure drop is 2 ft /100 ft, the diameter is 8 in. and the corresponding velocity is about 8 ft/s.

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FIGURE 7.2 Friction chart for w ater in steel pipes (Schedule 40). ( Source: ASHRAE Handbook 1989 Fundamentals. Reprinted with permission.)

It is a common practice to limit the w ater velocity to no more than 4 ft /s (1.2 m /s) for w ater pipes 2 in. (50 mm) or less in diameter in order to pre vent an e xcessive H f . The pressure drop should not exceed 4 ft/100 ft (400 Pa/m) for water pipes of greater than 2-in. (50-mm) diameter. An open water system or a closed w ater system that is connected with an open e xpansion tank, all the pressure dif ferences between tw o points or le vels, and pressure drops across a piece of equipment or a device are expressed in feet of water column or psi (head in meters of water column or pressure loss in kP a). The total or static pressure of w ater at a certain point in a w ater system is actually measured and expressed by that part of pressure which is greater or smaller than the atmospheric pressure, often called gauge pressure, in feet of water column gauge or psig (meters gauge or kPa  g). The relationships between the steady f ow energy equation and the f uid pressure, and between the pressure loss and f uid head, are discussed in Sec.17.1.

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FIGURE 7.3 Friction chart for w ater in copper tubing (types K, Reprinted with permission.)

L, and M). ( Source: ASHRAE Handbook 1989 Fundamentals.

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FIGURE 7.4 Friction chart for w ater in plastic pipes (Schedule 80). ( Source: ASHRAE Handbook 1989 Fundamentals. Reprinted with permission.)

7.2 WATER PIPING Piping Material For water systems, the piping materials most widely used are steel, both black (plain) and galv anized (zinc-coated), in the form of either welded-seam steel pipe or seamless steel pipe; ductile iron and cast iron; hard copper; and polyvin yl chloride (PVC). The piping materials for various services are shown below: Chilled water Hot water Cooling water and drains

Black and galvanized steel Black steel, hard copper Black steel, galvanized ductile iron, PVC

Copper, galvanized steel, galvanized ductile iron, and PVC pipes ha ve better corrosion resistance than black steel pipes. Technical requirements, as well as local customs, determine the selection of piping materials. Piping Dimensions The steel pipe w all thicknesses currently used were standardized in 1930. The thickness ranges from Schedule 10, light wall, to Schedule 160, very heavy wall. Schedule 40 is the standard for a pipe with a diameter up to 10 in. (250 mm). F or instance, a 2-in. (50-mm) standard pipe has an outside diameter of 2.375 in. (60.33 mm) and an inside diameter of 2.067 in. (52.50 mm). The nominal pipe size is only an approximate indication of pipe size, especially for pipes of small diameter . Table 7.1 lists the dimensions of commonly used steel pipes. The outside diameter of extruded copper is standardized so that the outside diameter of the copper tubing is 1/8 in. (3.2 mm) larger than the nominal size used for soldered or brazed sock et joints. As in the case with steel pipes, the result is that the inside diameters of copper tubes seldom equal the nominal sizes. Types K, L, M, and DWV designate the wall thickness of copper tubes: type K is the heaviest, and DWV is the lightest. Type L is generally used as the standard for pressure copper tubing. Type DWV is used for drainage at atmospheric pressure.

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Nominal size and pipe OD D, in.

Surface area

Cross-sectional

Weight of

Inside diameter d, in.

Outside, ft2/ft

Inside, ft2/ft

Metal area, in.2

Flow area, in.2

Pipe, lb/ft

Water, lb/ft

1 4

40 ST 80 XS

0.088 0.119

0.364 0.302

0.141 0.141

0.095 0.079

0.125 0.157

0.104 0.072

0.424 0.535

0.045 0.031

3 8

40 ST 80 XS

0.091 0.126

0.493 0.423

0.177 0.177

0.129 0.111

0.167 0.217

0.191 0.141

0.567 0.738

1 2

40 ST 80 XS

0.109 0.147

0.622 0.546

0.220 0.220

0.163 0.143

0.250 0.320

0.304 0.234

D  1.050

3 4

40 ST 80 XS

0.113 0.154

0.824 0.742

0.275 0.275

0.216 0.194

0.333 0.433

1 D  1.315

40 ST 80 XS

0.133 0.179

1.049 0.957

0.344 0.344

0.275 0.251

1 41 D  1.660

40 ST 80 XS

0.140 0.191

1.380 1.278

0.435 0.435

1 12 D  1.900

40 ST 80 XS

0.145 0.200

1.610 1.500

2 D  2.375

40 ST 80 XS

0.154 0.218

212 D  2.875

40 ST 80 XS

3 D  3.500

psig

CW CW

Thrd Thrd

188 871

0.083 0.061

CW CW

Thrd Thrd

203 820

0.850 1.087

0.131 0.101

CW CW

Thrd Thrd

214 753

0.533 0.432

1.13 1.47

0.231 0.187

CW CW

Thrd Thrd

217 681

0.494 0.639

0.864 0.719

1.68 2.17

0.374 0.311

CW CW

Thrd Thrd

226 642

0.361 0.335

0.669 0.881

1.50 1.28

2.27 2.99

0.647 0.555

CW CW

Thrd Thrd

229 594

0.497 0.497

0.421 0.393

0.799 1.068

2.04 1.77

2.72 3.63

0.881 0.765

CW CW

Thrd Thrd

231 576

2.067 1.939

0.622 0.622

0.541 0.508

1.07 1.48

3.36 2.95

3.65 5.02

1.45 1.28

CW CW

Thrd Thrd

230 551

0.203 0.276

2.469 2.323

0.753 0.753

0.646 0.608

1.70 2.25

4.79 4.24

5.79 7.66

2.07 1.83

CW CW

Weld Weld

533 835

40 ST 80 XS

0.216 0.300

3.068 2.900

0.916 0.916

0.803 0.759

2.23 3.02

7.39 6.60

7.57 10.25

3.20 2.86

CW CW

Weld Weld

482 767

4 D  4.500

40 ST 80 XS

0.237 0.337

4.026 3.826

1.178 1.178

1.054 1.002

3.17 4.41

12.73 11.50

10.78 14.97

5.51 4.98

CW CW

Weld Weld

430 695

6 D  6.625

40 ST 80 XS

0.280 0.432

6.065 5.761

1.734 1.734

1.588 1.508

5.58 8.40

28.89 26.07

18.96 28.55

12.50 11.28

ERW ERW

Weld Weld

696 1209

D  0.540 D  0.675 D  0.840

Mfr. process

Chapter 7

Joint type

bb

Wall thickness t, in.

CONF PASS

Schedule number or weight*

Working pressure† ASTM A538B to 400°F

Wang (MCGHP)

TABLE 7.1 Dimensions of Commonly Used Steel Pipes

7/19/00 pg 7.8

7.8

TABLE 7.1

Nominal size and pipe OD D, in.

Dimensions of Commonly Used Steel Pipes (Continued) Surface area

Cross-sectional

Inside diameter d, in.

Outside, ft2/ft

Inside, ft2/ft

Metal area, in.2

8 D  8.625

30 40 ST 80 XS 30

0.277 0.322 0.500 0.307

8.071 7.981 7.625 10.136

2.258 2.258 2.258 2.814

2.113 2.089 1.996 2.654

7.26 8.40 12.76 10.07

51.16 50.03 45.66 80.69

24.68 28.53 43.35 34.21

10 D  10.75

40 ST XS 80 30

0.365 0.500 0.593 0.330

10.020 9.750 9.564 12.090

2.814 2.814 2.814 3.338

2.623 2.552 2.504 3.165

11.91 78.85 16.10 74.66 18.92 71.84 12.88 114.8

ST

0.375 0.406 0.500 0.687 0.375

12.000 11.938 11.750 11.376 13.250

3.338 3.338 3.338 3.338 3.665

3.141 3.125 3.076 2.978 3.469

14.58 15.74 19.24 26.03 16.05

40

XS 80 30 ST

Pipe, lb/ft

Water, lb/ft

Mfr. process

Joint type

psig

22.14 21.65 19.76 34.92

ERW ERW ERW ERW

Weld Weld Weld Weld

526 643 1106 485

40.45 54.69 64.28 43.74

34.12 32.31 31.09 49.68

ERW ERW ERW ERW

Weld Weld Weld Weld

606 887 1081 449

113.1 111.9 108.4 101.6 137.9

49.52 53.48 65.37 88.44 54.53

48.94 48.44 46.92 43.98 59.67

ERW ERW ERW ERW ERW

Weld Weld Weld Weld Weld

528 583 748 1076 481

XS

0.437 0.500 0.750

13.126 13.000 12.500

3.665 3.665 3.665

3.436 3.403 3.272

18.62 135.3 21.21 132.7 31.22 122.7

63.25 72.04 106.05

58.56 57.44 53.11

ERW ERW ERW

Weld Weld Weld

580 681 1081

16 D  16.00

30 ST 40 XS ST

0.375 0.500 0.375

15.250 15.000 17.250

4.189 4.189 4.712

3.992 3.927 4.516

18.41 182.6 24.35 176.7 20.76 233.7

62.53 82.71 70.54

79.04 76.47 101.13

ERW ERW ERW

Weld Weld Weld

421 596 374

18 D  18.00

30

XS

0.437 0.500 0.562

17.126 17.000 16.876

4.712 4.712 4.712

4.483 4.450 4.418

24.11 230.3 27.49 227.0 30.79 223.7

81.91 93.38 104.59

99.68 98.22 96.80

ERW ERW ERW

Weld Weld Weld

451 530 607

20 D  20.00

ST 30 XS 40

0.375 0.500 0.593

19.250 19.000 18.814

5.236 5.236 5.236

5.039 4.974 4.925

23.12 291.0 30.63 283.5 36.15 278.0

78.54 104.05 122.82

125.94 122.69 120.30

ERW ERW ERW

Weld Weld Weld

337 477 581

80

40

CONF PASS

40

Chapter 7

14 D  14.00

Wang (MCGHP)

Wall thickness t, in.

Working pressure† ASTM A538B to 400°F

39445

Schedule number or weight*

12 D  12.75

Flow area, in.2

Weight of

bb 7/19/00

7.9

*Numbers are schedule number per ASTM B36.10; ST  standard weight; XS  extra strong. † Working pressures have been calculated per ASME/ANSI B31.9 using furnace butt weld (continuous weld, CW) pipe through 4 in. and electric resistance weld (ERW) thereafter. The allowance A has been taken as (a) 12.5 percent of t for mill tolerance on pipe wall thickness, plus (b) an arbitrary corrosion allowance of 0.025 in. for pipe sizes through NPS 2 and 0.065 in. from NPS 2 12 through 20 plus (c) a thread cutting allowance for sizes through NPS 2. Because the pipe wall thickness of threaded standard weight pipe is so small after deducting the allowance A, the mechanical strength of the pipe is impaired. It is good practice to limit standard-weight threaded pipe pressures to 90 psig for steam and 125 psig for water. Source: ASHRAE Handbook 1988, Equipment. Reprinted with permission.

pg 7.9

TX

DF

39445

Type

Surface area

Cross-sectional

Wall thickness t, in.

Outside diameter D, in.

Inside diameter d, in.

Outside, ft2/ft

Inside, ft2/ft

Metal area, in.2

Flow area, in.2

Weight of Tube, lb/ft

Water, lb/ft

Working pressure* ASTMB 888 to 250°F Annealed, psig

Drawn, psig

0.098 0.098

0.080 0.082

0.037 0.033

0.073 0.078

0.145 0.126

0.032 0.034

851 730

1596 1368

3 8

K L M

0.049 0.035 0.025

0.500 0.500 0.500

0.402 0.430 0.450

0.131 0.131 0.131

0.105 0.113 0.008

0.069 0.051 0.037

0.127 0.145 0.159

0.269 0.198 0.145

0.055 0.063 0.069

894 638 456

1676 1197 855

1 2

K L M

0.049 0.040 0.028

0.625 0.625 0.625

0.527 0.545 0.569

0.164 0.164 0.164

0.138 0.143 0.149

0.089 0.074 0.053

0.218 0.233 0.254

0.344 0.285 0.203

0.094 0.101 0.110

715 584 409

1341 1094 766

5 8

K L

0.049 0.042

0.750 0.750

0.652 0.666

0.196 0.196

0.171 0.174

0.108 0.093

0.334 0.348

0.418 0.362

0.144 0.151

596 511

1117 958

3 4

K L M

0.065 0.045 0.032

0.875 0.875 0.875

0.745 0.785 0.811

0.229 0.229 0.229

0.195 0.206 0.212

0.165 0.117 0.085

0.436 0.484 0.517

0.641 0.455 0.328

0.189 0.209 0.224

677 469 334

1270 879 625

1

K L M

0.065 0.050 0.035

1.125 1.125 1.125

0.995 1.025 1.055

0.295 0.295 0.295

0.260 0.268 0.276

0.216 0.169 0.120

0.778 0.825 0.874

0.839 0.654 0.464

0.336 0.357 0.378

527 405 284

988 760 532

1 14

K L M DWV

0.065 0.055 0.042 0.040

1.375 1.375 1.375 1.375

1.245 1.265 1.291 1.295

0.360 0.360 0.360 0.360

0.326 0.331 0.338 0.339

0.268 0.228 0.176 0.168

1.217 1.257 1.309 1.317

1.037 0.884 0.682 0.650

0.527 0.544 0.566 0.570

431 365 279 265

808 684 522 497

1 12

K L M DWV

0.072 0.060 0.049 0.042

1.625 1.625 1.625 1.625

1.481 1.505 1.527 1.541

0.425 0.425 0.425 0.425

0.388 0.394 0.400 0.403

0.351 0.295 0.243 0.209

1.723 1.779 1.831 1.865

1.361 1.143 0.940 0.809

0.745 0.770 0.792 0.807

404 337 275 236

758 631 516 442

2

K L M DWV

0.083 0.070 0.058 0.042

2.125 2.125 2.125 2.125

1.959 1.985 2.009 2.041

0.556 0.556 0.556 0.556

0.513 0.520 0.526 0.534

0.532 0.452 0.377 0.275

3.014 3.095 3.170 3.272

2.063 1.751 1.459 1.065

1.304 1.339 1.372 1.416

356 300 249 180

668 573 467 338

pg 7.10

0.305 0.315

7/19/00

0.375 0.375

bb

0.035 0.030

CONF PASS

K L

Chapter 7

1 4

Wang (MCGHP)

Nominal diameter, in.

RH__

TX

SH__ ST__ LG__

DF

TABLE 7.2 Dimensions of Copper Tubes

7.10

Dimensions of Copper Tubes (Continued)

TABLE 7.2

Nominal diameter, in.

Type

Surface area

Wall thickness t, in.

Outside diameter D, in.

Inside diameter d, in.

Outside, ft2/ft

Inside, ft2/ft

Cross-sectional Metal area, in.2

Flow area, in.2

Weight of Tube, lb/ft

Water, lb/ft

Working pressure* ASTMB 888 to 250°F Annealed, psig

Drawn, psig

0.637 0.645 0.653

0.755 0.640 0.523

4.657 4.772 4.889

2.926 2.479 2.026

2.015 2.065 2.116

330 278 226

619 521 423

3

K L M DWV

0.109 0.090 0.072 0.045

3.125 3.125 3.125 3.125

2.907 2.945 2.981 3.035

0.818 0.818 0.818 0.818

0.761 0.771 0.780 0.795

1.033 0.858 0.691 0.435

6.637 6.812 6.979 7.234

4.002 3.325 2.676 1.687

2.872 2.947 3.020 3.130

318 263 210 131

596 492 394 246

3 12

K L M

0.120 0.100 0.083

3.625 3.625 3.625

3.385 3.425 3.459

0.949 0.949 0.949

0.886 0.897 0.906

1.321 1.107 0.924

8.999 9.213 9.397

5.120 4.291 3.579

3.894 3.987 4.066

302 252 209

566 472 392

4

K L M DWV

0.134 0.110 0.095 0.058

4.125 4.125 4.125 4.125

3.857 3.905 3.935 4.009

1.080 1.080 1.080 1.080

1.010 1.022 1.030 1.050

1.680 1.387 1.203 0.741

11.684 11.977 12.161 12.623

6.510 5.377 4.661 2.872

5.056 5.182 5.262 5.462

296 243 210 128

555 456 394 240

5

K L M DWV

0.160 0.125 0.109 0.072

5.125 5.125 5.125 5.125

4.805 4.875 4.907 4.981

1.342 1.342 1.342 1.342

1.258 1.276 1.285 1.304

2.496 1.963 1.718 1.143

18.133 18.665 18.911 19.486

9.671 7.609 6.656 4.429

7.846 8.077 8.183 8.432

285 222 194 128

534 417 364 240

6

K L M DWV

0.192 0.140 0.122 0.083

6.125 6.125 6.125 6.125

5.741 5.845 5.881 5.959

1.603 1.603 1.603 1.603

1.503 1.530 1.540 1.560

3.579 2.632 2.301 1.575

25.886 26.832 27.164 27.889

13.867 10.200 8.916 6.105

11.201 11.610 11.754 12.068

286 208 182 124

536 391 341 232

8

K L M DWV

0.271 0.200 0.170 0.109

8.125 8.125 8.125 8.125

7.583 7.725 7.785 7.907

2.127 2.127 2.127 2.127

1.985 2.022 2.038 2.070

6.687 4.979 4.249 2.745

45.162 46.869 47.600 49.104

25.911 19.295 16.463 10.637

19.542 20.280 20.597 21.247

304 224 191 122

570 421 358 229

10

K L M

0.338 0.250 0.212

10.125 10.125 10.125

9.449 9.625 9.701

2.651 2.651 2.651

2.474 2.520 2.540

10.392 7.756 6.602

70.123 72.760 73.913

40.271 30.054 25.584

30.342 31.483 31.982

304 225 191

571 422 358

12

K L M

0.405 0.280 0.254

12.125 12.125 12.125

11.315 11.565 11.617

3.174 3.174 3.174

2.962 3.028 3.041

14.912 10.419 9.473

100.554 105.046 105.993

57.784 40.375 36.706

43.510 45.454 45.863

305 211 191

571 395 358

7/19/00

0.687 0.687 0.687

bb

2.435 2.465 2.495

CONF PASS

2.625 2.625 2.625

Chapter 7

0.095 0.080 0.065

Wang (MCGHP)

K L M

39445

2 12

pg 7.11

7.11

*When using soldered or brazed f ttings, the joint determines the limiting pressure. Working pressures calculated using ASME B31.9 allowable stresses. A 5 percent mill tolerance has been used on the wall thickness. Higher tube ratings can be calculated using allowable stress for lower temperatures. If soldered or brazed f ttings are used on hard-drawn tubing, use the annealed ratings. Full-tube allowable pressures can be used with suitably rated f are or compression-type f ttings. Source: ASHRAE Handbook 1988, Equipment. Reprinted with permission.

__RH

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__SH __ST __LG

DF

39445

RH__

7.12

Wang (MCGHP)

Chapter 7

CONF PASS

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7/19/00

pg 7.12

CHAPTER SEVEN

TX

Copper tubes are also cate gorized as hard and soft copper . Soft pipes should be used in applications for which the pipe will be bent in the f eld. Table 7.2 lists the dimensions of copper tubes. Thermoplastic plastic pipes are the most widely used plastic pipes in air conditioning. They are manufactured with dimensions that match steel pipe dimensions. The advantages of plastic pipes include resistance to corrosion, scaling, and the growth of algae and fungi. Plastic pipes ha ve smooth surfaces and negligible age allowance. Age allowance is the allowance for corrosion and scaling for plastic pipes during their service life. Most plastic pipes are lo w in cost, especially compared with corrosion-resistant metal tubes. The disadvantages of plastic pipes include the f act that their pressure ratings decrease rapidly when the w ater temperature rises abo ve 100 °F (37.8 °C). PVC pipes are weak er than metal pipes and must usually be thick er than steel pipes if the same w orking pressure is to be maintained. Plastic pipes may e xperience e xpansion and contraction during temperature changes that is 4 times greater than that of steel. Precautions must be tak en to protect plastic pipes from e xternal damage and to account for its beha vior during f re. Some local codes do not permit the use of some or allplastic pipes. It is necessary to check with local authorities.

Pipe Joints Steel pipes of small diameter (2 in. or 50 mm less) threaded through cast-iron f ttings are the most widely used type of pipe joint. F or steel pipes of diameter 2 in. (50 mm) and more, welded joints,

TABLE 7.3 Maximum Allowable Pressures at Corresponding Temperatures System

Fitting Application Recirculating water 2 in. and smaller

2.5 – 12 in.

Refrigerant

SH__ ST__ LG__ DF

Pipe material

Weight

Joint type

Class

Material

Temperature, °F

Maximum allowable pressure at temperature, psig

Steel (CW) Copper, hard PVC CPVC PB

Standard Type L Sch. 80 Sch. 80 SDR-11

Thread 95-5 solder Solvent Solvent Heat fusion Insert crimp

125 — Sch. 80 Sch. 80 — —

Cast iron Wrought copper PVC CPVC PB Metal

250 250 75 150 160 160

125 150 350 150 115 115

A53 B ERW steel

Standard

PB

SDR-11

Weld Flange Flange Flange Groove Heat fusion

Standard 150 125 250 —

Wrought steel Wrought steel Cast iron Cast iron MI or ductile iron PB

250 250 250 250 230 160

400 250 175 400 300 115

Copper, hard A53 B SML steel

Type L or K

Braze

—

Wrought copper

—

—

Standard

Weld

Wrought steel

—

—

Note: Maximum allowable working pressures have been derated in this table. Higher system pressures can be used for lower temperatures and smaller pipe sizes. Pipe, f ttings, joints, and valves must all be considered. Note: A53 ASTM Standard A53 PVC Polyvinyl chloride CPVC Chlorinated polyvinyl chloride PB Polybutylene Source: Abridged with permission from ASHRAE Handbook 1988, Equipment.

39445

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Chapter 7

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7/19/00

pg 7.13

WATER SYSTEMS

7.13

bolted f anges, and grooved ductile iron joined f ttings are often used. Galv anized steel pipes are threaded together by galvanized cast iron or ductile iron f ttings. Copper pipes are usually joined by soldering and brazing sock et end f ttings. Plastic pipes are often joined by solvent welding, fusion welding, screw joints, or bolted f anges. Vibrations from pumps, chillers, or cooling to wers can be isolated or dampened by means of f exible pipe couplings. Arch connectors are usually constructed of n ylon, dacron, or polyester and neoprene. They can accommodate de f ections or dampen vibrations in all directions. Restraining rods and plates are required to pre vent e xcessive stretching. A f exible metal hose connector includes a corrugated inner core with a braided co ver. It is a vailable with f anged or groo ved end joints.

__RH TX

Working Pressure and Temperature In a water system, the maximum allowable working pressure and temperature are not limited to the pipes only; joints or the pipe f ttings, especially valves, may often be the weak links. Table 7.3 lists types of pipes, joint, and f ttings and their maximum allo wable w orking pressures for speci f ed temperatures.

Expansion and Contraction During temperature changes, all pipes e xpand and contract. The design of w ater pipes must tak e into consideration this expansion and contraction. Both the temperature change during the operating period and the possible temperature change between the operating and shutdo wn periods should

FIGURE 7.5 Expansion loops. (a) U bends; ( b) L bends; ( c ) Z bends.

__SH __ST __LG DF

39445

RH__ TX

7.14

Wang (MCGHP)

Chapter 7

CONF PASS

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7/19/00

pg 7.14

CHAPTER SEVEN

also be considered. For chilled and condenser water, which has a possible temperature change of 40 to 100°F (4.4 to 37.8 °C), expansion and contraction cause considerable mo vement in a long run of piping. Unexpected expansion and contraction cause e xcess stress and possible f ailure of the pipe, pipe support, pipe joints, and f ttings. Expansion and contraction of hot and chilled w ater pipes can be better accommodated by using loops and bends. The commonly used bends are U bends, Z bends, and L bends, as sho wn in Fig. 7.5. Anchors are the points where the pipe is f xed so that it will e xpand or contract between them. Reaction forces at these anchors should be considered when the support is being designed. ASHRAE Handbook 1992 , HVAC Systems and Equipment , gives the required calculations and data for determining these stresses. Guides are used so that the pipes expand laterally. Empirical formulas are often used instead of detailed stress analyses to determine the dimension of the offset leg Lo [ft (m)]. Waller (1990) recommended the following formulas: 0.46 T U bends: Lo  0.041D 0.48L ac

Z bends: Lo  (0.13DLac T) 0.5 L bends: Lo  (0.314DLac T)

(7.2)

0.5

where D  diameter of pipe, in. (mm) Lac  distance between anchors, hundreds of ft (m) T  temperature difference, °F (°C) If there is no room to accommodate U, Z, or L bends (such as in high-rise b uildings or tunnels), mechanical e xpansion joints are used to compensate for mo vement during e xpansion. P acked expansion joints allow the pipe to slide to accommodate movement during expansion. Various types of packing are used to seal the sliding surfaces in order to prevent leakage. Another type of mechanical joint uses bello ws or f exible metal hose to accommodate mo vement. These types of joints should be carefully installed to avoid distortion. TABLE 7.4 Recommended Pipe Hanger Spacing, ft Nominal pipe diameter, in. 1 2 3 4

SH__ ST__ LG__ DF

1 1 12 2 2 12 3 4 6 8 10 12 14 16 18 20

Standard-weight steel pipe (water) 7 7 7 9 10 11 12 14 17 19 20 23 25 27 28 30

Copper tube (water)

Rod size, in.

5 5 6 8 8 9 10 12 14 16 18 19

1 4 1 4 1 4 3 8 3 8 3 8 3 8 1 2 1 2 5 8 3 4 7 8

1 1 1 14 1 14

Note: Spacing does not apply where concentrated loads are placed between supports such as f anges, valves, and specialties. Source: ASHRAE Handbook 1988, Equipment. Reprinted with permission.

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7/19/00

pg 7.15

WATER SYSTEMS

7.15

__RH TX

Piping Supports Types of piping support include hangers, which hang the pipe from abo ve; supports, which usually use brackets to support the pipe from belo w; anchors to control the mo vement of the piping; and guides to guide the axial mo vement of the piping. Table 7.4 lists the recommended spacing of pipe hangers. Piping support members should be constructed based on the stress at their point of connection to the pipe as well as on the characteristics of the structural system. Pipe supports must ha ve suff cient strength to support the pipe, including the w ater inside. Except for the anchors, they should also allow for expansion movement. Pipes should be supported around the connections to the equipment so that the pipe’s weight and expansion or contraction do not af fect the equipment. For insulated pipes, heavy-gauge sheet-metal half-sleeves are used between the hangers and the insulation. Corrosion protection should also be carefully considered. Piping Insulation External pipe insulation should be pro vided when the inside w ater temperature 105 °F  Tw  60°F (41°C  Tw  15.6°C) for the sak e of ener gy sa ving, surface condensation, and high-temperature safety protection. The optimum thickness of the insulation of pipes depends mainly on the operating temperature of the inside water, the pipe diameter, and the types of service. There is a compromise between initial cost and energy cost. ASHRAE/IESNA Standard 90.1-1999 specif es the minimum pipe insulation thickness for w ater systems, as listed in Table 7.5. Insulation shall be protected from damage including that because of sunlight, moisture, equipment maintenance, and wind. TABLE 7.5 Minimum Pipe Insulation Thickness*, in. Fluid design operating temperature range, °F

Insulation conductivity Conductivity, Btuin./h ft2°F

Mean rating temp. °F

Nominal pipe or tube size, in. 95 percent. The f lter media are often made of glass f bers of submicrometer and micrometer diameter. They are often in the form of a pleated mat in a cartridge, as shown in Fig. 15.38, or in the form of a bag f lter. The air v elocity f owing through the f lter media is lo wer, and the minimum f nal pressure drops across the f lter media are 1.4 in. WC (350 P a). High-eff ciency air f lters are often protected by lo w- or medium-ef f ciency pre f lters to e xtend their service life. High-eff ciency f lters are widely used in air systems for hospitals, high-demand commercial buildings, and precision manufacturing workshops.

Ultrahigh-Efficiency Air Filters Ultrahigh-eff ciency f lters include high-ef f ciency particulate air (HEP A) f lters, ultralow penetration air (ULP A) f lters, and gaseous adsorbers and chemisorbers. Activated carbon adsorbers are discussed in Sec. 15.16, and chemisorbers are discussed in Sec. 24.5. HEPA f lters have an eff ciency of 99.97 percent for dust particles  0.3 m using the DOP test method. ULPA f lters have an ef f ciency of 99.999 percent for dust particles  0.12 m using the DOP method. A typical HEPA f lter is shown in Fig. 15.38. The dimensions of this f lter are 24 in. by 24 in. by 11.5 in. (600 by 600 by 287 mm). The f lter media are made of glass f bers of submicrometer diameter that are formed into pleated paper mats. Some of the lar ger f bers act as the carrier of the web . The performance of the f lter medium is often assessed by an inde x called the alpha value , which can be calculated as follows:



SH__ ST__ LG__ DF

(2  log P)100 25.4 pt

(15.51)

where pt  pressure drop of the f lter medium, in. WC (Pa). Both penetration P, in percent, and pressure drop pt are measured at an air v elocity f owing through the f lter medium of 10.5 fpm (0.05 m /s). The  value is usually between 10 and 11. Because of the de velopment of ne w media with lower pressure drops, an  value of 13 or e ven higher can no w be achieved. The surface area of the f lter medium may be 50 times the f ace area of the ultrahigh-ef f ciency f lter, and the rated face velocity may vary from 190 to 390 fpm (0.95 to 1.95 m / s) for ultrahigh-eff ciency f lters at a pressure drop of 0.65 to 1.35 in. WC (162 to 337 P a) for clean f lters. The face velocity of highcapacity ultrahigh-eff ciency f lters can be raised to 500 fpm (2.5 m / s). The f lter media themselv es have eff ciencies higher than that of the mounted f lter. Sealing of the f lter pack within its frame and sealing between the frame and gask et are critical f actors that affect HEPA and ULP A f lter penetration and ef f ciency. Penetration of dusts represents aerosol passing through the medium and through pinholes in the f lter medium, as well as leaks between the pack and the frame and between the frame and the gasket. To e xtend the service life of an ultrahigh-ef f ciency f lter, it should be protected by either a medium-eff ciency f lter or two f lters: a low-eff ciency f lter and a medium-ef f ciency f lter located just upstream from the ultrahigh-ef f ciency f lter. The removal of lar ge particles in the pre f lter reduces the dust load and prolongs the life of the ultrahigh-eff ciency f lter. HEPA and ULPA f lters are used to remo ve air contaminants such as unattached viruses, carbon dust, combustion smoke, and radon progeny of particles in sizes  0.3 m. They are widely used in clean rooms and clean spaces for the microelectronics industry , pharmaceutical industry, precision manufacturing, and operating theaters in hospitals.

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pb

6/27/2000

AIR SYSTEMS: COMPONENTS — FANS, COILS, FILTERS, AND HUMIDIFIERS

15.69

15.15 ELECTRONIC AIR CLEANERS

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An electronic air cleaner uses the attraction between particles of opposite char ges. Dust particles charged within the cleaner attract and agglomerate to greater sizes at the collecting plates. They are therefore easily removed from the airstream. A typical electronic air cleaner is shown in Fig. 15.39. A high dc potential of 1200 V is supplied to the ionizing f eld. The positive ions generated from the ionizer wire char ge the dust particles. Right after the ionizing section, the dust particles come to a collecting section, which consists of several plates that are alternately grounded and insulated. A strong electric f eld is produced by supplying a dc potential of 6000 V to these plates. The positively charged dust particles are attracted by the grounded plates of opposite charge, and attach themselves to the plates. Because of the numerous points of contact, the bond between particles held together by intermolecular forces is greater than that between the particles and plates. Therefore, the dust particles agglomerate and grow to such sizes that the y are blown off and carried a way by the airstream. The agglomerates are then collected by a medium-ef f ciency air f lter located downstream from the collecting section. Kemp et al. (1995) sho wed that electronic air cleaners are ef f cient for removing small particles of 0.5 to about 8 m. When dust particle sizes were approached and e xceeded 10 m, the f lter’s eff ciency dropped from its highest peak. In a year-long test with the outdoor ambient air, no microbial growth was observed on any of the f lters. The pressure drop across the ionizer section and collecting section is lo w and ranges from 0.15 to 0.25 in. WC (37 to 62 P a) against an air v elocity of 300 to 500 fpm (1.5 to 2.5 m / s). Safety

FIGURE 15.39

An electronic air cleaner.

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measures must be pro vided for protection against the high dc potential. If the positi ve charges are not removed at the collecting section, the accumulation of these positi vely charged dust particles may build up a space char ge. This space char ge drives the char ged particles to the w alls and other building envelope surfaces in the conditioned space, causing them to be smudged.

15.16 ACTIVATED CARBON ADSORBERS Granular activated carbon adsorbers are most widely used to remove objectionable odors and irritating vapors (including indoor v olatile organic compounds, VOCs) of v ery small gaseous molecules from the airstream by adsorption. Adsorption is the physical condensation of a gas or v apor onto an activated substance. Activated substances are highly porous. One pound of extremely porous carbon contains more than 5,000,000 ft 2 (465,000 m2) of internal surface. Gas molecules diffuse to the micropores or macropores of acti vated carbon, bond to these surf aces, and come in contact with the carbon granules. One pound (0.45 kg) of acti vated carbon may adsorb 0.2 to 0.5 lb (0.1 to 0.25 kg) odorous gases. Activated carbon in the form of granules or pellets is made from coal, coconut shells, or petroleum residues. They are heated in steam and carbon dioxide to remo ve foreign matter such as hydrocarbons and to produce internal porosity . Granular acti vated carbon is placed in special trays, which slide easily into position, to form activated carbon beds that are sealed into the cell housing by face plates, as shown in Fig. 15.40. A typical carbon tray, which is 23 by 23 by 85 in. thick (0.58 by 0.58 by 0.015 m thick), weighs 12 lb (5.4 kg). Lo w- or medium-eff ciency air f lters are used as pref lters for protection. When air f ows through a typical assembly with a f ace velocity of 375 to 500 fpm (1.88 to 2.5 m /s), the corresponding pressure drops are between 0.2 and 0.3 in. WC (50 and 75 Pa). Activated carbon can also be mounted in f xed frames with perforated sheets in continuous pleats. Adsorption capacity is de f ned as the amount of carbon tetrachloride adsorbed by a given weight of acti vated carbon. For various odors, the adsorption capacity is also af fected by the operating temperature and humidity. In general, a higher humidity or higher operating temperature

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FIGURE 15.40 Granular activated carbon assembly.

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usually decreases the adsorption capacity of the activated carbon. The maximum operating temperature is 100°F (37.8°C). The removal eff ciency of an activated carbon adsorber is the ratio of pollutant concentrations in the air upstream and do wnstream from the absorber . Toluene is used as the surrogate for total volatile organic compounds. The removal eff ciency of the carbon bed remains relati vely constant during its service life. Service life is the operating period between when an acti vated carbon absorber be gins to remo ve pollutant and the time when the remo val ef f ciency drops belo w an acceptable minimum. When the removal eff ciency drops below an acceptable minimum, it must be replaced either by reacti vation or re generation. Reacti vation is the process of remo ving spent carbon and replacing it with fresh carbon. Re generation is the process by which spent carbon is converted to fresh carbon. Re generation can only be performed by the acti vated carbon manuf acturer. The simplest w ay to in vestigate whether the carbon needs to be replaced is to detect do wnstream the odor that the carbon f lter is supposed to remo ve. A sample of carbon can also be tested to determine its remaining capacity. Burroughs (1997b) reported that active particle fabrics demonstrate excellent indoor air contaminant control of VOC gases. These new fabric-based sorption f lter media can be pleated into v arious conf gurations to maximize sorption capacity and airf ow.

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15.17 SELECTION OF AIR FILTERS During the selection of air f lters, electronic air cleaners, and activated carbon f lters to remo ve air contaminants for the health and comfort of the occupants in indoor conditioned space, the following requirements and arrangements should be carefully considered: ●

●

●

●

●

Characteristics of the air contaminants, the size of the particles, and the concentration of dust particles or irritating vapors to be removed must be identif ed. The degree of air cleanliness required in the conditioned space must be speci f ed, especially the design criteria for clean spaces or clean rooms. The minimum eff ciency of the air f lter at speci f c particle sizes and loadings must be speci f ed. The initial, average, and f nal pressure drops during the operating period, which affect the energy consumption and service life of the f lter and the air system, must be determined. Service life of the air f lter inf uences the installation cost, the pressure drop, and the eff ciency of the air f lter. The following f ltration tactics are recommended:

●

●

●

●

●

Select an air f lter based on the dust particle size MERV to remove air contaminants. Burroughs (1997) recommended the use of medium-ef f ciency particulate f lters for protection of coils and air distrib ution systems, preventing the nutrition for biological gro wth; high-eff ciency particulate f lters for control of respirable particulate and bacteria; and gaseous absorbers to control objectionable odors and harmful volatile organic compounds. High-eff ciency and ultrahigh-ef f ciency air f lters, gaseous adsorbers such as granular acti vated carbon, and electronic air cleaners must be protected by a pref lter of medium eff ciency to extend the life of these high-ef f ciency and ultrahigh-ef f ciency gaseous adsorbers and electronic air cleaners. Ho wever, the addition of tw o lo w- or tw o medium-ef f ciency f lters does not signi f cantly improve the eff ciency to collect submicron dust and irritating vapors. Monitoring the pressure drop of the air f lters and periodic maintenance of lo w-eff ciency viscous air f lters have a direct impact on f lter performance. Muller (1995) reported that granular acti vated carbon is very good to remove most hydrocarbons, many aldehydes and or ganic acids, as well as nitrogen dioxide. It is not particularly ef fective against sulfur oxide, nitric oxide, formaldehyde, hydrogen sul f de, or lo wer-molecular-weight

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aldehydes and or ganic acids. Potassium permanganate – impregnated alumina (chemisorbers are discussed in Sec. 24.5) is ef fective against sulfur and nitric oxides, hydrogen sul f de, and lowmolecular-weight aldehydes.

15.18 HUMIDIFICATION AND HUMIDIFIERS Space Relative Humidity As mentioned in Sec. 4.9, for comfort air conditioning systems the space relative humidity is tolerable between 35 and 65 percent in summer , preferably from 40 to 50 percent. During winter , for space served by a comfort air conditioning system that is installed with a humidi f er, the space relative humidity should not e xceed 30 percent e xcept in hospitals. The actual humidifying capacity of a humidif er should not exceed the humidifying requirement so that wet surfaces do not occur inside the AHU, PU, or supply ducts. Wet surfaces and dirt often cause the growth of microorganisms and thus indoor air quality problems. For air conditioning systems without humidi f ers, space relati ve humidity is usually not specif ed. For processing air conditioning systems, the space relati ve humidity should be speci f ed as required by the manufacturing process. If the space temperature is maintained at 72 °F (22.2°C) and only free con vection is provided in the occupied space, the space relati ve humidity should not e xceed 27 percent when the outdoor temperature is 30 °F ( 1.1°C), in order to pre vent condensation on the inner surf ace of singleglazed windo ws during winter . To pre vent condensation on the inner surf ace of double-glazed windows, the space relative humidity should not exceed 33 percent when the outdoor temperature is 0°F (17.8°C).

Humidifiers A humidif er adds moisture to the air. Humidif ers may (1) inject steam directly into air or add heat and evaporate steam from w ater supplied to the conditioned space; (2) atomize or spray liquid w ater, so that water evaporates and is added to the air; or (3) force air to f ow over a wetted element so that as water evaporates, it is added to the air as v apor. All these increase the humidity ratio of the space air and, therefore, its relative humidity. One important inde x of a humidi f er is its humidifying capacity m˙ cap in pounds of w ater per hour , lb / h (kg / h), or the rate at which w ater v apor is added to the air.

Humidifying Load Humidifying load m˙ hu , in lb / h or kg / h, is the amount of water vapor required to be added to the air by a humidif er so as to maintain a predetermined space relati ve humidity. The humidifying load of an air system installed with a humidif er can be calculated as (see Fig. 15.41) m˙ hu  60V˙s s (wlv  wen)  60V˙s s (ws  wm )

SH__ ST__ LG__ DF

where V˙s  supply volume f ow rate of AHU or PU, cfm (m3 /min) s  density of supply air, lb / ft3 (kg / m3) wen, wlv  humidity ratio of air entering and leaving humidif er, lb / lb (kg / kg) ws  humidity ratio of supply air, lb / lb (kg / kg) wm  humidity ratio of mixture of outdoor air and recirculating air, lb / lb (kg / kg)

(15.52)

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FIGURE 15.41

Humidifying load for a typical air system.

In Fig. 15.41, s is the state of supply air when the space has a heating load, and s is the condition when the space has a cooling load, ws  ws. The humidifying load can also be calculated from the following relationship: m˙ hu  60(V˙o  V˙inf )o (wr  wo)  m˙ wr

(15.53)

where V˙o  supply volume f ow rate of outdoor air intake, cfm (m3/min) V˙inf  volume f ow rate of inf ltrated air, cfm (m3/min) wr,wo  humidity ratio of space air and outdoor air, lb / lb (kg / k g) In Eq. (15.53), m˙ wr represents the space moisture gains, in lb / h (kg / h). Space moisture gains include the latent load from the occupants, appliances, equipment, and products. The moisture gains from the building structures are often ignored. Types of Humidifier According to the mechanism used for evaporation of water vapor from water, humidif ers can be classif ed as steam and heating element humidi f ers, atomizing humidif ers, and wetted element humidif ers.

15.19 STEAM AND HEATING ELEMENT HUMIDIFIERS At a lo w partial pressure when mix ed with dry air , steam is the ready-made w ater v apor. With proper water treatment at the boiler , steam is free of mineral dust and odor , and it does not support the growth of bacteria that create sanitation problems. Two types of steam humidif ers are currently used in air systems: steam grid humidif ers and steam humidif ers with separators. Steam Grid Humidifiers A steam grid humidif er installed inside ductwork is shown in Fig. 15.42a. A steam grid humidif er may have a single distrib ution manifold or multiple manifolds. Each distrib ution manifold has an

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FIGURE 15.42 Steam grid humidif ers: (a) steam grid and piping connections; (b) cross section of jacketed distribution manifold.

inner steam pipe and an outer jack et. The inner steam tube is connected to a header and controlled by a control v alve. Steam is supplied at a pressure less than 10 psig (69 kP ag) and throttled to a lower pressure after the controlled v alve. It then enters the inner tube and dischar ges through the small holes to the outer jack et. The dry steam is again dischar ged through the ori f ces of the jack et to the ambient airstreams to humidify them, as shown in Fig. 15.42 b. The condensate inside the jacket is dischar ged to a drainpipe and steam trap located at the opposite side. The inner tube and the outer jacket are slightly pitched toward the drainpipe and steam trap.

Steam Humidifiers with Separators

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A steam humidi f er with a separator is illustrated in Fig. 15.43. Steam is supplied to a jack eted distribution manifold and then enters a separating chamber with its condensate. It then f ows through a control v alve, is throttled to slightly abo ve atmospheric pressure, and enters a drying chamber. Because of the lo wer pressure and temperature in the drying chamber compared with the higher pressure and temperature in the surrounded separating chamber , the steam is superheated. Dry steam is then dischar ged into the ambient airstream through the ori f ces of the inner steam discharging tubes. Noise is produced mainly during the throttling of the steam at the control v alve. It is attenuated as steam f ows through the silencing materials.

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FIGURE 15.43 Steam humidif er with a separator. (Adapted with permission from The Armstrong Humidif cation Handbook, 1991.)

Heating Element Humidifiers When a steam boiler is not a vailable or is too e xpensive to install, steam humidi f cation is still possible by using heating element humidi f ers. A heating element humidi f er has a w ater pan with electric or gas-f red heated elements installed at the bottom of the pan to e vaporate the liquid water to w ater v apor and then add the v apor to the airstream f owing o ver the w ater pan, as sho wn in Fig. 15.44. The humidifying capacity of a heated element humidi f er is limited by its heat-transfer surface and the characteristics of the heating element.

FIGURE 15.44 Heating element humidif er.

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Characteristics and Requirements of Steam and Heating Element Humidifiers Steam humidif ers are widely used in commercial and industrial applications. Both steam and heating element humidif ers need heat energy from gas, fossil fuel, or electricity to evaporate the liquid water in the steam generator. The temperature of the ambient air or airstream in which moisture has been added is approximately unchanged. Steam and heating element humidi f ers are therefore often called isothermal humidif ers. For a conditioned space that has a cooling load year round and needs a cold air supply in winter , it is a w aste of energy to evaporate the liquid w ater by using fossil fuel instead of the excess heat gains at the conditioned space. Control is essential in steam and heating element humidi f ers. The capacity of a steam humidif er is often controlled by a microprocessor-based controller by modulating the valve pin of the control valve according to the signal of a humidity sensor and, therefore, the steam f ow rate. Morton (1996) reported an accuracy of humidity control of 5 to 7 percent for on / off control and 3 to 5 percent for modulation control. The size of the control v alve should be carefully matched with the humidifying load with an adequate turndo wn ratio. The control valve should be inte grated with the humidif er and steam-jack eted at supply pressure to pre vent condensation. An interlocking control should be installed to drain all condensate before the steam humidif er is started. For an electric heating element humidi f er, or simply an electric humidi f er, the input w attage can be adjusted when the space relative humidity is too high or too low. A gas-f red heating element humidif er, or simply a gas- f red humidif er, uses infrared b urners located within a heat e xchanger. A microprocessor-based controller stages and c ycles the b urner to v ary the heat output and therefore the humidifying capacity during part-load operation. The makeup water supply to the steam boiler and the w ater pan must be well treated so that the steam is free of mineral deposits and odor . F or electric and gas- f red humidi f ers, according to DeBat (1996), recent de velopments, such as ionic bed technology , overcome the problem of mineral deposits. Dissolved mineral salts begin to precipitate out of w ater when their concentration becomes too high. Ionic beds consist of inert f bers that attract these precipitated minerals. The mineral salts deposit on the ionic bed instead of on the w alls of the heating chamber . When throughly encrusted with crystalline solids, the ionic bed cartridge is easily remo ved and discarded. Steam humidif ers should be located where noise is not objectionable. Access to equipment for inspection and maintenance should be provided. Condensate drainage must be properly designed. If the steam humidi f er is located in ductw ork, install a duct high-limit humidistat and an interlock ed airf ow proving switch. The humidity sensor is located preferably in the return air duct or a representative location.

15.20 ATOMIZING AND WETTED ELEMENT HUMIDIFIERS Humidification Process

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Atomizing means producing a f ne spray. When liquid w ater is atomized, the smaller the diameter of the w ater droplets, the greater the interf acial area between w ater and air and thus the higher the rate of evaporation and humidif cation. When air f ows through an atomized w ater spray from an atomizing humidi f er, if the dif ference between the temperature of the w ater spray and the wet-b ulb temperature of the ambient air is small, the result is an increase in the humidity ratio of air from the addition of e vaporated water vapor and a corresponding drop in air temperature because of the absorbtion of required latent heat of vaporization from the ambient air . Such a humidi f cation process is an adiabatic saturation or e vaporative cooling process and follows the thermodynamic wet-bulb temperature line on the psychrometric chart. Atomizing humidi f ers can be classi f ed, according to their con f guration and mechanism of atomizing, as ultrasonic humidi f ers, centrifugal atomizing humidi f ers, pneumatic atomizing humidif ers, and air washers.

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Ultrasonic Humidifiers

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An ultrasonic humidif er uses a piezoelectric transducer to convert high-frequency mechanical electric signals to a high-frequency ocscillation. The transducer is submerged into a water pan in which water is well treated and demineralized. The ne gative oscillation produces a momentary v acuum that causes the w ater to cavitate into vapor at low pressure. The positive oscillation creates a highcompression w ave which forces the w ater particles to lea ve the w ater surf ace and inject into the airstream. Water f lms are then brok en into a f ne mist. Air containing this mist can be blo wn into the space directly or through an air system by a fan. If the w ater is well treated, it will be free of mineral deposits, and no f lter is required do wnstream from the humidif er. Ultrasonic humidif ers create less equipment noise than other atomizing humidif ers. In the mid-1990s, ultrasonic humidi f cation had been installed in telecommunication projects. Ultrasonic atomizing humidi f ers with small humidifying capacities ha ve been used directly in residences.

Case Study: White Plains Ultrasonic Humidification Project Longo (1994) reported on a retro f t ultrasonic humidi f cation project which is located in a 1950s building in White Plains, New York. The humidif cation system in a totally ne w air system serv es space that houses solid-state comple x multiswitching processors. Static char ges created in a lo whumidity environment can be detrimental to system components and cause failure. The 100,000-cfm (47,200-L / s) air system consists of four supply f ans (tw o standby), higheff ciency f lters, and cooling coils with outside air economizer for free cooling. The ultrasonic humidif ers were placed after the f lters and upstream from the supply f an. There are altogether 43 ultrasonic units mounted on stainless-steel racks, each rated at 39.6 lb / h (18 kg / h) and 960 W. A humidity sensor is installed in the return air duct. The controller turns the humidity system on one stage at a time up to a total of six stages. A high-limit sensor set at 80 percent relati ve himidity shuts down the system in case of overhumidif cation. Water is treated by a series of pre f lters, a reverse osmosis system, and deionization canisters. Water quality is controlled by a resisti vity probe. When w ater resistance reaches 0.02 M , the deionization canisters will be changed. After completion, the actual running results in winter are e xcellent. As the equipment w as installed gradually, the supply air temperature was rescheduled, and the space relative humidity was maintained precisely at 50 percent. The initial cost of the White Plains humidity system w as $350,000. Annual operating cost was $23,000, which was less than one-tenth of the annal operating cost of an electric heating element humidity system.

Centrifugal Atomizing Humidifiers These humidif ers use the centrifugal force produced by a rotating de vice such as a rotating cone, blades of an axial fan, rotating disk, or rotating drum to break the liquid w ater f lm into f ne mist or to f ing it into f ne water droplets. When air is forced through an atomizing w ater spray, produced by a pulv erizing fan, water vapor is evaporated and added to air as a result of the mass concentration difference between the saturated air f lm at the surface of the droplets and the ambient air . The rotating device receives liquid w ater either from a pressurized w ater supply or by dipping into the surface of a nonpressurized supply. The humidifying capacity of a pulv erizing fan depends mainly on the volume f ow rate of the f an. Because of the o versaturation characteristics of the pulv erizing fan, m˙ cap may vary from 50 to 150 lb / h (23 to 68 kg / h). Oversaturation is the e xcess amount of water particles present in the moist air , and it is discussed in Chap. 20. Man y different types of rotating humidif ers with limiting humidifying capacities are used directly in the conditioning space in residential and industrial applications. Most are portable.

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Maintaining a predetermined space relati ve humidity with a centrifugal atomizing humidi f er can be performed by modulating the f ow rate of water supplied to the rotating devices.

Pneumatic Atomizing Humidifiers A typical pneumatic atomizing nozzle consists of tw o concentric brass tubes, as sho wn in Fig. 15.45a. The inner tube has an outer diameter of 323 in. (2.4 mm) and an inner diameter of 321 in. (0.8 mm). It is connected to a w ater tank that can be mo ved to adjust the dif ference in water levels between the w ater tank and the centerline of the nozzle. The conical outer tube has a minimum inner diameter of 81 in. (3.2 mm) and is connected to a compressed air line with a pressure at 15 psig (103 kPag). When compressed air is dischar ged from the annular slot at a v ery high v elocity, it extracts water from the inner tube and breaks the w ater into a v ery f ine mist. A typical pneumatic atomizing nozzle may ha ve a humidifying capacity m˙ cap from 6 to 10 lb / h (2.7 to 7.5 kg / h). The magnitude of m˙ cap is af fected by the conf iguration of the nozzle, the pressure of the compressed air, and the dif ference in w ater le vels. In a pneumatic humidif ier, m˙ cap is usually controlled by adjusting the dif ference in w ater levels between the w ater tank and the centerline of the nozzle. A pneumatic humidi f er produces high-frequenc y noise. This type of humidi f er is usually applied in industries with a certain sound le vel of machine noise and can be used for direct in-space humidif cation.

Wetted Element Humidifiers Wetted element humidi f ers include a wetted element, such as an e vaporative pad, plastic, or impregnated cellulose, that is dipped with w ater from the top. Such humidi f ers have been installed in air-handling units and packaged units to humidify the air . Characteristics of wetted element humidif ers are similar to those of wetted element e vaporative coolers (see Chap. 27). Modulation of the w ater supply to the w ater dipping de vice varies the humidifying capacity and maintains the desirable space relative humidity.

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FIGURE 15.45

Pneumatic atomizing humidif er: (a) nozzle; (b) pneumatic atomizing humidifying system.

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15.21 AIR WASHERS

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The air washer was the f rst air conditioning equipment de veloped by Carrier in 1904, and it is still used today to humidify, cool, and clean the air in many factories.

Construction of an Air Washer An air washer has an outer casing, shown in Fig. 15.46a. It is usually made of plastic or galv anized steel sheet with w ater-resistant paint for protection. All joints are well sealed by w ater-resistant resin. A water tank either forms a part of the casing at the bottom or is installed separately on the f oor to collect or sometimes to mix the recirculating and incoming chilled w ater. A separately installed water tank is usually made of steel, stainless steel, or reinforced concrete with an insulation layer mounted on the outer surf ace if chilled w ater is used for spraying. A bank of guided baf f es installed at the entrance provides an evenly distributed air-water contact. Eliminators in the shape of a sinusoidal curve at the exit are installed to remove entrained water droplets from the air. They are preferably made of plastic or stainless steel for con venient cleaning and maintenance. Access doors

FIGURE 15.46

Schematic diagram of a typical air washer: (a) air washer; (b) spray nozzle.

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mounted on the outer casing must be provided for inspection and maintenance. They should be well sealed to prevent water leakage. In most air w ashers, two banks of spraying nozzles f ace each other. Nozzles are often made of brass, plastic, or nylon. The nozzle consists of a cover with an orif ce diameter between 116 and 316 in. (1.6 and 4.8 mm), a discharge chamber, and an inlet with its centerline slightly eccentric from the centerline of the ori f ce, as shown in Fig. 15.46 b. This offset causes the w ater stream to rotate inside the discharge chamber and breaks the w ater into f ne droplets. A larger orif ce diameter does not clog easily , and a smaller ori f ce can produce f ner w ater sprays. The distance between tw o spraying banks is from 3 to 4.5 ft (0.9 to 1.35 m), and the total length of the air w asher varies from 4 to 7 ft (1.2 to 2.1 m). A centrifugal water pump is used to recirculate the spraying w ater. For air containing much dirt and lint, as in air conditioning systems in the te xtile industry, an automatic washing and collecting water strainer, using copper or brass f ne-mesh screen, often precedes the pump. The opening of the f ne-mesh screen should be smaller than the diameter of the hole of spraying nozzles to a void clogging.

Functions of an Air Washer Currently, air washers are used to perform one or more of the following functions: ●

●

●

Cooling and humidif cation Cooling and dehumidif cation Washing and cleaning

Whether it is a cooling and humidi f cation or cooling and dehumidi f cation process is determined by the temperature of the spraying w ater. If recirculating w ater is used as the spraying w ater, the temperature of w ater then approaches the wet-b ulb temperature of the air entering the air w asher, and the air is humidif ed and evaporatively cooled. If chilled w ater or a mixture of chilled w ater and recirculating w ater is used for spraying water, and its temperature is lo wer than the de w point of the entering air , the air is cooled and dehumidif ed. Water spraying is sometimes practiced for the purpose of w ashing and cleaning air when an objectionable gas is kno wn to be soluble in w ater and is to be remo ved from the air . For example, exhaust gas from many industrial applications may be removed by water spraying. For humidi f cation purposes, nozzles must ha ve a smaller ori f ce and a higher w ater pressure, such as a 116 -in. (1.6-mm) diameter and a pressure of 40 psig (275 kP ag). For cooling and dehumidif cation, a larger ori f ce and lo wer w ater pressure, such as a 316 -in. (4.8-mm) ori f ce and 25 psig (172 kPag), are often used.

Performance of an Air Washer For a humidi f cation process along the thermodynamic wet-b ulb temperature line, the performance of an air washer can be illustrated by the saturation eff ciency εsat and calculated as sat 

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w2  w1 T2  T1  w*s  w1 T* s  T1

(15.54)

where T1, T2  temperature of air entering and leaving air washer, °F (°C) w1,w2  humidity ratio of air entering and leaving air washer, lb / lb (kg / kg) ws*, Ts*  humidity ratio and temperature of saturated air at thermodynamic wet-bulb temperature, lb / lb (kg / kg) and °F (°C), respectively

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For a cooling and dehumidif cation process, the performance of an air washer is better described by a factor called the performance factor Fp as follows: Fp 

h en  h lv h en  h *s

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(15.55)

where hen, hlv  enthalpy of air entering and leaving air washer, Btu / lb (kJ / kg) hs*  enthalpy of saturated air at chilled water spraying temperature, Btu / lb (kJ / kg) The performance of an air w asher for a humidi f cation process, indicated by the saturation ef f ciency εsat, and for a cooling and dehumidi f cation process, expressed by Fp, depends mainly on the contact time between air and w ater and the contact surf ace area, i.e., the w ater-air ratio m˙ w /m˙ a (which indicates the mass f ow rate of the spraying w ater to the mass f ow rate of the air f owing through the air washer, in same units), the length of the air w asher, and the direction of spray with respect to airf ow, (whether it is opposing or following the airf ow). An air w asher is usually designed at an air v elocity between 500 and 800 fpm (2.5 and 4 m /s) with respect to its cross-sectional area at the w ater sprayers. Its greater w asher length and the attached water tank result in a b ulkier volume than that of DX coils and w ater cooling coils. Total pressure loss of the airstream f owing through an air w asher pt,w, in in. WC (Pa), depends mainly on the con f guration of the eliminators and the air v elocity f owing through them. Usually pt,w varies from 0.25 to 1 in. WC (62.5 to 250 Pa), and typically it is 0.5 in. WC (125 Pa). For humidif cation and evaporative cooling, the water-air ratio m˙ w /m˙a is usually 0.3 to 0.6. F or a 5-ft (1.5-m) air washer with m˙ w /m˙ a  0.45, that is, 4 gpm of w ater per 1000 cfm of air (0.534 L / s of w ater per 1000 L / s of air), εsat  0.85 to 0.9. F or cooling and dehumidi f cation, the water-air ratio is usually 0.5 to 1.2. F or a 5-ft- (1.5-m-) long air w asher with m˙ w /m˙ a  0.9, that is, 8 gpm / 1000 cfm (1 L / s of water per 1000 L / s of air), Fp  0.5 to 0.75. For either the humidi f cation process or the cooling and dehumidi f cation process in an air washer, oversaturation always exists at the e xit. Oversaturation (wo  ws*) usually varies between 0.0002 and 0.001 lb / lb (0.0002 and 0.001 kg / kg). It depends mainly on the construction of the eliminator and the f ow-through air v elocity. Ov ersaturation is bene f cial to a humidi f cation process with space heat gains and interferes with a cooling and dehumidi f cation process. Because of the existence of oversaturation, the saturation effectiveness is no longer the dominant index that affects the performance of the air w asher during humidi f cation. Ov ersaturation is discussed in Sec. 20.13.

Bypass Control Bypass control is def ned as modulating the face damper and the bypass damper of the air washer to vary the proportion of humidifying air and bypass air in order to maintain the space relati ve humidity within predetermined limits. Such a control is often used in an air w asher to pro vide better humidity control than modulating the water pressure or the water f ow rate to the spraying nozzles.

Single-Stage or Multistage A group of nozzles that are supplied with w ater at the same pressure and produce spray at the same temperature are said to be in the same stage. A stage may ha ve one or tw o spraying banks. A twostage air washer is illustrated in Fig. 15.47. For humidif cation processes or for cooling and dehumidi f cation processes using chilled w ater from the refrigeration plant, it is more economical to operate a single-stage air w asher, usually with two water-spraying banks. When low-temperature deep well or groundwater is used for cooling and dehumidif cation, it is often adv antageous to use tw o-stage, or even three-stage, water spraying to make full use of the cooling capacity of the groundwater.

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FIGURE 15.47

Two-stage air washer using deep well water.

Groundwater can be used again as condenser w ater, or as f ushing water in plumbing systems after its dischar ge from the air w asher follo wing necessary w ater treatment. Used groundw ater should be drained to another well (rechar ge well) in the vicinity to pre vent subsidence and damage to the soil, which may affect the foundation of the building in the vicinity of the deep well. The use of groundwater is regulated by local and federal environmental protection agencies.

15.22 CHARACTERISTICS OF ATOMIZING AND WETTED ELEMENT HUMIDIFIERS Atomizing and wetted element humidif ers, including air washers for humidif cation purposes have the following operating characteristics: ●

●

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Conditioned air follo ws the thermodynamic wet-b ulb temperature line on a psychrometric chart. The humidity ratio of the air increases while it is e vaporatively cooled. Atomizing humidif ers are often called nonisothermal humidi f ers. This characteristic indicates that atomizing and wetted element humidif ers are especially suitable for conditioned spaces that need a cold air supply to offset a space cooling load during or after the humidi f cation process. The heat energy required to evaporate the liquid water, as in steam or heating element humidi f ers, has been saved. For a conditioned space which needs a w arm air supply during winter operation, the temperature of e vaporatively cooled air after an atomizing or wetted element humidi f cation process should be raised before supplying it to the conditioned space, so as to offset the transmission loss through the b uilding envelope. The heat energy required is approximately equal to that required to e vaporate water in a steam boiler or in a water pan when a steam humidif er or a heating element humidif er is used. For atomizing humidi f ers that need a cold air supply during winter , when the humidifying capacity is modulated, both the space relati ve humidity and temperature v ary accordingly. Many

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●

●

industrial manufacturing processes need to maintain a stable space relati ve humidity (RH). RH is often the controlled v ariable, and temperature is usually maintained within predetermined limits. Oversaturation exists at the e xit of the atomizing humidi f er even if an eliminator is used. Ov ersaturation is often adv antageous for humidi f ers installed in conditioned spaces ha ving e xcess heat gains because it increases the humidifying capacity. The size of the w ater droplets, the local humidifying capacity, and the excessive heat gains determine whether liquid w ater droplets may f all on the f oor, products, or equipment. According to f eld measurements, the diameter of w ater droplets suspended in the airstream dischar ged from atomizing humidif ers is as follows: Pneumatic atomizing humidif ers Centrifugal atomizing humidif ers

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30 to 80 m 50 to 300 m

When the diameter of the w ater droplets e xceeds 80 m and the local humidifying capacity is excessive, water may drip. Water dripping must be a voided, even if reducing the humidifying capacity is necessary . It causes product damage and corrosion of appliances, instruments, and equipment. In the water tank and water pan of atomizing and wetted element humidi f ers, algae, bacteria, and other microor ganisms may gro w. Water spray and the airstream dischar ged from the atomizing and wetted element humidi f ers may distrib ute the bacteria and sometimes odors. Water treatments, regular cleaning and f ushing, and periodic inspection and blo wdown help to minimize these problems. The designer must include convenient access for such maintenance.

15.23 SELECTION OF HUMIDIFIERS AND DESIGN The following factors should be considered in selecting a humidif er: ●

●

●

●

●

●

●

●

●

Energy consumption and oper ating cost. For a space that needs a cold air supply e ven in winter, atomizing and wetted element humidi f ers sa ve ener gy and, therefore, have lo wer operating costs. Among the atomizing humidif ers, pneumatic atomizing humidif ers require compressed air to atomize the liquid w ater and, therefore, have a higher operating cost than other atomizing humidif ers. Quality of humidi f cation. Humidif ed air should be clean and free of odor , bacteria, particulate matter, and water droplets to pre vent microbial gro wth due to wetted surf aces or standing w ater and improving IAQ. Humidifying capacity. Is a large, medium, or small humidifying capacity required? Is o versaturation benef cial in the conditioned space? Capacity control. Is this an on / off control or a modulation control, i.e., a stepless control? Does the control have a fast or slow response? What is the turndown ratio of the control valve? Morton (1996) recommended that the best location of a humidi f er in an AHU or PU is in the f rst section of supply main duct do wnstream from the supply fan. The next choice is between the coil section and the supply fan if there is no oversaturation in the airstream. Equipment noise . For in-space humidi f ers such as pneumatic humidi f ers, equipment noise is a primary concern. Initial cost. For a steam humidif er, the initial cost should also include the installation costs of the steam boiler and the corresponding steam piping system. Maintenance. This includes the amount and cost of required maintenance w ork for the humidif ers. Space occupied. This is the volume occupied by the humidif er per unit humidifying capacity.

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Description

Heating element

Steam

Humidifying capacity Oversaturation Quality of humidif cation Dripping of water in conditioned space Energy consumption and operation cost Initial cost Maintenance work and maintenance cost Capacity control Response to control Space occupied Equipment noise

Air washer

Pneumatic atomizing

Ultrasonic

Small to large No Good

Small No Good

Large Yes Good

Small to large Possible Good

Medium and large Possible Good

No

No

No

No

No

Need more energy if cold air supply is needed, so operating cost is higher High Low

Same as steam

Saving energy if cold air supply is needed in winter

Saving energy if cold air supply is needed in winter

Saving energy if cold air supply is needed in winter

High Medium

Medium High

Medium Medium

High Medium

Modulation

On/off or modulation Slow Medium Low

Modulation

Multiple stages

Modulation

Medium Large Medium

Fast Small Low

Fast Medium High

Fast Small Medium

Various types of humidi f ers are compared in Table 15.4. F or commercial b uildings that need a warm air supply during winter and for which a steam supply is a vailable, a steam humidi f er is often the most suitable choice. F or a small packaged unit serving a health care f acility, a heating element humidif er may be selected. F or a telecommunications center with an appreciable amount of space heat gain, an ultrasonic humidif er is often the right choice. Ultrasonic humidi f ers are also often used in residences for their comparati vely low equipment noise. F or a factory in which space air contains lint and other dirt that should be cleaned, or for which a cold air supply is needed to offset the cooling load in winter, an air washer may be the suitable choice. When air w ashers or atomizing humidi f ers are selected as the humidi f ers, oversaturation may result in the supply air dischar ged from the air w asher or atomizing humidi f er. As oversaturation causes wetted surfaces and dampness along the air distrib ution passage, the following remedies are recommended to prevent IAQ problems: ●

●

●

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All-metal (f ber-free) silencer or active silencer should be used, if required. Interior surfaces of ducts that contact supply air must be galvanized metal sheets. Duct heat gain may be benef cial to eliminate oversaturation. Periodic monitoring of the moisture and dirt conditions of the ducts and silencers should be provided.

ASHRAE / IESNA Standard 90.1-1999 mandates that when a zone is serv ed by a system or systems with both humidi f cation and dehumidi f cation, means (such as limit switches, mechanical stops, software program for DDC control) shall be pro vided to prevent simultaneous operation of humidif cation and dehumidi f cation equipment. Exceptions are zones serv ed by desiccant dehumidi f cation systems, which are used with direct e vaporative cooling in series, or zones where speci f c humidity levels are required such as computer rooms, museums, and hospitals. Standard 90.1-1999 also mandates that humidif ers with preheating jackets mounted in the airstream shall ha ve an automatic valve to shut off preheat operation when humidif cation is not in use.

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REFERENCES ANSI/ASHRAE Standard 52.2-1999, Method of Testing General Ventilation Air-Cleaning Devices for Removal Eff ciency by Particle Size, ASHRAE Inc., Atlanta, GA, 1999. ARI Guideline — Fouling Factors, Heating / Piping / Air Conditioning, no. 2, 1988, pp. 109 – 110. Armstrong Machine Works, The Armstrong Humidif cation Handbook, Armstrong Machine Works, Three Rivers, MI, 1991. ASHRAE, ASHRAE Handbook 1996, HVAC Systems and Equipment, ASHRAE Inc., Atlanta, GA, 1996. ASHRAE, ASHRAE Handbook 1997, Fundamentals, Atlanta, 1997. Avery, R. H., Selection and Uses of HEPA and ULPA Filters, Heating / Piping / Air Conditioning, no. 1, 1986, pp. 119 – 123. Branda M. R., A Primer on Adjustable Frequency Inverters, Heating / Piping / Air Conditioning, no. 8, 1984, pp. 83 – 87. Braun, J. E., Klein, S. A., and Mitchell, J. W., Effective Models for Cooling Towers and Cooling Coils, ASHRAE Transactions, 1990, Part I, pp. 164 – 174. Braun, R. H., Problem and Solution to Plugging of a Finned-Tube Cooling Coil, ASHRAE Transactions, 1986, Part I B, pp. 385 – 387. Brown, W. K., Humidif cation by Evaporation for Control Simplicity and Energy Savings, ASHRAE Transactions, 1989, Part I, pp. 1265 – 1272. Burroughs, H. E. B., IAQ: An Environmental Factor in the Indoor Habitat, HPAC, no. 2, 1997a, pp. 57 – 60. Burroughs, Filtration: An Investment in IAQ, HPAC no. 8, 1997b, pp. 55 – 65. Chen, J. C., Correlation for Boiling Heat Transfer to Saturated Fluids in Convective Flow, Ind. Eng. Chem. Process Design Develop., vol. 5, no. 3, 1966. Coward, Jr., C. W., Unhoused (Plug / Plenum) Fans: Is Their Performance Predictable? ASHRAE Journal, no. 10, 1997, pp. 71 – 76. Davis, W. J., Water Spray for Humidif cation and Air Flow Reduction, ASHRAE Transactions, 1989, Part II, pp. 351 – 356. Deacon, W. T., Important Considerations in Computer Rooms Humidif cation, ASHRAE Transactions, 1989, Part I, pp. 1273 – 1277. DeBat, R. J., Humidity: The Great Equalizer, HPAC, no. 10, 1996, pp. 66 – 71. Delaney, T. A., Maiocco, T. M., and Vogel, A. G., Avoiding Coil Freezeup, Heating / Piping / Air Conditioning, no. 12, 1984, pp. 83 – 85. Goldf eld, J., Use Total Pressure When Selecting Fans, Heating / Piping / Air Conditioning, no. 2, 1988, pp. 73 – 82. Goswami, D., Computer Room Humidity Control, Heating / Piping / Air Conditioning, November 1984, pp. 123 – 124. Green, G. H., The Positive and Negative Effects of Building Humidif cation, ASHRAE Transactions, 1982, Part I, pp. 1049 – 1061. Grimm, N. R., and Rosaler R. C, Handbook of HVAC Design, McGraw-Hill, New York, 1990. Gurock, D. R., and Aldworth, D. R., Fan-Motor Combination Saves Energy, Heating / Piping / Air Conditioning, no. 11, 1980, pp. 53 – 57. Hartman, T., Humidity Control, Heating / Piping / Air Conditioning, September 1989, pp. 111 – 114. Holtzapple, M. T., and Carranza, R. G., Heat Transfer and Pressure Drop of Spined Pipe in Cross Flow, Part III: Air Side Performance Comparison to Other Heat Exchangers, ASHRAE Transactions, 1990, Part II, pp. 136 – 141. Hunt, E., Benson, D. E., and Hopkins, L. G., Fan Eff ciency vs. Unit Eff ciency for Cleanroom Application, ASHRAE Transactions, 1990, Part II, pp. 616 – 619. Kemp, S. J., Kuehn, T. H., Pul, D. Y. H., Vesley, D., and Streifel, A. J., Filter Collection Eff ciency and Growth of Microorganisms on Filters Loaded with Outdoor Air, ASHRAE Transactions, 1995, Part I, pp. 229 – 238. Larocca, D. V., Chilled Water Coil Freeze Protection via Internal Drying, HPAC, no. 12, 1997, pp. 67 – 72. Longo, F., Ultrasonic Humidif cation for Telecommunications, Heating / Piping / Air Conditioning, no. 3, 1994, pp. 65 – 66.

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McQuinston, F. C., Finned Tube Heat Exchangers: State of the Art for the Air Side, ASHRAE Transactions, 1981, Part I, pp. 1077 – 1085. Morton, B. W., Humidif cation Handbook, B. W. Morton and Dri-Steam Humidifying Co., Hopkins, MN, 1981. Morton, C. B., Six Steps to Follow That Ensure Proper Humidif cation System Design and Control, ASHRAE Transactions, 1996, Part II, pp. 618 – 627. Muller, C. O., and England, W. G., Achieving Your Indoor Air Quality Goals: Which Filtration System Works Best? ASHRAE Journal, no. 2, 1995, pp. 24 – 32. Nader, J. C., and Kanis, T. W., Industrial vs. Commercial Fans, Heating / Piping / Air Conditioning, no. 2, 1983, pp. 55 – 62. National Air Filtration Association, Air Filter Test Standards, Air Conditioning, Heating & Refrigeration News, Aug. 11, 1997, pp. 5 – 6. Obler, H., Humidif cation Alternatives for Air Conditioning, Heating / Piping / Air Conditioning, December 1982, pp. 73 – 77. Osborne, W. C., Fans, 2d ed., Pergamon Press, Oxford, England, 1977. Remiarz, R. J., Johnson, B. R., and Agarwal, J. K., Filter Testing with Submicrometer Aerosols, ASHRAE Transactions, 1988, Part II, pp. 1850 – 1858. Rivers, R. D., Interpretation and Use of Air Filter Particle-Size-Eff ciency Data for General Ventilation Applications, ASHRAE Transactions, 1988, Part II, pp. 1835 – 1849. Rosenbaum, D. P., Obtaining Proper Fan Performance from Fans Installed in Air Handlers, ASHRAE Transactions, 1983, Part I B, pp. 790 – 794. Sheringer, J. S., and Govan, F., How to Provide Freeze-up Protection for Heating and Cooling Coils, Heating / Piping / Air Conditioning, no. 2, 1985, pp. 75 – 84. Symonds, S., Internal Coil Care: What’s the Best Solution? Air Conditioning, Heating and Refrigeration News, Aug. 21, 1997, pp. 9 – 10. Tao, W., and Chyi, D. P., Coil Design and Selection, Heating / Piping / Air Conditioning, no. 12, 1985, pp. 66 – 73. The Trane Company, Fans and Their Application in Air Conditioning, Application Engineering Seminar, Trane Company, La Crosse, WI, 1971. The Trane Company, Variax Fans, La Crosse, WI, 1983. Thornburg, D., Filter Testing and IAQ Control Move Forward, HPAC, no. 10, 1999, pp.54 – 56. Trent, W., and Trent, C., Condensate Control, Engineered Systems, no. 7, 1997a, pp. 40 – 44. Trent, W., and Trent, C., Seal It with More Than a Kiss, Engineered Systems, no. 9, 1997b, pp. 52 – 58. Wang, S. K., Air Conditioning, vol. 3, Hong Kong Polytechnic, Hong Kong, 1987. Wood, R. A., Sheff eld, J. W., and Sauer, Jr., H. J., Thermal Contact Conductance of Finned Tubes: The Effect of Various Parameters, ASHRAE Transactions, 1987, Part II, pp. 798 – 810.

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AIR SYSTEMS: EQUIPMENT — AIR-HANDLING UNITS AND PACKAGED UNITS 16.1 AIR-HANDLING UNITS 16.1 Functions of Air-Handling Units 16.1 Classifications of Air-Handling Units 16.2 16.2 MAIN COMPONENTS 16.4 Casing 16.4 Fans 16.4 Coils 16.5 Filters 16.5 Humidifiers 16.5 Outdoor Air Intake, Mixing, and Exhaust Section 16.6 Controls 16.6 Component Layout 16.6 Coil Face Velocity 16.8 16.3 SELECTION OF AIR-HANDLING UNITS 16.9 16.4 PACKAGED UNITS 16.12 Types of Packaged Unit 16.12

Rooftop Packaged Units 16.12 Indoor Packaged Units 16.15 Split Packaged Units 16.16 16.5 PERFORMANCE AND SELECTION OF PACKAGED UNITS 16.17 Indoor Environmental Control 16.17 Indoor Air Quality 16.18 Scroll Compressors and Evaporative Condensers 16.18 Controls 16.18 Minimum Performance 16.19 Selection of Packaged Units 16.19 16.6 FAN ROOM 16.24 Types of Fan Room 16.24 Layout Considerations 16.25 REFERENCES 16.28

16.1 AIR-HANDLING UNITS Functions of Air-Handling Units An air-handling unit (AHU) is the primary equipment in an air system of a central hydronic system; it handles and conditions the air and distrib utes it to v arious conditioned spaces. In an AHU, the required amounts of outdoor air and recirculating air are often mix ed and conditioned. The temperature of the dischar ge air is then maintained within predetermined limits by means of control systems. After that, the conditioned supply air is pro vided with moti ve force and is distrib uted to various conditioned spaces through ductwork and space diffusion devices. Many air -handling units are modular so that the y ha ve the fl xibility to add components as required. An AHU basically consists of an outdoor air intak e and mixing box section, a fan section including a supply fan and a fan motor, a coil section with a w ater cooling coil, a filter section and a control section. A return or relief f an, a heating coil, a precooling coil, and a humidifier may als be included depending on the application. Supply v olume fl w rates of AHUs vary from 2000 to 63,000 cfm (945 to 29,730 L/s). Whether a return f an or a relief f an should be added to an air system depends on the construction and operating characteristics of the air system and the total pressure loss of the return system (see Chap. 22). A heating coil is mainly used in the air -handling unit that serves the perimeter zone, or for morning w arm-up in the heating season. The use of a precooling coil, to dra w 16.1

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cooling water from the cooling to wer as a w ater economizer is discussed in Chap. 21. Humidifier are employed for processing air conditioning and health care f acilites where space humidity must be controlled.

Classifications of Air-Handling Units Air-handling units may be classified according to their structure location, and conditioning characteristics. Horizontal or Vertical Unit. In a horizontal unit, the supply fan, coils, and filters are all installe at the same le vel, as shown in Fig. 16.1 a. Horizontal units need more floor space for installation and they are mainly used as lar ge AHUs. Most horizontal units are installed inside the f an room. Occasionally, small horizontal units may be hung from the ceiling inside the ceiling plenum. In such a circumstance, fan noise and vibration must be carefully controlled if the unit is adjacent to the conditioned space. In a vertical unit, the supply fan is not installed at the same level as the coils and filters ut is often at a higher level, as shown in Fig. 16.1b. Vertical units require less floor space. They are usually smaller, so that the height of the coil section plus the f an section, and the height of the ductw ork that crosses over the AHU under the ceiling, is less than the head room (the height from the floor t the ceiling or the beam of the f an room). The fan room is the room used to house AHUs and other mechanical equipment. Draw-Through Unit or Blo w-Through Unit. In a dra w-through unit, the supply f an is located downstream from the cooling coil section, and the air is drawn through the coil section, as shown in Fig. 16.1a and b. In a draw-through unit, conditioned air is evenly distributed over the entire surface of the coil section. Also the discharge air from the AHU can be easily connected to a supply duct of similar higher velocity. Draw-through units are the most widely used AHUs. In a blo w-through unit, the supply f an is located upstream from the coil section, and the air blows through the coil section, as sho wn in Fig. 16.1 c. Usually , a multizone air -handling unit adopts a blo w-through unit. In a multizone AHU, the coil section is di vided into the hot deck and the cold deck. The heating coil is installed in the hot deck just abo ve the cold deck, where the cooling coil is located. The hot deck is connected to ductw ork that supplies w arm air to the perimeter zone through the warm duct. The cold deck is connected to a cold duct that supplies cold air to both the perimeter and interior zones. A blow-through unit also has the adv antage of treating the supply f an heat gain as part of the coil load and thus reduces the supply system heat gain. Outdoor Air (or Makeup Air) AHU or Mixing AHU. Most mixing AHUs can be used to condition either outdoor air only or a mixture of outdoor air and recirculating air, whereas an outdoor air AHU is used only to condition 100 percent outdoor air , as shown in Fig. 16.1d. An outdoor air, or makeup air, AHU is a once-through unit; there is no return air and mixing box. It may be a constant-v olume system or a variable-air-volume (VAV) system if the number of occupants v aries. In an outdoor-air AHU, the cooling coil is usually a six- to eight-ro w depth coil because of the greater enthalp y difference during cooling and dehumidification in summe . Freeze protections for w ater coils are necessary in locations where the outdoor temperature may be below 32°F (0°C) in winter. A heat recovery coil or a water economizer precooling coil is often installed in makeup air AHUs for energy savings.

SH__ ST__ LG__ DF

Single-Zone AHU or Multizone AHU. A single-zone AHU serves only a single zone. A multizone AHU serves two or more zones, as shown in Fig. 16.1c. A zone can be a large perimeter or an interior zone or one of the man y control zones which connect to a multizone AHU through ducts and terminals. A multizone AHU with a hot and cold deck is no w often used for a dual-duct VAV system (see Chap. 21). Another kind of multizone unit which has man y separate w arm air ducts and cold air

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FIGURE 16.1 Type of air -handling units (AHUs): (a) horizontal, draw-through unit; ( b) vertical draw-through unit; (c) blow-through unit, multizone AHU; (d ) makeup air AHU, custom-built, rooftop unit.

16.3

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ducts with associated w arm and cold air dampers for each of the zones became obsolete because this kind of multizone unit wastes energy, needs complicated control, and is expensive. Factory-Fabricated AHU or F ield-Built AHU, Custom-Built or Standard F abrication. One important reason to use factory-fabricated AHUs or standard fabrications is their lower cost and higher quality. Factory labor and controlled manuf acturing techniques pro vide more ef ficient and bette quality construction than field labor and assembl . Custom-built and field- uilt AHUs pro vide more fl xibility in structure, system component arrangements, dimensions, and specialized functions than standard f abricated products. Custombuilt and field- uilt AHUs also need more comprehensi ve, detailed specifications. Standard abricating products are usually less expensive and can be delivered in a shorter time. Rooftop AHU or Indoor AHU. A rooftop AHU is an outdoor penthouse, as shown in Fig. 16.1 d. It is usually curb-mounted on the roof and should be completely weathertight. The outside casing is usually made of hea vy-gauge galvanized steel or aluminum sheets with corrosion-resistant coating and sealant at the joints, both inside and outside. The fan motor, water valves, damper actuator linkages, and controls are all installed inside the casing. Access doors are necessary for service and maintenance of f ans, coils, and filters. An indoor AHU is usually located in the f an room. Small AHUs are sometimes ceiling-hung.

16.2 MAIN COMPONENTS Because of the impact of the indoor air quality (IA Q), the design and the construction of the AHU have been affected in many ways, as discussed in Gill (1996).

Casing Two kinds of casings are more widely used for new AHUs today: (1) a double-wall sheet-metal casing in which the insulation material is sandwiched between tw o sheet-metal panels of 1- to 2-in. (25- to 50-mm) thickness with a U value from 0.12 to 0.25 Btu / hft2 °F (0.68 to 1.42 W / m2 °C) and (2) single sheet-metal panel with inner insulation layer and perforated metal liners. Although insulating materials such as glass fibers and mineral ool are inert, when they become wet and collect dirt, both the glass fiber and the glass fiber liner p vide the site and source of microbial growth. In addition, glass fiber liner is susceptible to deterioration and erosion ver time. With a double-w all sheet-metal casing, glass fibers are not xposed to the moisture of the ambient air. Its inner surf ace can also be cleaned easily . Perforated metal liners cannot isolate the isulating material from the ambient moisture, but they are helpful to attenuate the fan noise. The outside surf ace of the casing is often coated with an ultra violet-resistant epoxy paint. The interior surface is better coated with a light color paint which increases the ability to spot the debris and microbial gro wth. Hinged access panels to the f an, coils, and filter sections must be pr vided for inspection and maintenance. Thermal break construction employs a resin bridge between the exterior and interior panels to interrupt the through-metal path heat transfer . A well-sealed doublewall metal panel should ha ve an air leakage of only 1 cfm / ft2 (5.078 L / sm2) panel at a 4 in. WC (1000 Pa) of outer and inner static pressure difference.

Fans SH__ ST__ LG__ DF

A double-inlet airfoil, backward-inclined centrifugal f an is often used in lar ge AHUs with greater cfm (L / s) and higher f an total pressure for its higher ef ficien y and lo wer noise. Vane-axial fans with carefully designed sound-absorpti ve housings, sound attenuators at inlet and outlet, and other

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AIR SYSTEMS: EQUIPMENT—AIR-HANDLING UNITS AND PACKAGED UNITS

16.5

attenuations can no w pro vide a sound rating of NC 55 in the f an room. Although the forw ardcurved centrifugal fan has a lower efficien y at full load, it is more compact and its part-load operating characteristics are better than those of a backw ard-inclined centrifugal fan. It is often used in small AHUs and where cfm (L / s) and f an total pressure are lo wer. For VAV systems, a dedicated outdoor air injection f an is sometimes used to pro vide outdoor ventilation air according to demand at both full and part load. An axial relief fan or an unhoused centrifugal return f an may be added as an optional system component. A return f an is used when the total pressure loss of the return system is considerable. This is discussed in Chap. 22. Large fans are usually belt-driven. Only small fans are sometimes direct-driven. As mentioned in Sec. 15.4, an adjustable-frequency variable-speed drive saves more energy than inlet vanes for VAV systems during part-load operation. It is often cost-ef fective for large centrifugal fans although a variable-speed drive is expensive. Inlet vanes are not suitable for small airfoil or backward-inclined centrifugal fans because they block the air passage at the fan inlet. Generally, a centrifugal f an has a higher ef ficien y and at the same time a lo wer noise. Gi ven two fans of the same model, both with the same cfm (L / s) and fan total pressure, a centrifugal fan that is greater in size and slower in speed creates less noise.

__RH TX

Coils In AHUs, the following types of coil are often used: water cooling coils, water heating coils, electric heating coils, and water precooling coils. The construction and characteristics of these coils are discussed in Chap. 15. Electric heating coils are made with nick el-chromium wire as the heating element (see Sec. 8.4). Ceramic b ushes float the heating elements and vertical brackets prevent the elements from sagging. In a finned tu ular element sheathed construction, the electric heating coil is usually made with a spiral fin brazed to a steel sheath. An electric heating coil is usually di vided into several stages for capacity control. When an electric heating coil is installed inside an AHU, the manuf acturer should ha ve the assembly tested by the Underwriters’ Laboratory (UL) to ensure that its requirements are met. Otherwise, the heater may only be installed outside the AHU as a duct heater, with a minimum distance of 4 ft (1.2 m) from the AHU. Various safety cutoffs and controls must be provided according to the National Electrical Codes and other related codes. Cooling and heating coils need periodic cleaning and freeze-up protection in locations where the outdoor air temperature may drop belo w 32°F (0°C) in winter . Condensate pan and condensate drain line must be properly designed and installed. All these are discussed in Chap. 15.

Filters Air filtration is an important component to achi ve an acceptable indoor air quality . In AHUs, earlier low-efficien y filters of the panel type are g ving way to the medium- and high-ef ficien y bag type and cartridge type of filters. Carbon-act vated gaseous absorption filters are also used t remove objectionable odors or v olatile organic compounds (VOCs) in b uildings. Newly developed air filters are more e ficient at rem ving air contaminants of particle size between 0.3 and 5 m which are lung-damaging dust.

Humidifiers Usually, there is no humidifier installed in the AHU for comfort air conditioning systems; b ut the outdoor climate is v ery cold in winter so that if a humidifier is not empl yed, the winter indoor relative humidity may be too low. Humidifiers are necessary for health care acilities and processing systems in pharmaceutical, semiconductor, textile, communication centers, and computer rooms. Steam grid or electric heating element humidifiers are widely used in AHUs where a w arm air

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CHAPTER SIXTEEN

TX

supply and humidity control are needed in winter . Ultrasonic humidifiers are often used for uildings in which a cold air supply and humidity control are required. F or industrial applications such as textile mills where humidity control, air washing, and cold air supply are needed all year round, an air washer is often used for these purposes.

Outdoor Air Intake, Mixing, and Exhaust Section An outdoor air intake, mixing, and exhaust section includes an outdoor air intake, an exhaust outlet, dampers, a mixing box, and a return f an or a relief f an. The location of outdoor air intak e has a direct impact on space IAQ. ●

●

●

●

The outdoor air intak e for each AHU should install with wind shield and louv ers to pre vent rain and birds. If the AHU is located on the roof, the bottom of the outdoor intake louvers should be at least 3 ft above the the roof. The outdoor air intak e must be located as f ar away from the e xhaust outlets and plumbing v ent stacks (horizontally and v ertically) as possible, to pre vent the intak e of e xhaust contaminants, condensation, and freezing which may pro vide a means of gro wth of microorganisms. Codes and local regulations should be followed. The outdoor air intake should reflect the influence of the p vailing winds. An outdoor intak e system should be pro vided with air filters or ven air cleaners in locations where outdoor air contaminants e xceed the National Primary Air Quality Standard, as discussed in Sec. 4.10.

For better outdoor v entilation air control, an outdoor damper should split into tw o dampers: a minimum outdoor v entilation damper and an economizer damper of 100 percent outdoor air free cooling except in small AHUs. Both should pro vide a short ducted outside path for air balancing and install with airfl w measuring station, or a minimum outdoor v entilation air injection f an and control if necessary. Poor mixing, such as parallel outdoor and recirculating airstreams in the mixing box, causes stratification of the mixture as discussed in Sec. 15.11. This is one of the important causes of coil freezing in locations where the outdoor air temperature drops belo w 32°F (0°C). F or good mixing, airstreams should meet at a 90° angle or opposite each other, as shown in Fig. 16.2. Recently, there is a trend to use an unhoused plug / plenum fan as a return f an in man y AHUs and PUs. Usually, a return fan is located nearly in the center of an e xhaust compartment, as shown in Fig. 16.3. The exhaust and recirculating dampers form tw o sides of the e xhaust compartment. A plug / plenum fan has the advantage of discharging from both exhaust and recirculating dampers and is quieter. After the recirculating damper, the mixture enters the coil section.

Controls As discussed in Sec. 5.14, AHU controls include generic, specific safety, and diagnostic controls. The economizer control, discharge air temperature control, duct static pressure control, outdoor ventilation air control, humidity, and w arm-up or cool-do wn controls are discussed in Chap. 21, 22, and 23.

SH__ ST__ LG__ DF

Component Layout In a typical horizontal, drawn-through AHU, the layout of the components in serial order is usually as follows:

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16.7

__RH TX

FIGURE 16.2 Mixing of outdoor and recirculating airstreams: (a) parallel airstreams (poor mixing); (b) airstreams at 90° (good mixing); (c) opposite airstreams (good mixing).

1. 2. 3. 4. 5. 6. 7. 8. 9. 10.

Return or relief fan, exhaust air passage and damper (optional) Mixing box with outdoor air and recirculating air dampers Filters: (prefilte , optional) medium-efficien y filter Preheating coil (optional) Precooling coil (optional) Cooling coil Heating coil (optional) Supply fan Humidifier (optional) High- or ultrahigh-efficien y filters (optional)

If there is a relief f an, it should be located on one side of the e xhausting compartment or in the relief or exhaust passage. If there is an unhoused return fan, it should be located nearly in the center

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CHAPTER SIXTEEN

TX

High-efficiency filters

Supply air

Supply fan

Exhaust

Mediumefficiency filters

Coils

air Exhaust compartment

Return air Unhoused return air fan

Outdoor air

FIGURE 16.3 A typical AHU with unhoused plug/plenum return fan.

of the e xhaust compartment. The v olume fl w and f an total pressure of the return f an should be carefully determined so that the pressure inside the e xhaust compartment is positi ve; and the pressure inside the mixing box must be ne gative in order to e xtract outdoor air , except an outdoor air projected fan is installed in the outdoor air passage (see Chaps. 21 and 23). In an AHU, to condition a mixture of outdoor and recirculating air is often simpler and less expensive than to condition the outdoor air and recirculating air separately . Therefore, the mixing box is usually located before the filters and the coils. A preheating coil is al ways located before the water heating and cooling coils for the sak e of freeze-up protection. A precooling coil is al ways located before a cooling coil for a greater temperature dif ference between the air and w ater. A steam grid humidifier is usually located after a heating coil because humidification is more fective at a higher air temperature. If there are ultrahigh-efficien y filters they should be located as near to the clean room or clean space as possible to pre vent pollution from the ductw ork or elsewhere. Low- or medium-efficien y prefilters must be installed ahead of the coils.

Coil Face Velocity

SH__ ST__ LG__ DF

The size, or more accurately , the width and height of a horizontal AHU, is mainly determined by the face velocity of the coil. A higher coil f ace velocity results in a smaller coil, a higher heattransfer coefficient a greater pressure drop across the coil and filte , and a smaller fan room, which directly affects the space required. On the contrary , a lower coil f ace velocity has a lar ger coil, a lower heat-transfer coefficient and a smaller pressure drop and fan energy use. The maximum f ace v elocity is usually determined according to the v alue required to pre vent carryover of water droplets due to condensate from a cooling coil. As mentioned in Secs. 10.2 and

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16.9

15.10, the cooling coil face velocity should not exceed 500 fpm (2.5 m / s) for a smooth f n coil and 550 fpm (2.75 m / s) for corrugated f ns. The lower limit of the coil face velocity depends mainly on the initial and ener gy cost analysis, including cost of the AHU, fan room size, the total number of annual operating hours, and the unit rate of electric po wer. A lower limit of face velocities between 400 and 450 fpm (2 and 2.25 m / s) at design conditions may be considered appropriate under man y specif c circumstances. For a fan room of adequate head room, it is preferable to use a high coil to reduce the f ace velocity and the pressure drop of coils and f lters. Use of a high coil has little inf uence on the f oor area of the fan room except that when a coil is higher than 42 in. (1070 mm), a cooling coil should split into two coils vertically and use two separate condensate pans vertically to prevent condensate carryover.

__RH TX

16.3 SELECTION OF AIR-HANDLING UNITS Table 16.1 lists general data of the supply fan and coil of a typical horizontal draw-through AHU; fan B means a class II f an. In Table 16.2, the volume f ow – fan static pressure performance of this horizontal draw-through modular AHU (unit size 30) with inlet v anes is presented. A backward-inclined centrifugal fan of 2214-in. (565-mm) diameter is used. In Table 16.2, the following items are also listed: ●

●

●

●

●

Supply volume f ow rate V˙s, in cfm of standard air. Air velocity at fan outlet, fpm. Fan static pressure, in in. WC. Fan total pressure can be obtained by adding the f an velocity pressure to the f an static pressure; v elocity pressure pv, in in. WC, can be calculated by pv  (vout / 4005)2 here vout indicates the fan outlet velocity, in fpm. Revolutions per minute (rpm) of fan impeller. Brake horsepower (bhp) input to fan shaft. The following are recommendations for selection of AHUs from a manufacturer’s catalog:

●

●

●

●

●

●

Draw-through AHUs are widely used. F or a small AHU, a vertical unit sa ves f oor space if the headroom of the f an room is suf f cient except the AHU is ceiling-hung. F or a large AHU, a horizontal unit is often the right choice. The size of the AHU is selected so that the f ace velocity of the cooling coil vcoil is optimum. For corrugated f ns, maximum vcoil should not exceed 550 fpm (2.75 m / s). For lar ge AHUs, choose a backw ard-curved centrifugal f an with airfoil blades or a backw ardinclined fan for higher ef f ciency. Select the rpm of the supply f an or supply and return f ans in order to meet the required system total pressure loss, i.e., external total pressure plus the total pressure loss of the AHU. For VAV systems, an adjustable-frequenc y v ariable-speed dri ve should be compared with inlet vanes via life-c ycle cost analysis. Ener gy consumption also should be analyzed at part-load operation. For a lar ge AHU, an adjustable-frequency variable-speed drive may be cost-ef fective. When an AHU is equipped with a small airfoil or backw ard-inclined centrifugal f an, inlet vanes for capacity control are not recommended because of the e xtremely high air velocity at the fan inlet. The required coil load is met through the v ariation of the number of rows of coil and the f n spacing. An even number of rows are often used so that the inlet and the outlet of the coils are on the same side. F our-row coils are often used for a mixing AHU, and a mak eup AHU seldom uses a coil that exceeds eight rows. In locations where outdoor air temperatures go belo w 32 °F (0 °C), coil freeze-up protection attained by installing a preheating coil and the impro vement of the mixing of airstreams in the mixing box should be considered.

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SH__ ST__ LG__

DF 16.10

39445

Unit size number Description

3

6

8

10

12

14

17

21

25

30

18 41

20

22 14

22 41

Fan data Fan BI

Size

—

—

—

13 12

15 1 15 16

16 21 15 1 16

2 163

2 167

2 11 16

2 11 16

—

—

—

Outlet area (ft )

—

1.41

1.90

2.31

2.79

3.39

4.14

5.05

6.30

6.30

Max. rpm/static pressure (in. WC)

—

—

—

3536 / 8

3183 / 8

2894 / 8

2616/8

2387/8

2164/8

2164/8

Motor hp range

—

—

—

1– 10

1– 15

1– 15

1– 20

1– 25

1– 25

1– 30

2

Coil data Unit coils ( 21 -in. tube) Cooling

Ft2

Heating

5.86

7.54

9.64

12.3

14.2

16.8

20.8

24.4

29.0

21  23

23  36

27  40

27  51

32  55

35  59

37  65

45  67

51  69

51  82

Ft2 Dimensions (in.)

2.34 15  23

4.31 17  36

5.49 20  40

7.01 20  51

9.46 25  55

10.2 25  59

12.3 27  65

15.0 32  67

17.8 37  69

21.2 37  82

2.75 18  22

4.38 18  35

6.50 24  39

8.33 24  50

11.3 30  54

12.1 30  58

14.7 33  64

16.5 218  66*

19.8 118  68* 124  68

23.6 118  81* 124  81

or 1-in. tube coils Ft2 Dimensions (in.)

JC 7/24/2000

BI: Backward-inclined centrifugal fan *Two coils are used for this unit. Source: The Trane Company. Reprinted with permission.

pg 16.10

5 8 -in.

3.32

Dimensions (in.)

SECOND PASS

Shaft size (in.)

ch16

1 11 16

Wang (MCGHP)

TABLE 16.1 General Data on Supply Fans and Coils of a Typical Horizontal Draw-through AHU

39445

TABLE 16.2 Volume Flow and Fan Static Pressure of a Typical Horizontal Draw-through Modular AHU (Size 30) with Inlet Vanes (BackwardInclined Centrifugal Fan of 22.25-in. Diameter) Fan static pressure, in. WC 3.75

4.00

4.25

4.50

4.75

5.00

rpm

bhp

rpm

bhp

rpm

bhp

rpm

bhp

rpm

bhp

rpm

bhp

rpm

bhp

1,524

10.39

1,561

10.91

1,597

11.40

1,632

11.87

1,667

12.32

1,701

12.77

1,734

13.23

1,746

1,570

11.81

1,607

12.54

1,643

13.23

1,678

13.86

1,712

14.44

1,745

14.99

1,777

15.51

12,000

1,905

1,620

13.21

1,656

14.00

1,691

14.81

1,725

15.62

1,759

16.41

1,792

17.16

1,824

17.85

13,000

2,063

1,678

14.90

1,710

15.67

1,742

16.46

1,774

17.28

1,806

18.14

1,838

19.02

1,870

19.90

14,000

2,222

1,741

16.82

1,771

17.61

1,800

18.41

1,830

19.23

1,860

20.08

1,889

20.94

1,919

21.84

15,000

2,381

1,809

18.98

1,836

19.80

1,864

20.63

1,891

21.47

1,919

22.32

1,946

23.20

1,974

24.09

16,000

2,540

1,880

21.36

1,906

22.23

1,931

23.10

1,957

23.97

1,982

24.85

2,008

25.75

2,034

26.65

17,000

2,698

1,953

23.97

1,978

24.88

2,002

25.80

2,026

26.72

2,050

27.64

2,074

28.57

2,098

29.71

18,000

2,857

2,028

26.83

2,052

27.78

2,075

28.74

2,098

29.71

19,000

3,016

2,104

SECOND PASS

1,587

11,000

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Wang (MCGHP)

CFM Outlet std. air velocity

3.50

29.96 5.25

5.50

5.75

6.00

6.25

6.50

bhp

rpm

bhp

rpm

bhp

rpm

bhp

rpm

bhp

rpm

bhp

10,000

1,587

1,766

13.70

1,798

14.19

1,830

14.71

1,861

15.26

1,892

15.82

1,923

16.41

11,000

1,746

1,809

16.01

1,840

16.51

1,871

17.00

1,901

17.51

1,931

18.02

1,960

18.55

12,000

1,905

1,854

18.48

1,885

19.09

1,915

19.68

1,944

20.24

1,973

20.79

2,002

21.33

13,000

2,063

1,901

20.76

1,932

21.58

1,962

22.35

1,991

23.07

2,020

23.76

2,048

24.41

14,000

2,222

1,949

22.77

1,979

23.71

2,008

24.66

2,037

25.60

2,066

26.50

2,094

27.36

15,000

2,381

2,002

25.00

2,030

25.94

2,057

26.90

2,085

27.9

2,113

28.90

2,140

29.92

16,000

2,540

2,060

27.57

2,086

28.51

2,112

29.46

7/24/2000

rpm

JC

Outlet velocity

pg 16.11

CFM std. air

Outlet vel–outlet velocity, fpm rpm–revolutions/minute bhp–brake horsepower Source: The Trane Company. Reprinted with permission.

16.11

__RH

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CHAPTER SIXTEEN ●

●

●

To improve indoor air quality , medium- and high-ef f ciency f lters should be used to pro vide an acceptable IAQ and to protect coils and air distrib ution devices. Dirty coils and condensate pans signif cantly degrade the IAQ. A pref lter must be installed to e xtend the service life of high- and ultrahigh-eff ciency f lters, or gas absorbers. Use an air or w ater economizer to sa ve ener gy. Use indirect and direct e vaporative coolers to replace part of the refrigeration if the y are applicable and economical. Ev aporative cooling is discussed in Chap. 27. Select an AHU with adequate specif c and safety controls and diagnostics.

16.4 PACKAGED UNITS Types of Packaged Unit A packaged unit (PU) is a unitary, self-contained air conditioner. It is also the primary equipment of a unitary packaged system. A packaged unit not only conditions the air and pro vides the moti ve force to supply the conditioned air to the space, but also pro vides gas heating, or electric heating, and refrigeration cooling from its o wn gas- f red furnace and refrigerating equipment or from its own heat pump. It is actually the primary equipment in an air conditioning system. A packaged unit is al ways equipped with DX coil(s) for cooling. This characteristic is the primary dif ference between a packaged unit and an air -handling unit. The portion that handles conditioned air in a packaged unit is called an airhandler , the air system of the packaged system, to distinguish it from an air-handling unit, as shown in Fig. 16.4. Another portion is a condensing unit, the refrigerant plant. Refrigerants HCFC-22, HFC-134a, HFC-404A, HFC-410A, HFC-407A, and HFC-407C are no w used in packaged units. Most PUs are f actory-built standard f abrication units. A packaged unit can be either enclosed in a single package or split into tw o units: an indoor air handler and an outdoor condensing unit. A packaged unit can also be a packaged heat pump; most are air-source heat pumps. In a packaged heat pump, in addition to the f an, DX coil, f lters, compressors, condensers, expansion valves and controls, there are four-way reversing valves to reverse the refrigerant f ow when cooling mode operation is changed to heating mode operation. The construction and size of a packaged unit depend mainly on its model and cooling capacity , in refrigeration tons or Btu / h (W). P ackaged units can be classi f ed according to their place of installation as rooftop packaged units, indoor packaged units, and split packaged units. Among these units, the rooftop packaged units are most widely used in commercial buildings.

Rooftop Packaged Units A rooftop packaged unit is mounted on the roof of the conditioned space, as shown in Fig. 16.4. It is usually enclosed in a weatherproof outer casing. The mixture of outdoor air and recirculating air is often conditioned in the rooftop packaged unit and supplied to the conditioned space on the f oors below. Based on the types of heating and cooling sources, rooftop packaged units can be subdivided into the following: ●

●

SH__ ST__ LG__ DF

●

Gas / electric rooftop packaged unit. In this unit, heating is pro vided by a gas- f red furnace, and cooling is provided by electric-driven reciprocating or scroll compressors. Electric/ electric rooftop packaged unit. In this unit, heating is pro vided by electric heating coils and cooling by reciprocating or scroll compressors. Rooftop packaged heat pump . In this unit, heating and cooling are pro vided by the heat pump, with auxiliary electric heating if necessary.

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AIR SYSTEMS: EQUIPMENT—AIR-HANDLING UNITS AND PACKAGED UNITS

16.13

__RH TX

FIGURE 16.4 permission.)

Cutaway vie w of a typical rooftop packaged unit. (

Source: the Trane Company . Reprinted with

A rooftop packaged unit is a single packaged unit composed of tw o main components: an air handler and a condensing unit. Its cooling capacity may v ary from 3 to 220 tons (10 to 774 kW), and its supply volume f ow rate may vary from 1200 to 80,000 cfm (565 to 37,750 L / s). An air handler of a typical rooftop packaged unit consists of mainly a casing, an indoor f an, DX coils, f lters, a mixing box, and controls; a gas-f red heater, a relief or return fan, and a humidif er are optional. The construction and characteristics of the casing, fans, f lters, and outdoor air intak e and mixing box are similar to those discussed in AHUs. Curb. A rooftop packaged unit is mounted on a curb which is a perimeter frame supporting the unit. A curb is often made of galv anized or aluminized steel sheets or angle iron and sits on a deck. It usually has additional structural support at or beneath the deck. On the top of the curb, there is always a mating f ange that matches the size of the rooftop unit with a sealing gask et to pro vide a weatherproof joint. Curbs are either factory-prefabricated or f eld-assembled. Curbs should be tall enough abo ve the structural support for a sloped roof. The manuf acturer and the structural engineer should be consulted to ensure that the unit can resist the local anticipated wind pressure. The requirements of National Roo f ng Contractors Association (NRCA) should be followed. DX Coils. For a speci f c model and size of rooftop packaged unit, the coil surf ace area is a f xed value. DX coils are usually of tw o, three, and four rows (except makeup units) with a f n spacing of 12 to 17 f ns / in. (1.5- to 2.1-mm f n spacing). F or large units, two separate refrigerant circuits with their own coils, associated expansion valves, and distributors are often used for a better capacity control.

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16.14

Wang (MCGHP)

ch16

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pg 16.14

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The cooling capacity of a rooftop packaged unit should include the additional solar heat absorbed by the unit on the rooftop. Because a DX coil is a wet coil, the same as the w ater cooling coil, condensate drain pans, condensate traps, and the condensate drain line should be properly designed and installed. Since refrigerant is used as the coolant instead of chilled w ater, coil freeze-up protection is no longer necessary. Supply, Return, and Relief or Exhaust Fans. For a rooftop packaged unit with a cooling capacity of 10 tons (35 kW) or less, there is often only a supply (indoor) f an. For a rooftop packaged unit of cooling capacity of 15 to 30 tons (53 to 105 kW), there is often a supply f an and a relief (e xhaust) fan. For a rooftop packaged unit of cooling capacity of 60 tons (210 kW) and greater , some manufacturers offer a supply f an and a return f an. As in AHUs, an unhoused plug / plenum return f an is also often located nearly in the center of an exhaust compartment in a rooftop packaged unit. Supply and return f ans in rooftop packaged units are usually belt-dri ven. For each refrigeration ton (3.5 kW) of cooling capacity , a rooftop packaged unit usually has a nominal supply v olume f ow rate of 350 to 450 cfm / ton (47 to 60 L / s kW). However, a rooftop packaged unit can v ary its speed of supply fan and return fan to provide various volume f ow rates and fan total pressure for a specif c model and size. External pressure is the pressure loss of the duct system and terminals. Most of the rooftop packaged units can v ary their supply v olume f ow rate at a range between 200 and 500 cfm / ton (27 and 67 L / skW). A maximum of 6-in. (1500-P a) fan total static pressure or 4-in. (1000-Pa) external pressure can be pro vided by a supply f an in rooftop packaged units of 30 tons (105 kW) and greater. There are also manuf acturers that offer inlet vanes, inlet cone, or adjustable frequency variablespeed drive for rooftop packaged units of 20 tons (70 kW) and greater to modulate v olume f ow rates of supply and return f ans at part-load operation in v ariable-air-volume systems. F or variablespeed drives, cooling air must be supplied to the elecronic dri ve mechanism to prevent encountered high temperature because of the solar heat on the rooftop. Gas-Fired Furnace and Electric Heating Coil. The gas b urners in a rooftop packaged unit of a heating capacity of 40,000 Btu / h (11,720 W) and greater are power burners of the induced combustion type. A centrifugal blo wer is used to e xtract comb ustion air and comb ustion products to a vent. There are often tw o tub ular heat e xchangers: a 16-gauge (1.6-mm-thick) stainless-steel or aluminum-silicon alloy coated steel primary heat exchanger, and an 18-gauge (1.3-mm-thick) stainless-steel or corrosion-resistant allo y coated steel secondary e xchanger. Both are of free- f oating design to eliminate expansion and contraction stresses and noises. Firing rate is often controlled by modulating the gas v alve through a DDC controller . Safety controls for gas- f red furnace include the pro ving of comb ustion air supply prior to ignition and continuous electronic f ame supervision. An electric heater is often made of internally wired hea vy-duty nickel-chromium elements. It is usually divided into multiple stages or units for capacity control. Humidifiers A humidi f er is optional. Indoor packaged units for computer room and data processing systems are often installed with steam or heating element humidi f ers in a position between the coil section and the supply fan inlet. As in the AHU, the outdoor intak e of a rooftop unit should be shielded from the wind ef fect, located as f ar a way as possible from the contaminated air source, and the bottom of the louv ers should be kept at least 3 ft (0.9 m) from the roof. In a rooftop packaged unit, the condensing unit mainly consists of compressors; an air -cooled, evaporatively cooled, and water-cooled condenser; and controls.

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Compressors. Semihermetic and hermetic reciprocating compressors and scroll compressors are often used. Scroll compressors are more ener gy-eff cient than reciprocating compressors. They need fewer parts and are quieter . Scroll compressors are gradually replacing reciprocating compressors in packaged units. F or medium-size and lar ge rooftop packaged units, two, three, or four

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compressors of equal or sometimes unequal horsepo wer input are preferable for better capacity control in steps.

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Condensers. Multirow, 38-in. (10-mm) copper tube and aluminum f n air-cooled condensing coil connected with subcooling coils are used. Condensing coil may co ver the tw o sides of the condenser or be in a V shape at the middle. The ratio of the face area of the condensing and subcooling coils to the cooling capacity is often 1.5 to 2 ft 2 / ton ref (0.04 to 0.053 m 2 / kWref). Propeller fans are used for the induced-draft condenser f an. The ratio of volume f ow of cooling air to the cooling capacity is between 600 and 900 cfm / tonref (81 and 121 L / s kWref). Air-cooled condensers are most widely used in rooftop packaged units today for their lo wer initial cost and simple operation and maintenance. An evaporative condenser has the adv antage of low condensing temperature, similar to a w atercooled condenser with cooling to wers, as well as lo wer energy consumption b ut at a lo wer initial cost. Today, an e vaporatively cooled condenser has become one of the optional alternati ves in a rooftop packaged unit and is used more and more frequently than before. Heat Pump. A rooftop packaged heat pump is a packaged unit installed with four -way reversing valves to change the refrigerant f ow after the compressor. In an air -source heat pump, the DX coil is often called an indoor coil, and the air -cooled condensing coil the outdoor coil. During cooling mode operation, hot gas from the compressor is f rst discharged to the outdoor coil for condensing and subcooling. The liquid refrigerant then enters the e xpansion valve and indoor coil to produce refrigeration. During the heating mode operation, the reversing valve changes its connections and the direction of refrigerant f ow. Hot gas from the compressor no w enters the indoor coil to release its condensing heat f rst. The liquid refrigerant is then dischar ged to the outdoor coil to absorb heat from the ambient air for e vaporation. Detailed operation and system performance are discussed in Sec. 12. 2. Indoor Packaged Units An indoor packaged unit is also a single packaged, factory-built unit. It is usually installed indoors inside a f an room or a machinery room, as described in Sec. 9.21. A small or medium-size indoor packaged unit may sometimes be f oor-mounted directly inside the conditioned space with or without connected ductw ork, such as the indoor packaged unit in computer rooms, as sho wn in Fig. 16.5. The cooling capacity of the indoor packaged unit may v ary from 3 to 100 tons (10 to 350 kW), and its supply volume f ow rate from 1200 to 40,000 cfm (565 to 18,880 L / s). Indoor packaged units can be classif ed as follows: ●

●

●

Indoor packaged cooling units. Only cooling is provided by the DX coil. Indoor packaged heating / cooling units. These units not only pro vide cooling from the DX coil, but also provide heating from a hot water coil, steam heating coil, or electric heater. Indoor packaged heat pump. When an indoor packaged unit is connected to an air -cooled condenser and equipped with re versing valves, the change of refrigerant f ow also causes the change of cooling mode and heating mode operation and provides heating and cooling as required.

In indoor packaged units, usually only a supply f an is installed in small units. Because of the compact size of the units, generally a forward-curved centrifugal fan is used. For large indoor packaged units, an additional return f an section is added to e xtract recirculating air from the conditioned space through the return duct. For computer room indoor packaged units, return air entering the unit at high le vel and supply air dischar ged at lo w le vel are common. Medium-ef f ciency f lters or sometimes high-ef f ciency f lters are usually used. A steam humidif er or other type of humidi f er is always an integral part of computer room units, so as to maintain a required space relati ve humidity in winter and pre vent static electricity. Multiple semihermetic and hermetic reciprocating compressors or scroll compres-

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FIGURE 16.5 A typical indoor packaged unit.

sors — with dual refrigerant circuits — are used in medium-size and lar ge units for capacity control. A microprocessor-based DDC control is always used for new and retrof t indoor packaged units. An indoor packaged unit differs from a rooftop packaged unit in condensing arrangements. Usually there are two alternatives in indoor packaged units: ●

●

With an air -cooled condenser, hot gas from the compressor is dischar ged to an air -cooled condenser through the dischar ge line located on the rooftop. Liquid refrigerant is returned to the DX coil from the air-cooled condenser through the liquid line. A shell-and-tube or a double-tube water-cooled condenser is installed inside the unit, and the condenser water is supplied from the cooling tower or from other sources.

An economical analysis based on the local conditions should be made to determine whether an air-cooled or a w ater-cooled condenser should be installed. If a w ater-cooled condenser using condenser water from the cooling to wer is used, a precooling coil is sometimes installed in the indoor packaged unit as a component of the water-side economizer.

Split Packaged Units

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A split packaged unit, sometimes called a split system, splits the packaged unit into an indoor air handler and an outdoor condensing unit, which is most probably mounted outdoors, on the rooftop, on a podium, or in some other adjacent place, as shown in Fig. 16.6. The indoor air handler and outdoor condensing unit are connected by refrigerant pipes. An air handler in a split packaged unit is similar to the air handler in a rooftop unit e xcept that a large air handler in a split packaged unit is usually installed inside the f an rooms, whereas the small air handler may be hung under the ceiling. An air handler for split packaged units usually has a cooling capacity from 3 to 80 tons (10 to 280 kW), a supply volume f ow rate of 1200 to 32,000 cfm (565 to 15,100 L / s), and a maximum f an total pressure of 5.0 in. WC (1250 P a) for mediumsize and large units.

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FIGURE 16.6 A typical split packaged unit.

Reciprocating and scroll compressors are used in split packaged units. The condenser in the outdoor condensing unit can be either air -cooled or w ater-cooled. Compared with indoor packaged units, a split packaged unit al ways has its compressors in its outdoor condensing unit, whereas an indoor packaged unit has its compressor, water-cooled condenser indoors. A split packaged heat pump is also a kind of split packaged unit. In such a unit, additional reversing valves or changeo ver arrangements force the refrigerant f ow from the compressor to the outdoor air -cooled condensing coil during the cooling mode operation, and change to the indoor DX coil during the heating mode operation. The cooling capacity of a split packaged heat pump varies from 10 to 30 tons (35 to 105 kW), and the heating capacity ranges from 100,000 to 400,000 Btu / h (29,300 to 117,200 W) at rated conditions.

16.5 PERFORMANCE AND SELECTION OF PACKAGED UNITS A packaged unit is the primary equipment of a packaged air conditioning system. Its function and performance represent the quality and capability of the packaged system. Because of the de velopment of microprocessor-based DDC control systems in the 1980s, and the effort toward energy conservation since the energy crisis in 1973, the difference in indoor environmental control and indoor air quality between a custom-b uilt central system and a packaged system becomes smaller in the late 1990s, according to the following analyses.

Indoor Environmental Control Today, there are so man y types, models, and sizes of packaged units, such as rooftop, indoor, and split units with gas- f red, electric, hot w ater, and steam heating at a v olume f ow from 1200 to 48,000 cfm (565 to 22,650 L /s) and a f an total pressure from 2.0 to 6.0 in. WC (500 to 1500 P a)

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to meet v arious requirments of HV AC&R systems. F actory-fabricated and -assembled packaged units usually ha ve a higher quality than the f eld-assembled central system of nearly the same construction. Many manuf acturers of fer v ariable-air-volume packaged units of cooling capacity of 20 tons (70 kWref) and greater. A VAV packaged unit maintains a required zone temperature by v arying the volume f ow rate of supply air using proportional-inte gral control mode through each terminal. Such a control is a modulation control, offset-free, fast response, and energy-eff cient. Some packaged units pro vide humidi f ers to control the space relati ve humidity as required. Most of the packaged units of fer simultaneously operated air economizer and refrigeration compressor at a cooling capacity of 5 tons (18 kW) and greater. Indoor Air Quality Adequate outdoor v entilation air, higher-eff ciency air f lters, and a clean and ef fective air system directly affect the indoor air quality . A packaged unit is as ef fective at pro viding adequate outdoor ventilation air for the occupants in the conditioned space as an AHU in a central system. With the increase of the f an total pressure (e xternal pressure), most medium-size and lar ge packaged units can now be installed with high-eff ciency f lters.

Scroll Compressors and Evaporative Condensers Many packaged units no w use scroll compressors instead of reciprocating compressors. In 1997, one scroll compressor manuf acturer in the United States announced that the ener gy eff ciency ratio (EER) of its product was 11.5 Btu / h W (3.37 COP). A rooftop packaged unit using an e vaporative condenser can lo wer the condensing temperature, approaching a v alue that only a w ater-cooled condenser can achieve, and therefore is energy-eff cient.

Controls Microprocessor-based speci f c controls, safety controls, and diagnostics for 1997 manuf actured rooftop packaged units with a cooling capacity between 20 to 130 tons (70 to 455 kW ref) are as follows: 1. Discharge air temperature control ● ● ● ●

Cooling: air economizer control Cooling: staging of scroll compressors Heating: modulation of gas valve — continuous f ame supervision Reset based on outdoor or zone temperature

2. Duct static pressure control ● ●

Inlet vanes / variable-speed drive control Maximum duct static pressure control

3. Space pressure control 4. Morning warm-up — full or cycling capacity 5. Minimum ventilation control ●

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Space positive pressurization, exhaust mode, and purge mode

6. DX coil frost protection 7. Occupied / Unoccupied switching — night setback 8. Low ambient compressor lockout

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9. Compressor lead / lag control — a more balanced run time among compressors 10. Humam interf ace — to monitor all temperatures, pressures, humidities, inputs and outputs; to edit set points; and to select services 11. Diagnostics — to detect f aults and diagnostics (a total of 49 dif ferent diagnostics can be read from the display) 12. Emergency stop input

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A satisfactory indoor environment surely will be maintained, and at the same time the rooftop packaged unit is energy-eff cient with all these controls.

Minimum Performance To reduce energy use in the packaged units, ASHRAE / IESNA Standard 90.1-1999 mandates minimum eff ciency requirements of various PUs as listed in Table 11.6, and also specif es the minimum eff ciency requirements for various air source heat pumps as listed in Sec. 12.2. Almost all PUs and air source heat pumps are driven and operated electrically. In Table 11.6, EER indicates the ener gy eff ciency ratio, SEER the seasonal ener gy eff ciency ratio, IPLV the integrated part-load value, and HSPF the heating seasonal performance factor. All of these energy index are def ned in Sec. 9.20. When using the ef f ciency ratings listed in Table 11.6 and Sec. 12.2 to compare dif ferent types of HVAC&R equipment, the following conditions must be considered: ●

●

●

The eff ciency ratings for w ater-cooled equipment cannot be compared directly to those for air cooled equipment. Water-cooled equipment does not include the energy use of required condenser water pumps and cooling to wer fans whereas air-cooled packaged unit ratings include the ener gy use of condenser fan. The ratings for condensing units cannot be directly compared with single or split packaged units. Condensing unit ratings do not include the energy use of fans in indoor air handlers. The eff ciency ratings of a w ater chiller cannot be compared with a packaged unit using DX coil, as the eff ciency of a water chiller does not include the energy use of chilled-water pumps.

Selection of Packaged Units The procedure of selection of PUs is dif ferent from that for an AHU because the number of ro ws and f n spacing of a DX coil are f xed for a specif c model and size of PU. The size of a PU is determined primarily by the required cooling capacity of the DX coil at v arious operating conditions. The cooling capacity of a typical rooftop PU is listed in Table 16.3, and Table 16.4 presents the supply f an performance of this typical rooftop PU. The selection procedure for PUs, based on data from manufacturers’ catalogs, can be outlined as follows: 1. Calculate the cooling coil load and sensible cooling coil load (refer to Chap. 6) of the conditioned space that is serv ed by the PU. F or a unitary packaged system using DX coils, coil load is equal to the refrigeration load of the refrigeration system. Select the model and size of the PU based on the cooling capacity that is equal to or greater than the required coil load and sensible coil load at the speci f ed design conditions. These include the dry- and wet-b ulb temperatures of air entering the DX coil and the outdoor air entering the air cooled condenser or evaporatively cooled condenser. The equipment capacity may e xceed the design load only when the equipment selected is the smallest size needed to meet the load.

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TABLE 16.3 Cooling Capacity of a Typical Rooftop PU, Nominal Capacity 55 Tons; Volume Flow Rate, Standard Air, of 19,250 cfm Ambient temperature, °F 85

95 Entering wet-bulb temperature, °F

67

73

61

67

73

SHR

MBH

SHR

MBH

SHR

MBH

SHR

MBH

SHR

MBH

SHR

75 80 85 90

502 506 530 558

81 95 100 100

559 559 559 564

55 69 83 95

621 620 620 619

33 46 59 71

473 479 505 533

83 97 100 100

528 528 528 536

56 71 85 96

587 586 586 585

33 47 60 73

105°F 61°F

115°F

67°F

73°F

61°F

67°F

73°F

MBH

SHR

MBH

SHR

MBH

SHR

MBH

SHR

MBH

SHR

MBH

SHR

75 80 85 90

443 453 480 508

86 98 100 100

496 495 496 507

57 73 88 98

552 551 551 550

33 47 61 75

413 427 454 481

89 100 100 100

463 462 464 481

58 75 91 100

516 516 515 515

33 48 63 78

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MBH: 1000 Btu/h SHR: Sensible heat ratio Source: The Trane Company. Reprinted with permission.

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TABLE 16.4 Supply Fan Performance of a Typical Rooftop PU, Nominal Capacity 55 Tons Total static pressure, in. WC CFM standard air

rpm

bhp

rpm

bhp

rpm

bhp

rpm

bhp

rpm

bhp

rpm

bhp

rpm

bhp

rpm

bhp

10,000 12,000 14,000 16,000 18,000 20,000 22,000 22,500 24,000

842 843 857 876 900 926 946 950 963

8.09 9.17 10.72 12.59 14.80 17.31 19.85 20.50 22.44

887 887 897 914 936 962 986 991 1,004

9.18 10.31 11.85 13.81 16.08 18.75 21.57 22.28 24.40

927 931 936 952 972 996 1,022 1,027

10.28 11.50 13.01 15.05 17.40 20.14 23.22 23.97

965 972 974 988 1,006 1,029 1,055 1,061

11.37 12.75 14.25 16.31 18.76 21.56 24.79 25.64

1,000 1,012 1,012 1,023 1,040 1,061 1,086 1,093

12.47 14.03 15.54 17.60 20.14 23.01 26.33 27.24

1,034 1,050 1,050 1,057 1,073 1,093 1,116 1,123

13.59 15.34 16.90 18.93 21.54 24.50 27.88 28.84

1,067 1,085 1,087 1,091 1,105 1,124 1,146

14.74 16.65 18.31 20.30 22.96 26.01 29.46

1,099 1,118 1,123 1,124 1,137 1,154

15.90 17.96 19.76 21.74 24.40 27.54

2.500

2.750

3.000

3.250

3.500

3.750

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If the required supply air v olume f ow rate deviates from the nominal rated v alue in percentage, as shown below, then the cooling capacity and sensible cooling capacity can be roughly multiplied by a multiplier because of the change of the heat-transfer coeff cient as follows: Volume f ow Cooling capacity Sensible cooling

 20% 0.965 0.94

 10% 0% 0.985 0.97

10% 1.0 1.015 1.0 1.03 1.06

20% 1.025

2. Calculate the heating coil load at the winter design condition or at the w arm-up period. Determine the capacity of the gas- f red furnace, electric heater , water or steam heating coil. F or a packaged heat pump, calculate the supplementary heating capacity of the gasf red or electric heater. 3. Evaluate the external total pressure loss of the duct system, terminal, and space diffusion devices (see Chaps. 17 and 18). Determine the speed of the supply f an, relief fan, exhaust fan, or return fan, such that the v olume f ow and the f an total pressure of the supply f an, supply fan / return fan combination is equal to or greater than the sum of the e xternal total pressure loss and the total pressure loss in the PU. For fans for which only a f an static pressure is given, roughly a 0.4-in. WC (100-Pa) of velocity pressure can be added to the f an static pressure for purposes of rough estimation of f an total pressure. Check that the face velocity of the DX coil does not exceed 550 fpm (2.75 m / s) so that condensate will not carry over. 4. For small PUs, a medium-eff ciency air f lter of lower pressure drop, such as a f nal pressure drop between 0.4 and 0.6 in. WC (100 to 150 Pa), should be selected. Example 16.1. Select an AHU or a rooftop PU for a typical f oor in a commercial b uilding with the following operating characteristics: Supply volume f ow rate Cooling coil load Sensible cooling coil load Outdoor air temperature Entering coil dry-bulb temperature Entering coil wet-bulb temperature External pressure drop Total pressure loss of AHU or PU

16,000 cfm (7550 L / s) 520,000 Btu / h or 43 tons (152,360 W) 364,000 Btu / h (106,650 W) 95°F (35°C) 80°F (26.7°C) 67°F (19.4°C) 2.0 in. WC (500 Pa) 2.25 in. WC (563 Pa)

For a 4-ro w, 12-f ns/in. w ater-colling coil in an AHU, chilled w ater enters the coil at 45 °F (7.2°C) and is expected to leave the coil at 55°F (12.7°C). Solution: 1. Divide the supply volume f ow rate by 500 fpm, that is, 16,000 / 500  32. From Tables 16.1 and 16.2, select an AHU of size 30, which gives a maximum static pressure up to 8 in. WG and a face area of cooling coil of 29 ft2. The actual face velocity of cooling coil is vcoil 

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16,000 29

 551 fpm (2.75 m / s)

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There will be no carryover of condensate water droplets; therefore, it is acceptable. 2. From Table 16.2, select a size-30 horizontal draw-through unit with a backward-inclined centrifugal fan at an impeller diameter of 2214 in. and a fan speed of 1931 rpm. The fan static pressure is then 4.00 in. WC at a volume f ow of 16,000 cfm. From Table 16.2, because the fan outlet velocity is 2540 fpm, its velocity pressure is then pv 

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 0.40 in. WC  2540 4005  2

And the fan total pressure provided by this fan is 4.00  0.40  4.40 in. WC (1100 Pa) This value is greater than the required fan total pressure 2.0  2.25  4.25 in. WC (1063 Pa) From the manufacturer’s catalog, the heights of the fan and coil modules are both 4 ft 6 in. (1.37 m). For such a height, the headroom of the fan room is usually suff cient. 3. As in Example 15.2, the cooling and dehumidifying capacity per ft 2 of coil f ace area of the water cooling coil given by the manufacturer’s catalog is 19.1 MBtu / h ft2, and the sensible cooling capacity is 13.89 MBtu / h ft2. For a coil f ace area of 29 ft 2, the total cooling and dehumidifying capacity is Qc  19,100  29  553,900 Btu / h (162,290 W) and the sensible cooling capacity is Qcs  13,890  29  402,810 Btu / h (118,020 W) These capacities both are greater than the required cooling capacity of 520,000 Btu / h (152,360 W) and the sensible cooling capacity of 364,000 Btu / h (106,650 W). 4. From Table 16.3, select a rooftop PU with a cooling capacity of 55 tons and a sensible heat ratio SHR c  0.71. When the supply v olume f ow rate of this PU is 19,250 cfm, at an air entering coil dry-bulb temperature of 80 °F and a wet-bulb temperature of 67 °F, and an outdoor air temperature of 95 °F, the cooling capacity is 528 MBtu / h, and the sensible cooling capacity is 0.71  528  375 MBtu / h. Both are greater than the required v alues. The selected 55-ton PU is suitable. From the manufacturer’s catalog, the face area of this 55-ton DX coil is 37.9 ft2. If the supply v olume f ow rate is reduced to 17,625 cfm (19,250  17,625) / 16,000  0.10, or 10 percent greater than the required v olume f ow rate of 16,000 cfm, then the cooling capacity is reduced to 528,000  0.985  520,080 Btu / h (152,380 W) and the sensible cooling capacity will be reduced to 375,000  0.97  363,750 Btu / h (106,580 W) Only the sensible cooling capacity is slightly less than the required value of 364,000 Btu / h. 5. Assume that the supply f an outlet v elocity of the rooftop PU is 2000 fpm (10 m / s); that is, the velocity pressure at fan outlet is pv 

 0.25 in. WC  2000 4005  2

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From Table 16.4, we see that at a supply v olume f ow of 17,625 fpm and a f an speed of about 1135 rpm, this rooftop PU can provide a fan total pressure of pt  4.0  0.25  4.25 in. WC (1060 Pa) This fan total pressure can meet the required value. Check the DX coil’s face velocity 17,625 37.9

 465 fpm (2.32 m / s)

Coil f ace v elocity is less than 550 fpm (2.75 m carryover.

/ s); therefore, there is no condensate droplet

16.6 FAN ROOM Types of Fan Room A fan room is an enclosure in which an AHU, an air handler, an indoor PU, and other accessories and air-handling equipment are located. According to ASHRAE Standard 15-1994, for refrigerating systems of 100 hp (74.6 kW) or less, a fan room may contain refrigeration machinery if ●

●

●

The fan room is a separated, tight construction with tight f tting doors Access by authorized personnel is controlled Detectors (refrigerant, oxygen, etc.) are located in refrigerant leaking areas

Refer to Standard 15-1994 for details. In low-rise buildings of three stories and less, a fan room may be used to serve up to three f oors. In high-rise buildings of four stories and more, a fan room may be used to serve one or more f oors, usually up to 20 f oors, depending on the characteristics of the air system, its initial cost, and its operating cost. Fan rooms can be classif ed according to their pressure characteristics as open or isolated. Open Fan Room. An open fan room is open to the f lter end of the AHU or air handler, as shown in Fig. 16.7. The return ceiling plenum is directly connected to the f an room through an inner-lined return duct or a return duct with a sound attenuator . Outdoor air is often forced to the f an room by an outdoor air f an or a mak eup AHU, or extracted to the f an room by the supply f an in the AHU. The fan room becomes the mixing box of the AHU or air handler . Its pressure is often lo wer than the static pressure in the return plenum and the outdoor atmosphere. Return air is then e xtracted to the fan room through the return duct. An exhaust fan is often installed on the e xternal wall of the fan room to maintain this pressure difference. The advantages of this kind of f an room are a positi ve outdoor air supply and less ductw ork in the fan room. The disadvantages include the following: ●

●

●

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There may be inf ltration of uncontrolled outdoor air if the fan room is not airtight. The fan room is entirely exposed to outdoor air. Suff cient sound attenuation must be pro vided in the transfer duct or in the air passage that transfers the return air from the return ceiling plenum to the fan room.

Isolated Fan Room. In this kind of f an room, shown in Fig. 16.8, the outdoor air , the return air , and the exhaust air are all isolated from the f an room air because of the ductw ork. The static pressure in the f an room depends mainly on the air leakage from or to the AHU or air handler, and the

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16.25

__RH TX

FIGURE 16.7 Open fan room: (a) plan view; (b) sectional view.

air passage connecting the f an room and outside atmophere. This kind of f an room is most widely used in commercial buildings. Fan rooms located at the perimeter of the b uilding often ha ve direct access from the outside walls. They are more con venient to pro vide the outdoor air intak e and e xhaust. F or f an rooms

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TX

FIGURE 16.8 Isolated fan room: (a) plan view; (b) sectional view.

SH__ ST__ LG__ DF

located in the interior core, large outdoor and exhaust air risers are required in multistory b uildings when an air economizer cycle is used. Figure 16.9 shows the plan and sectional vie ws of an isolated f an room for an indoor packaged unit located in the interior core of the b uilding. The unit is also equipped with a coil module for water cooling and heating coils and water economizer precooling coils.

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FIGURE 16.9 Interior core fan room: (a) plan view; (b) sectional view.

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Layout Considerations

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CHAPTER SIXTEEN

A satisfactory fan room layout should meet the following requirements: ●

●

●

●

●

●

It should be compact yet pro vide suf f cient space for the maintenance w orkers to pull out f an shafts, coils, and f lters. Piping connections on the same side of the AHU or PU are preferable. Outdoor intake and exhaust outlets are located either on outside w alls perpendicular to each other or on different walls with a certain distance between them. Fire dampers should be installed to separate the f an room and the f re compartment according to the f re codes. F an room v entilation and e xhaust should be pro vided to meet the requirements of the codes. Inner-lined square elbows or elbows with 2- or 3-in. -(50- or 75-mm) thick duct liners are used for better sound attenuation at lo w frequencies. Sound attenuating de vices are required for both the supply and return sides of the fan. A v ertical AHU occupies less f oor space than a horizontal unit; therefore, it is often the f rst choice if the headroom is suff cient. An unhoused centrifugal f an located in the e xhaust compartment may be the suitable choice of return fan for its lower noise, and it has the advantage of discharging on both sides.

In Fig. 16.8a, at point rf after the return fan, the pressure is positive. At point m in the mixing box, the pressure must be ne gative in order to e xtract outdoor air. There must be a damper and an appropriate pressure drop between point rf and m to guarantee such a positive-to-negative pressure conversion.

REFERENCES

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ASHRAE, ASHRAE / IES Standard 90.1-1989, User’s Manual, ASHRAE Inc., Atlanta, GA, 1992. ASHRAE, Energy Code for Commercial and High-Rise Residential Buildings, Atlanta, 1993. ASHRAE, ASHRAE Handbook 1996, HVAC Systems and Equipment, Atlanta, 1996. ASHRAE / IES Standard 90.1-1999, Energy Standard for Buildings Except New Low Rise Residential Buildings, Atlanta, 1999. Bierwirth, H. C., Packaged Heat Pump Primer, Heating / Piping / Air Conditioning, July 1982, pp. 55 – 59. Brasch, J. F., Electric Duct Heater Principles, Heating / Piping / Air Conditioning, March 1984, pp. 115 – 130. Carrier Corporation, Products and Systems 1992 / 1993 Master Catalog, Carrier Corp., Syracuse, NY, 1992 / 1993. Gill, K. E., IAQ and Air Handling Unit Design, HPAC, no. 1, 1996, pp. 49 – 54. Gill, K. E., Rooftop HVAC, HPAC, no. 7, 1997, pp. 51 – 55. Haessig, D. L., A Solution for DX VAV Air Handlers, Heating / Piping / Air Conditioning, no. 5, 1995, pp. 83 – 86. Haines, R. W., Stratif cation, Heating / Piping / Air Conditioning, November 1980, pp. 70 – 71. McGuire, A. B., Custom Built HVAC Units, Heating / Piping / Air Conditioning, January 1987, pp. 115 – 122. Pannkoke, T., Rooftop HVAC for the 90s, Heating / Piping / Air Conditioning, no. 7, 1993 pp. 33 – 42. Riticher, J. J., Low Face Velocity Air Handling Units, Heating / Piping / Air Conditioning, December 1987, pp. 73 – 75. Scolaro, J. F., and Halm, P. E., Application of Combined VAV Air Handlers and DX Cooling HVAC Packages, Heating / Piping / Air Conditioning, July 1986, pp. 71 – 82. The Trane Company, Packaged Rooftop Air Conditioners, The Trane Co., Clarksville, TN, 1997. Waller, B., Economics of Face Velocities in Air Handling Unit Selection, Heating / Piping / Air Conditioning, March 1987, pp. 93 – 94. Wang, S. K., Air Conditioning, vol. 3, Hong Kong Polytechnic, Hong Kong, 1987. Weisgerber, J., Custom Built HVAC Penthouses, Heating / Piping / Air Conditioning, November 1986, pp. 115 – 117.

39445 Wang (MCGHP) Chap_17

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CHAPTER 17

AIR SYSTEMS: AIR DUCT DESIGN 17.1 BASICS OF AIRFLOW IN DUCTS 17.2 Bernoulli Equation 17.2 Steady Flow Energy Equation 17.2 Static Pressure, Velocity Pressure, and Total Pressure 17.3 Stack Effect 17.5 Laminar Flow and Turbulent Flow 17.6 Velocity Distribution 17.7 Equation of Continuity 17.7 17.2 CHARACTERISTICS OF AIRFLOW IN DUCTS 17.8 Types of Air Duct 17.8 Pressure Characteristics of the Airflow 17.8 System Pressure Loss 17.10 Criteria of Fan Energy Use 17.10 17.3 DUCT CONSTRUCTION 17.12 Maximum Pressure Difference 17.12 Material 17.12 Rectangular Ducts 17.13 Round Ducts 17.17 Flat Oval Ducts 17.17 Flexible Ducts 17.18 Fiberglass Ducts 17.18 17.4 DUCT HEAT GAIN, HEAT LOSS, AND DUCT INSULATION 17.19 Temperature Rise or Drop due to Duct Heat Gain or Loss 17.19 Duct Insulation 17.19 Temperature Rise Curves 17.21 17.5 FRICTIONAL LOSSES 17.22 Darcy-Weisbach Equation 17.22 Friction Factor 17.22 Duct Friction Chart 17.24 Roughness and Temperature Corrections 17.25 Circular Equivalents 17.27 17.6 DYNAMIC LOSSES 17.31 Elbows 17.31 Converging and Diverging Tees and Wyes 17.34 Entrances, Exits, Enlargements, and Contractions 17.38 17.7 FLOW RESISTANCE 17.38 Flow Resistances Connected in Series 17.40 Flow Resistances Connected in Parallel 17.41 Flow Resistance of a Y Connection 17.42 Flow Resistance of a Duct System 17.42 17.8 PRINCIPLES AND CONSIDERATIONS IN AIR DUCT DESIGN 17.43

Optimal Air Duct Design 17.43 Design Velocity 17.45 System Balancing 17.46 Critical Path 17.48 Air Leakage 17.48 Shapes and Material of Air Ducts 17.50 Ductwork Installation 17.50 Fire Protection 17.50 17.9 AIR DUCT DESIGN PROCEDURE AND DUCT LAYOUT 17.51 Design Procedure 17.51 Duct System Characteristics 17.52 Duct Layout 17.52 17.10 DUCT SIZING METHODS 17.53 Equal-Friction Method 17.53 Constant-Velocity Method 17.53 Static Regain Method 17.54 T Method 17.55 17.11 DUCT SYSTEMS WITH CERTAIN PRESSURE LOSSES IN BRANCH TAKEOFFS 17.56 Design Characteristics 17.56 Cost Optimization 17.56 Condensing Two Duct Sections 17.59 Local Loss Coefficients for Diverging Tees and Wyes 17.60 Return or Exhaust Duct Systems 17.63 17.12 DUCT SYSTEMS WITH NEGLIGIBLE PRESSURE LOSS AT BRANCH DUCTS 17.66 Supply Duct Systems 17.66 Pressure Characteristics of Airflow in Supply Ducts 17.66 Return or Exhaust Duct Systems 17.71 17.13 REQUIREMENTS OF EXHAUST DUCT SYSTEMS FOR A MINIMUM VELOCITY 17.72 17.14 COMPUTER-AIDED DUCT DESIGN AND DRAFTING 17.72 Drafting 17.72 Schedules and Layering 17.72 Design Interface 17.73 Running Processes 17.73 17.15 DUCT LINER AND DUCT CLEANING 17.74 Duct Liner 17.74 Duct Cleaning 17.74 17.16 PRESSURE AND AIRFLOW MEASUREMENTS 17.75 Equal-Area versus Log Tchebycheff Rule 17.77 REFERENCES 17.78

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17.1 BASICS OF AIRFLOW IN DUCTS Bernoulli Equation The Bernoulli equation relates the mean v elocity v, in ft / s (m / s), the pressure p, in lb f / ft2 absolute (abs.) or psia (P a abs.), and the elevation z, in ft (m), of a frictionless or ideal fluid at steady state When a fluid motion is said to be in steady state the variables of the fluid at a y point along the fluid f w do not v ary with time. Assuming constant density , the Bernoulli equation can be expressed in the following form: p





gz v2   constant 2gc gc

(17.1)

where p  static pressure, lbf / ft2 abs. (Pa abs.)   fluid densit , lbm / ft3 (kg / m3) g  gravitational acceleration, ft / s2 (m / s2) gc  dimensional constant, 32.2 lbm ft / lbf s2 (1) For convenience, lb  lbm (mass). Steady Flow Energy Equation For a real fluid f wing between two cross sections in an air duct, pipe, or conduit, energy loss is inevitable because of the viscosity of the fluid the presence of the mechanical friction, and eddies. The energy used to o vercome these losses is usually transformed to heat ener gy. If we ignore the kinetic energy difference between the value calculated by the mean velocity of the cross section and the value calculated according to the v elocity distribution of the cross section, then the steady fl w energy equation for a unit mass of real fluid is g ven as p1

1

 u 1J 

gz 1 gz 2 p2 v 21 v 22   qJ   u 2J   W 2gc gc r2 2gc gc

(17.2)

where u  internal energy, Btu / lb (J / kg) J  Joule’s equivalent, 778 ft lbf / Btu (1) q  heat supplied, Btu / lb (J / kg) W  work developed, ft lbf / lb (J /s) In Eq. (17.2), subscripts 1 and 2 indicate the cross section 1 and 2, respectively, and p1 and p2 denote the absolute static pressure at cross section 1 and 2. Signs of q and W follow the convention in thermodynamics, i.e., when heat is supplied to the system, q is positive and when heat is released from the system, q is negative. When work is developed by the system, W is positive; and for work input to the system, W is negative. Multiply both sides of Eq. (17.2) by , ignore the dif ference in densities, and rearrange the terms. Then each term has the unit of pressure, in lbf / ft2 abs. (Pa abs.), or p1 

SH__ ST__ LG__ DF

1v 21 2gc



1gz 1 gc

 p2 

2v 22 2gc



2gz 2 gc

 W  J(u 2  u 1  q)

(17.3)

For an air duct or piping work without a fan, compressor, and pump, W  0. Let the pressure loss from viscosity , friction, and eddies between cross sections 1 and 2 be pf  J(u2  u1  q); then each term of Eq. (17.3) can be expressed in the form of pressure

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AIR SYSTEMS: AIR DUCT DESIGN

p1 

1v 21 2gc



1gz 1 gc

 p2 

2v 22 2gc



2gz 2 gc

17.3

(17.4)

 pf

If both sides of Eq. (17.2) are multiplied by gc / g and W  0, then each term of the equation is e xpressed in the form of head, in ft or in. (m) of f uid column. That is, gc p1



g1

gc pf gc p2 v 21 v 22  z1    z2  2g 2g g2 g

(17.5)

Static Pressure, Velocity Pressure, and Total Pressure Pressure is the force per unit area e xerted by a f uid or solid. In an air duct system, a water piping system, or a refrigerant piping system, f uid pressure including air , water, refrigerant pressure at a surface or a level, or inside an enclosed vessel, or pressure difference between two surfaces is often measured under the following conditions: ●

●

●

Fluid pressure is measured related to a datum of absolute vacuum. Such a measured f uid pressure is given as absolute pressure and is often represented by the pressure exerted at the bottom surface of a water column. Fluid pressure is often more con veniently measured related to a datum of atmospheric pressure. Such a measured f uid pressure is gi ven as gauge pressure. The measured gauge pressure that is greater than the atmosperic pressure is expressed as positive gauge pressure or simply gauge pressure. That part of measured gauge pressure which is less than the atmospheric pressure is e xpressed as negative gauge pressure or vacuum. Fluid pressure is measured as a pressure dif ference, pressure drop, or pressure loss between tw o surfaces, two levels, or two cross-sectional surf aces. The involved two measured pressures must be either both gauge pressure or both absolute pressure.

Consider a supply duct system in a multistory b uilding, as shown in Fig. 17.1. In Eq. (17.4), since p1  pat1  p1 and p2  pat2  p2, where p1 and p2 represent the gauge static pressure and pat1 and pat2 the atmospheric pressure added on the f uid at cross sections 1 and 2. The relationship of f uid properties between cross sections 1 and 2 can be expressed as pat1  p1 

1v 21 2gc



1gz 1 gc

 pat2  p2 

2v 22 2gc



2gz 2 gc

 pf

(17.6)

If the air temperature inside the air duct is equal to the ambient air temperature, and if the stack effect because of the dif ference in air densities between the air columns inside the air duct and the ambient air does not exist, then pat1  pat2  (2z 2  1z 1)

g gc

Therefore, Eq. (17.6) becomes p1 

1v 21 2gc

 p2 

2v 22 2gc

 pf

(17.7)

Equation (17.7) is one of the primary equations used to determine the pressure characteristics of an air duct system that does not contain a fan and in which the stack effect is negligible. Static Pressure. In Eq. (17.7), static pressures p1 and p2 are often represented by ps. In air duct systems, its unit can be either Pa (pascal, or newtons per square meter) in SI units, or the height of water

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FIGURE 17.1 Pressure characteristics of an air duct system.

column, in inches, for I-P units. Either is e xpressed in gauge pressure or absolute pressure. The relationship between the static pressure ps, in lbf / ft2, and the height of a water column H, in ft, is ps 

wgHA gc A



wgH gc

(17.8)

where A  cross-sectional area of water column, ft2   density of water, lb / ft3 When the static pressure is expressed as the height of 1 in. of w ater column absolute pressure (1 in. WC) or 1 in. of w ater column gauge pressure (1 in. WG), for a density of w ater w  62.3 lb / ft3 and numerically g  gc  32.2, we f nd from Eq. (17.8) ps  1 in. WG

and

ps  1 in. WC 

wgH gc



62.3 32.2 1  5.192 lbf / ft 2 32.2 12

That is 1 in. WC  5.192 lbf /ft2. Because 1 lb f /ft2  47.88 Pa, 1 in. WC  5.192 47.88  248.6 Pa. Velocity Pressure. In Eq. (17.7), the term v2 / (2gc) is called the v elocity pressure, or dynamic pressure, and is represented by the symbol pv, that is, SH__ ST__ LG__ DF

pv 

v 2 2gc

(17.9)

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AIR SYSTEMS: AIR DUCT DESIGN

17.5

where   air density , lb / ft3. F or air density   0.075 lb / ft3, if the v elocity pressure pv is expressed in in. WC and the air velocity in fpm or ft / min, according to Eq. (17.9),

 60v   2

5.192(2pvgc)





5.192 2 32.2 pv  4458pv 0.075

v  4005 √pv

and

pv 

v  4005 

2

(17.10)

In SI units, if v is expressed in m / s and  in kg / m3, then pv is in P a, and pv can be calculated by Eq. (17.9). For SI units, gc  1. Total Pressure. At any cross-sectional plane perpendicular to the direction of the air f ow, the total pressure of the airstream pt is def ned as the sum of the static pressure ps and the velocity pressure pv, that is, pt  ps  pv

pv  pt  ps

(17.11)

From Eq. (17.11) v elocity pressure pv is also a kind of pressure dif ference. The units of pt must be consistent with ps and pv. In I-P units, it is also indicated in inches WC or WG and in SI units in P a absolute (abs.) or gauge (g). Substituting Eq. (17.11) into Eq. (17.7), we see that pt1  pt2  pf

(17.12)

Equation (17.12) is another primary equation that relates the pressure loss from friction and other sources, pf, and the total pressure pt1 and pt2 at two cross sections of the air duct system.

Stack Effect When an air duct system has an ele vation difference and the air temperature inside the air duct is different from the ambient air temperature, the stack effect exists. It affects airf ow at different elevations. During a hot summer day, when the density of the outdoor air is less than the density of the cold supply air inside the air duct, the pressure e xerted by the atmospheric air column between z1 and z2, as shown in Fig. 17.1, is given as pat1  pat2 

o g(z 2  z 1) gc

where o  mean density of the ambient air , lb / ft3 (kg / m3). And the pressure e xerted by the air column inside the air duct between z1 and z2 is

2gz 2  1gz 1 gc



i g(z 2  z 1) gc

where i  mean density of the supply air inside the air duct, in lb / ft3 (kg / m3). If the dif ferences between the densities inside the air ducts, 1 and i, and 2 and i, are ignored, substituting these relationships into Eq. (17.6) yields p1 

and

1v 21 2gc



pt1 

g(o  i)(z 2  z 1) gc g(o  i)(z 2  z 1) gc

 p2 

2v 22 2gc

 pt2  pf

 pf

(17.13)

(17.14)

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CHAPTER SEVENTEEN

The third term on the left-hand side of Eq. (17.13) and the second term on the left-hand side of Eq. (17.14) are called the stack effect pst, in lbf / ft2 (Pa), pst 

g(o  i)(z 2  z 1) gc

(17.15)

If pst is expressed in in. WC (1 lbf / ft2  0.1926 in. WC), then pst 

0.1926g(o  i)(z 2  z 1) gc

(17.16)

For an upw ard supply duct system, z2 z1. If cold air is supplied, i o and pst is ne gative. If warm air is supplied, o i and pst is positive. For a downward supply duct system with a cold air supply, pst is positive; if there is a warm air supply, pst is negative. When supply air is at a temperature of 60 °F (15.5°C) and a relati ve humidity of 80 percent, its density is 0.075 lb / ft3 (1.205 kg / m3). Also, if the space air has a temperature of 75 °F (24°C) and a relative humidity of 50 percent, its density is 0.073 lb / ft3 (1.171 kg / m3). Numerically, g  gc, for a difference of z2  z1  30 ft (9.14 m), so pst  0.1926 (0.073  0.075)(30)  0.0116 in. WC (2.96 Pa) For an air-handling unit or a packaged unit that supplies air to the same f oor where it is located, the stack effect is usually ignored.

Laminar Flow and Turbulent Flow Reynolds identif ed two types of f uid f ow in 1883 by observing the beha vior of a stream of dye in a water f ow: laminar f ow and turbulent f ow. He also discovered that the ratio of inertial to viscous forces is the criterion that distinguishes these tw o types of f uid f ow. This dimensionless parameter is now widely known as Reynolds number Re, or Re 

vL 



vL v

(17.17)

where   density of f uid, lb / ft3 (kg / m3) v  velocity of f uid, ft / s (m / s) L  characteristic length, ft (m)   viscosity or absolute viscosity, lb / ft  s (N  s / m2) v  kinematic viscosity, ft2 / s (m2 / s) Many experiments have shown that laminar f ow occurs at Re 2000 in round ducts and pipes. A transition region exists between 2000 Re 4000. When Re 4000, the f uid f ow is probably a turbulent f ow. At 60°F (15.5°C), the viscosity   1.21 105 lb / ft s (18.0 106 N s / m2). For a round duct of 1-ft (0.305-m) diameter and an air f ow through it of 3 ft / s (180 ft / min or 0.915 m / s), the Reynolds number is Re  SH__ ST__ LG__ DF

vL 



0.075 3 1  18,595 1.21 105

The Re of such an air duct is f ar greater than 4000. Therefore, the airf ow inside the air duct is usually turbulent except within the boundary layer adjacent to the duct wall.

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AIR SYSTEMS: AIR DUCT DESIGN

17.7

FIGURE 17.2 Velocity distribution in a circular duct.

Velocity Distribution The velocity distributions of turbulent and laminar f ow at a specif c cross section in a circular duct that result from (1) the mechanical friction between the f uid particles and the duct w all and (2) the shearing stress of the viscous f uid are shown in Fig. 17.2. The difference between these tw o types of f ow is signif cant. For a fully de veloped turbulent f ow, the air v elocity v, in fpm (m /s), at various distances from the duct wall of a circular duct y, in ft (m), varies according to Prantdl ’s one-seventh power law as follows: v vmax



 Rv 

1/7

(17.18)

where vmax  maximum air velocity on centerline of air duct, fpm (m / s) R  radius of duct, ft (m) The mean air velocity vm lies at a distance about 0.33R from the duct wall. Equation of Continuity For one-dimensional f uid f ow at steady state, the application of the principle of conserv ation of mass gives the following equation of continuity: m˙  1v1A1  2v2 A2

(17.19)

where m˙  mass f ow rate, lb/s (kg/s) A  cross-sectional area perpendicular to f uid f ow, ft2 (m2) Subscripts 1 and 2 indicate the cross sections 1 and 2 along the f uid f ow. If the dif ferences in f uid density at v arious cross sections are ne gligible, then the equation of continuity becomes V˙  A1v1  A2v2 where V˙  volume f ow rate of airf ow, cfm (m3 / s) v  mean air velocity at any specif c cross section, fpm (m / s)

(17.20) __SH __ST __LG DF

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CHAPTER SEVENTEEN

Theoretically, the velocity pressure pvt should be calculated as pvt 

[ v 2 / (2gc)] vdA v dA

(17.21)

Its value is slightly different from pv calculated from Eq. (17.9), which is based on mean velocity v. For a fully de veloped turbulent f ow, pvt  1.06pv. Since most e xperimental results of pressure loss (indicated in terms of velocity pressure) are calculated by pv  v2 / (2gc), for the sake of simplicity pv is used here instead of pvt.

17.2 CHARACTERISTICS OF AIR FLOW IN DUCTS Types of Air Duct Air ducts can be classif ed into four types according to their transporting functions: 1. Supply duct. Conditioned air is supplied to the conditioned space. 2. Return duct. Space air is returned (1) to the f an room where the air -handling unit is installed or (2) to the packaged unit. 3. Outdoor air duct . Outdoor air is transported to the air -handling unit, to the f an room, or to the space directly. 4. Exhaust duct. Space air or contaminated air is e xhausted from the space, equipment, fan room, or localized area. Each of these four types of duct may also subdi vide into headers, main ducts, and branch ducts or runouts. A header is that part of a duct that connects directly to the supply or e xhaust fan before air is supplied to the main ducts in a lar ge duct system. Main ducts ha ve comparatively greater f ow rates and size, serve a greater conditioned area, and, therefore, allow higher air v elocities. Branch ducts are usually connected to the terminals, hoods, supply outlets, return grilles, and e xhaust hoods. A vertical duct is called a riser. Sometimes, a header or a main duct is also called a trunk. Pressure Characteristics of the Airflow

SH__ ST__ LG__ DF

During the analysis of the pressure characteristics of air f ow in a fan duct system such as the one in Fig. 17.3, it is assumed that the static pressure of the space air is equal to the static pressure of the atmospheric air, and the v elocity pressure of the space air is equal to zero. Also, for convenient measurements and presentation, as pre viously mentioned, the pressure of the atmospheric air is taken as the datum, that is, pat  0, and pressure is expressed as gauge pressure. When p pat, p is positive; and if p pat, then p is negative. In a f an duct system, a fan or f ans are connected to a ductwork and equipment. At cross section R1, as the recirculating air enters the return grille, both the total pressure pt and static pressure ps decrease as the result of the total pressure loss of the inlet. The velocity pressure pv, indicated by the shaded section in Fig. 17.3, will gradually increase until it is equal to the velocity of the branch duct. Both pt and ps are negative so that air will f ow from the conditioned space at a datum of 0 to a ne gative pressure. Because v elocity pressure pv is always positive in the direction of f ow, from Eq. (17.11) pt  ps  pv, so ps is then smaller, or more negative, than pt. When the recirculating air f ows through the branch duct se gment R1-11, elbow 1 1-12, branch duct segment 12-13, diffuser 13-14, and branch duct segment 14-15, both pt and ps drop because of the pressure losses. Velocity pressure pv remains the same between cross sections R1 and 1 3. It gradually decreases because of the diffuser 13-14 and remains the same between 14 and 15. As the recirculating air f ows through the converging tee 15-1, this straight-through stream meets with another branch stream of recirculating air from duct section R2-1, at node or junction 1. The

39445 Wang (MCGHP) Chap_17

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AIR SYSTEMS: AIR DUCT DESIGN

17.9

FIGURE 17.3 Pressure characteristics of a fan duct system.

combined airstream then becomes the main stream. Total pressure pt usually decreases when the straight-through stream f ows through the con verging tee. It may increase if the v elocity of the branch stream is much higher than that of the straight-through stream. Ho wever, the sum of the energies of the straight-through and the branch streams at the upstream side of the con verging tee is always higher than the energy at the downstream side because of the pressure loss of the converging tee. In the main duct section 1-2 1, pt and ps drop further while pv increases because of the higher air velocity. This is mainly because of the greater volume f ow rate in main duct section 1-2 1. When the recirculating air enters the air -handling unit, its velocity drops sharply to a v alue between 400 and 600 fpm (2 and 3 m / s). The air is then mix ed with the outdoor airstream from the fresh air duct at node 2. After the mixing section, pt and ps drop sharply when the mixture of recirculating and outdoor air f ows through the f lter and the coil. Total pressure pt and static pressure ps drop to their minimum values before the inlet of the supply fan Fi. At the supply f an, pt is raised to its highest v alue at the f an outlet Fo. Both pt and ps decrease in duct section Fo-3. At junction 3, the airstream di verges into the main stream or straight-through stream and the branch stream. Although there is a drop in pt after the main stream passing through the diverging tee 3-3 1, ps increases because of the smaller pv after 3 1. The increase in ps due to the decrease of pv is known as the static regain ps,r in in. WC (Pa). It can be expressed as ps,r  

v3

2

v31

2

 4005    4005  (v 23  v 231) 2

(17.22)

where v3, v31  air velocity at cross section 3 and 3 1, respectively, fpm (m / s). In duct sections 3-4 and 4-S3, pt decreases gradually along the direction of air f ow. Finally, pt, ps, and pv all drop to 0 after the supply air is discharged to the conditioned space.

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pg 17.10

CHAPTER SEVENTEEN

The pressure characteristics along the airf ow can be summarized as follows: ●

●

●

●

In most sections, pt of the main airstream decreases along the air f ow. However, pt of the main airstream may increase because of the higher velocity of the combined branch airstream. When air f ows through the supply f an, mechanical work is done on the air so that pt and ps are raised from a minimum ne gative value at the f an inlet to a maximum positi ve value at the f an outlet. The pressure characteristics between an y two cross sections of a duct system are go verned by the change of pt and the pressure loss pf between these tw o cross sections pt1  pt2  pf. Static pressure is always calculated as ps  pt  pv. In a constant-v olume air system, the airf ow inside an air duct is considered steady and continuous. Because the change in ps in a fan duct system is small when compared with pat, the airf ow is also considered incompressible.

System Pressure Loss For an air system (a f an duct system), system pressure loss psy, in in. WC (Pa), is the sum of the total pressure losses of the return air system pr,s, section R1-21 in Fig. 17.3; the air -handling unit pAHU, section 2 1-Fo, or the packaged unit pPU; and the supply air system ps,s, section Fo-S3; all are expressed in in. WC (Pa). That is, psy  pr,s  pAHU  ps,s  pr,s  pPU  ps,s

(17.23)

Both pr,s and ps,s may include the pressure losses of duct sections (including duct se gments); duct f ttings such as elbows, diffusers, and converging and diverging tees; components; and equipment. The sum of pr,s and ps,s is called the e xternal total pressure loss, or e xternal pressure loss pt,ex, as opposed to the pressure loss in the air -handling unit or packaged unit. The external pressure loss pt,ex  pr,s  ps,s, in in. WC (Pa). In commercial buildings, most air systems ha ve a system pressure loss of 2.5 to 6 in. WC (625 to 1500 Pa). Of this, pAHU usually has a value of 1.5 to 3 in. WC (375 to 750 Pa), and ps,s is usually less than 0.6 psy except in large auditoriums and indoor stadiums. Criteria of Fan Energy Use ASHRAE / IESNA Standard 90.1-1999 speci f es that each air system ha ving a total f an po wer exceeding 5 hp (3.7 kW) shall meet the follo wing specif ed allowable fan system power (AFSP) in order to encourage heat recovery, relief fan, and other energy eff cient means: (AFSP)  (fan power limitation)(temperature ratio)  (pressure credit)  (relief fan credit) (17.24) Standard 90.1-1999 speci f es f an po wer limitation (including supply , return, relief, and e xhaust fans) as follows: For constant-volume systems Psy 0.0012 hp / cfm (0.0019 W s / L) V˙s, d SH__ ST__ LG__ DF

Psy 0.0011 hp / cfm (0.00174 W s / L) V˙s, d

when V˙s,d 20,000 cfm when V˙s, d  20,000 cfm

(17.24a)

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pg 17.11

AIR SYSTEMS: AIR DUCT DESIGN

17.11

where Psy  each air system total power input to the fan motors, hp (kW) V˙s,d  volume f ow rate of air system at design conditions, cfm (L / s) For variable-air-volume (VAV) system: nameplate fan motor power is 0.0017 hp / cfm (0.0027 W per L / s); when supply v olume f ow  20,000 cfm (9440 L / s), allowable nameplate f an motor power is 0.0015 hp / cfm (W per L / s). That is: Psy 0.0017 hp / cfm (0.0027 W s / L) V˙s,d

when V˙s,d 20,000 cfm

Psy 0.0015 hp / cfm (0.0024 W s / L) V˙s, d

when V˙s,d  20,000 cfm

(17.24b)

Psy can be calculated as Psy 

psy,dV˙s,d

(17.25a)

6350f,dd,dm,d

where psy,d  system total pressure loss at design conditions, in. WC (Pa) f,d, d,d, m,d  fan total eff ciency, drive eff ciency, and motor eff ciency at design conditions Substituting Eq. (17.25 a) into Eq. (17.24), for constant-volume systems, the system total pressure loss is psy

f,dd,dm,d psy

f,dd,dm,d

7.6 in. WC (Pa)

when V˙s,d 20,000 cfm

7.0 in. WC (Pa)

when V˙s,d  20,000 cfm

(17.25b)

For VAV systems, the system total pressure loss is psy

f,dd,dm,d psy

f,dd,dm,d

10.8 in. WC (Pa)

when V˙s,d 20,000 cfm

9.5 in. WC (Pa)

when V˙s,d  20,000 cfm

(17.25c)

In Eq. (17.24), temperature ratio can be calculated as: Temperature ratio 

Tr, set  Ts,d

(17.26)

20

Pressure credit, in hp (kW), can be calculated as: Pressure credit 

 V˙s,d(pfil  1.0)

n1

3718



 V˙s,dpHR

m1

3718

(17.27a)

Relief fan credit, in hp (kW), can be calculated as:



Relief fan credit  Pr, f 1 

1  V˙r, f V˙s,d



(17.27b)

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pg 17.12

CHAPTER SEVENTEEN

where Tr, set  room set point temperature, °F (°C) Ts,d  design supply air temperature for the zone in which the thermostat is located, °F(°C) pf l, pHR  total pressure loss of the f lters and of the heat recovery coils, in WC (Pa) pr, f  name plate relief fan motor power, hp (kW) V˙r, f  volume f ow rate of relief fan at cooling design operation, cfm (L / s) Consider a VAV system with a design supply v olume f ow rate V˙s,d 20,000 cfm (9440 L / s). If a rooftop packaged system is used with a combined f an, drive, and motor ef f ciency f,dd,dm,d  0.45, from Eq. (17.25c), the allowable system pressure loss for this rooftop package system (including supply and return fans) is psy f,dd,dm,d (9.5) 0.45 9.5 4.3 in. WC (1070 Pa) If the design system pressure loss psy 4.3 in. WC (1070 P a), the HVAC&R system designer is recommended either to use a relief fan instead of a return fan for relief fan credit, or to take into account the pressure credit if f lter’s total pressure loss is greater than 1 in. WC (250 Pa) or if there is heat recovery coil, increase the supply temperature differential Ts  (Tr, set  Ts) to a value greater than 20°F (11.1°C) if they are cost ef fective to meet the f an power limitation in ASHRAE / IESNA Standard 90.1–1999. For a VAV reheat central system of V˙s,d 20,000 cfm (9440 L / s), if combined eff ciency is 55 percent, from Eq. (17.25c), the allowable system pressure loss is psy f,dd,dm,d (9.5) 0.55 9.5 5.2 in. WC (1300 Pa) If the design system total pressure loss 5.2 in. WC, the same as for the rooftop packaged system, means of relief f an credit, pressure credit due to f lters and heat reco very coils, and the increase of the supply temperature dif ferential greater than 20 °F (11.1 °C) should be considered. Refer to ASHRAE Standard 90.1-1999 for details.

17.3 DUCT CONSTRUCTION Maximum Pressure Difference Duct systems can be classi f ed according to the maximum pressure dif ference between the air inside the duct and the ambient air (also called the static pressure dif ferential) as  0.5 in. WC ( 125 Pa),  1 in. WC ( 250 Pa),  2 in. WC ( 500 Pa),  3 in. WC ( 750 Pa),  4 in. WC ( 1000 Pa),  6 in. WC ( 1500 Pa), and  10 in. WC ( 2500 Pa). In actual practice, the maximum pressure difference of the supply or return duct system in commercial buildings is usually less than  3 in. WC ( 750 Pa). In commercial b uildings, a low-pressure duct system has a static pressure dif ferential of 2 in. WC (500 P a) or less, and the maximum air v elocity inside the air duct is usually 2400 fpm (12 m / s). A medium-pressure duct system has a static pressure dif ferential of 2 to 6 in. WC (500 to 1500 Pa) with a maximum air velocity of about 3500 fpm (17.5 m / s). In industrial duct systems, including mechanical v entilation, mechanical e xhaust, and industrial air pollution control systems, the pressure difference is often higher. In residential buildings, the static pressure differential of the duct systems is classif ed as  0.5 in. WC ( 125 Pa) or  1 in. WC ( 250 Pa).

Material SH__ ST__ LG__ DF

Underwriters’ Laboratory (UL) classif es duct systems according to the f ame spread and smoke developed of the duct material during f re as follows:

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pg 17.13

AIR SYSTEMS: AIR DUCT DESIGN

17.13

FIGURE 17.4 Various types of air duct: (a) rectangular duct; ( b) round duct with spiral seam; ( c) f at oval duct; (d) f exible duct.

Class 0. Zero f ame spread, zero smoke developed. Class 1. A f ame spread rating of not more than 25 without e vidence of continued progressi ve combustion and a smoke developed rating of not more than 50. Class 2. A f ame speed of 50 and a smoke developed rating of 100. National Fire Protection Association (NFPA) Standard 90A speci f es that the material of the ducts be iron; steel including galv anized sheets, aluminum, concrete, masonary; or clay tile. Ducts f abricated by these materials are listed as class 0. UL Standard 181 allows class 1 material to be used for ducts when the y do not serv e as risers for more than tw o stories or are not used in temperatures higher than 250 °F (121 °C). Fibrous glass and man y f exible ducts that are f actory-fabricated are approved by UL as class 1. Ducts can be classi f ed according to their shapes into rectangular , round, f at oval, and f exible, as shown in Fig. 17.4. Rectangular Ducts For the space available between the structural beam and the ceiling in a b uilding, rectangular ducts have the greatest cross-sectional area. They are less rigid than round ducts and are more easily

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pg 17.14

CHAPTER SEVENTEEN

TABLE 17.1 Thickness of Galvanized Sheet for Rectangular Ducts Thickness, in. Gauge

Nominal

Minimum

Nominal weight, lb / ft2

30 28 26 24 22 20 18 16 14 13 12 11 10

0.0157 0.0187 0.0217 0.0276 0.0336 0.0396 0.0516 0.0635 0.0785 0.0934 0.1084 0.1233 0.1382

0.0127 0.0157 0.0187 0.0236 0.0296 0.0356 0.0466 0.0575 0.0705 0.0854 0.0994 0.1143 0.1292

0.656 0.781 0.906 1.156 1.406 1.656 2.156 2.656 3.281 3.906 4.531 5.156 5.781

Note: Minimum thickness is based on thickness tolerances of hot-dip galvanized sheets in cut lengths and coils (per ASTM Standard A525). Tolerance is valid for 48- and 60-in.wide sheets. Source: ASHRAE Handbook 1988, Equipment. Reprinted with permission.

SH__ ST__ LG__ DF

fabricated on-site. The joints of rectangular ducts ha ve a comparati vely greater percentage of air leakage than f actory-fabricated spiral-seamed round ducts and f at oval ducts, as well as f berglass ducts. Unsealed rectangular ducts may ha ve an air leakage from 15 to 20 percent of the supply v olume f ow rate. Rectangular ducts are usually used in low-pressure systems. The ratio of the long side a to the short side b in a rectangular duct is called the aspect ratio Ras. The greater Ras, the higher the pressure loss per unit length as well as the heat loss and heat gain per unit volume f ow rate transported. In addition, more labor and material are required. Galvanized sheet or, more precisely, galvanized coated steel sheet, and aluminum sheet are the materials most widely used for rectangular ducts. To prevent vibration of the duct w all by the pulsating airf ow, transverse joints and longitudinal seam reinforcements are required in ferrous metal ducts. The galv anized sheet gauge and thickness for rectangular ducts are listed in Table 17.1. Table 17.2 gives specif cations for rectangular ferrous metal duct construction for commercial systems based on the publication of the Sheet Metal and Air Conditioning Contractors’ National Association (SMACNA) titled HVAC Duct Construction Standards — Metal and Flexible. For design and construction of an economical duct system, it is recommended to select an optimum combination of minimum galv anized sheet thickness, type of transv erse joint reinforcement, and its maximum spacing for a speci f c duct dimension at a speci f c pressure dif ferential between the air inside the duct and the ambient air. For rectangular ducts, one uses the same metal thickness for all sides of the duct and e valuates duct reinforcement on each side separately. In Table 17.2, for a given duct dimension and thickness, letters indicate the type of duct reinforcement (rigidity class) and numbers indicate maximum spacing, in ft. Blanks indicate that reinforcement is not required, and dashes denote that such a combination is not allowed. Transverse joint reinforcements, abridged from SMA CNA’s publication HVAC Duct Construction Standar d — Metal and Fle xible and ASHRAE Handbook 1988, Equipment are presented in Table 17.3. These must be matched with the arrangements in Table 17.2. Duct hangers should be installed at right angles to the centerline of the duct. Habjan (1984) recommended the follo wing maximum duct hanger spacing:

TABLE 17.2 Rectangular Ferrous Metal Duct Construction for Commercial Buildings

Duct dimensions, in.

0.0466 (18)

0.0356 (20)

0.0296 (22)

0.0236 (24)

0.0187 (26)

Pressure, in. WG 2

C-10 C-10 D-10 E-10 E-8 G-8 G-5 H-5 I-5 J-4 K-4 H-2.5 plus rods

3

D-10 D-10 D-10 E-8 E-5 G-5 H-5 H-5 I-4 J-3 L-3 H-2.5 plus rods

1

C-10 C-10 D-10 E-10 E-8 F-8 G-8 H-5 I-5 I-4 H-2.5 plus rods

2

B-10 C-10 C-10 C-10 D-10 E-8 E-5 F-5 G-5 H-5 H-4 J-4 K-3 H-2.5 plus rods

3

B-10 C-10 C-10 D-10 D-10 D-8 E-5 E-5 G-5 H-5 H-4 H-3 J-3 K-2.5 H-2.5 plus rods

1

A-10 B-10 B-10 C-10 C-10 D-10 D-8 E-5 E-5 F-5 G-4 H-4 I-3 —

2

B-10 B-10 C-10 C-10 C-8 D-8 D-5 E-5 F-5 F-4 G-4 H-3 I-3 J-2.5 —

3

A-8 A-8 B-8 B-8 B-5 C-5 C-5 C-5 E-5 E-5 F-4 G-3 G-3 H-3 J-2.5 J-2 —

0.5

A-10 B-10 B-10 C-10 D-8 D-8 D-5 E-5 F-5 H-5 H-4 —

1

A-10 A-10 B-10 B-10 C-10 C-10 D-8 D-5 E-5 E-5 F-5 G-4 — — —

2

3

A-10 A-10 B-8 B-8 C-8 C-8 C-5 C-5 D-5 E-5 E-4 F-3 G-3 H-3 — — —

A-8 A-8 A-8 A-5 A-5 B-5 C-5 C-5 C-5 D-4 E-4 E-3 G-3 G-2.5 H-2.5 — — —

0.5

A-10 A-10 A-10 B-10 B-10 C-8 D-8 D-5 D-5 E-5 F-4 G-4 — —

1

2

3

A-10 A-10 A-10 A-10 B-10 B-8 C-8 C-8 C-5 D-5 E-5 E-4 F-4 H-2 — — —

A-8 A-8 A-8 A-5 A-5 B-5 B-5 C-5 C-5 D-4 E-4 E-3 F-3 G-2.5 H-2 — — —

A-8 A-5 A-5 A-5 A-5 B-5 B-5 C-5 C-4 C-4 D-4 E-3 E-2.5 E-2.5 G-2.5 H-2 — — —

0.5

1

2

A-10 A-10 A-10 A-10 B-8 B-8 C-5 D-5 D-5 D-5 E-4 — — — —

A-10 A-8 A-8 A-8 A-5 A-5 A-5 B-5 B-5 C-5 D-4 D-4 — — — — — —

A-8 A-5 A-5 A-5 A-5 A-5 B-5 B-5 B-4 C-4 — — — — — — — — —

pg 17.15

Note: For a given duct thickness, letters indicate type (rigidity class) of duct reinforcement (see Table 17.3); numbers indicate maximum spacing (ft) between duct reinforcement. Use the same metal duct thickness on all duct sides. Source: Adapted with permission from SMACNA, HVAC Duct Construction Standards — Metal and Flexible. Refer to this standard for complete details.

SECOND PASS bzm 6/28/00

Up through 10 12 14 16 18 20 22 24 26 28 30 36 42 48 54 60 72 84 96 Over 96

0.0575 (16)

39445 Wang (MCGHP) Chap_17

Minimum galvanized steel thickness, in. (gauge)

17.15

__SH __ST __LG

DF

SH__ ST__ LG__

DF 17.16

TABLE 17.3 Transverse Joint Reinforcement Pressure, 4 in. WG maximum*

Hs T (min), in.

Standing S W, in.

Hs T (min), in.

Standing S

Standing S

Hs T (min), in.

Hs T (min), in.

Use class B

B

1 8 0.0.187

C

118 0.0296

—

1 0.0187

1 0.0187

D

—

—

1 0.0236

1 0.0236

1 8 0.0187

E

—

3

16

1 8 0.0356

—

118 0.0466

F

—

3

158 0.0296

—

112 0.0236

Standing S (angle-reinforced)

Hs T (min) plus reinforcement (H T), in.

2 0.0187

1

Use class C

1

16

1

2 0.0296

1

Use class D Use class F 1

112 0.0236 plus 112 18 bar

—

3

16

158 0.0466

—

112 0.0466

112 0.0296 plus 112 18 bar

H

—

—

—

—

—

112 0.0356 plus 112 112 316 angle

I

—

—

—

—

—

2 0.0356 plus 2 2 18 angle

J

—

—

—

—

—

2 0.0356 plus 2 2 316 angle

Acceptable to 36 in. length at 3 in. WG and to 30 in. length at 4 in. WG. Source: Adapted with permission from SMA CNA, HVAC Duct Construction Standar ds — Metal and Fle xible and ASHRAE Handbook 1988, Equipment. Refer to SMA CNA Standard for complete details. *

pg 17.16

G

SECOND PASS bzm 6/28/00

A

Standing S (bar-reinforced)

39445 Wang (MCGHP) Chap_17

Minimum rigidity Standing class drive slip

39445 Wang (MCGHP) Chap_17

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pg 17.17

AIR SYSTEMS: AIR DUCT DESIGN

17.17

TABLE 17.4 Round Ferrous Metal Duct Construction for Duct Systems in Commercial Buildings Minimum galvanized steel thickness, in. Pressure,  2 in. WG Duct diameter, in.

Spiral seam duct

Longitudinal seam duct

Up through 8 14 26 36 50 60 84

0.0157 0.0187 0.0236 0.0296 0.0356 0.0466 0.0575

0.0236 0.0236 0.0296 0.0356 0.0466 0.0575 0.0705

Pressure,  2 in. WG Fittings

Spiral seam duct

Longitudinal seam duct

Fittings

Suggested type of joint

0.0236 0.0236 0.0296 0.0356 0.0466 0.0575 0.0705

0.0157 0.0157 0.0187 0.0236 0.0296 0.0356 0.0466

0.0157 0.0187 0.0236 0.0296 0.0356 0.0466 0.0575

0.0187 0.0187 0.0236 0.0296 0.0356 0.0466 0.0575

Beaded slip Beaded slip Beaded slip Beaded slip Flange Flange Flange

Source: Adapted with permission from SMA CNA, HVAC Duct Construction Standar ds — Metal and Fle xible. Refer to SMA CNA Standard for complete details.

Duct area

Maximum spacing, ft (m)

Up to 4 ft (0.37 m ) Between 4 and 10 ft2 (0.37 and 0.93 m2) Larger than 10 ft2 (0.93 m2)

8 (2.4) 6 (1.8) 4 (1.2)

2

2

Round Ducts For a speci f ed cross-sectional area and mean air v elocity, a round duct has less f uid resistance against air f ow than rectangular and f at o val ducts. Round ducts also ha ve better rigidity and strength. The spiral- and longitude-seamed round ducts used in commercial b uildings are usually factory-fabricated to impro ve the quality and sealing of the ductw ork. The pressure losses can be calculated more precisely than for rectangular ducts, and result in a better balanced system. Air leakage can be maintained at about 3 percent as a result of well-sealed seams and joints. Round ducts have much smaller radiated noise break out from the duct than rectangular and f at oval ducts. The main disadvantage of round ducts is the greater space required under the beam for installation. Factory-fabricated spiral-seamed round ducts are the most widely used air ducts in commercial buildings. The standard diameters of round ducts range from 4 to 20 in. in 1-in. (100 to 500 mm in 25-mm) increments, from 20 to 36 in. in 2-in. (500 to 900 mm in 50-mm) increments, and from 36 to 60 in. in 4-in. (900 to 1500 mm in 100-mm) increments. The minimum thickness of galvanized sheet and f ttings for round ducts in duct systems in commercial b uildings is listed in Table 17.4. Many industrial air pollution control systems often require a velocity around 3000 fpm (15 m / s) or higher to transport particulates. Round ducts with thick er metal sheets are usually used in such applications.

Flat Oval Ducts Flat oval ducts, as shown in Fig. 17.4, have a cross-sectional shape between rectangular and round. They share the advantages of both the round and the rectangular duct with less large-scale air turbulence and a small depth of space required during installation. Flat o val ducts are quick er to install and have lower air leakage because of the factory fabrication. Flat oval ducts are made in either spiral seam or longitudinal seam. The minimum thickness of the galv anized sheet and f ttings for f at o val duct systems used in commercial b uildings is presented in Table 17.5.

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pg 17.18

CHAPTER SEVENTEEN

TABLE 17.5 Flat Oval Duct Construction for Positive-Pressure Duct Systems in Commercial Buildings Minimum galvanized steel thickness, in. Major axis, in.

Spiral seam duct

Longitudinal seam duct

Fittings

Suggested type of joint

Up through 24 36 48 60 70 Over 70

0.0236 0.0296 0.0296 0.0356 0.0356 0.0466

0.0356 0.0356 0.0466 0.0466 0.0575 0.0575

0.0356 0.0356 0.0466 0.0466 0.0575 0.0575

Beaded slip Beaded slip Flange Flange Flange Flange

Source: Adapted with permission from SMACNA, HVAC Duct Construction Standards — Metal and Flexible. Refer to SMACNA Standard for complete details.

Flexible Ducts Flexible ducts are often used to connect the main duct or the dif fusers to the terminal box. Their f exibility and ease of remo val allo w allocation and relocation of the terminal de vices. Fle xible ducts are usually made of multiple-ply polyester f lm reinforced by a helical steel wire core or corrugated aluminum spiral strips. The duct is often insulated by a f berglass blanket 1 or 2 in. (25 to 50 mm) thick. The outer surface of the f exible duct is usually co vered with aluminum foil or other types of vapor barriers to prevent the permeation of water vapor into the insulation layer. The inside diameter of f exible ducts may range from 2 to 10 in. in 1-in. (50 to 250 mm in 25-mm) increments and from 12 to 20 in. in 2-in. (300 to 500 mm in 50-mm) increments. The f exible duct should be as short as possible, and its length should be fully e xtended to minimize f ow resistance.

Fiberglass Ducts

SH__ ST__ LG__ DF

Fiberglass duct boards are usually made in 1-in. (25-mm) thickness. They are fabricated into rectangular ducts by closures. A f berglass duct with a 1.5-in. (38-mm) thickness may be used in the Gulf area of the United States where the climate is hot and humid in summer, to minimize duct heat gain. Round molded f berglass ducts are sometimes used. Fiberglass ducts have a good thermal performance. For a 1-in. (25-mm) thickness of duct board, the U value is 0.21 Btu / h ft2 °F at 2000 fpm (1.192 W / m2 °C at 10 m / s) air v elocity, which is better than a galvanized sheet metal duct with a 1-in. (25-mm) inner liner . Fiberglass duct has good sound attenuation characteristics. Its air leakage is usually 5 percent or less, which is f ar less than that of a sheet-metal rectangular duct that is not well sealed. Another important advantage of f berglass duct is its lower cost. The closures, also called taping systems, are tapes used to form rectangular duct sections from duct boards and to join the sections and f ttings into an integrated duct system. The improved acrylic pressure-sensitive tapes pro vide a better bond than before. Heat-sensiti ve solid polymer adhesi ve closures show themselves to be good sealing tapes even if dust, oil, or water is present on the surface of the duct board. Mastic and glass fabric closures are also used in many f berglass duct systems. Tests show that the ongoing emission of f berglass from the duct board w as less than that contained in outdoor air. Fiberglass ducts have a slightly higher friction loss than galv anized sheet duct (0.03 in. WC or 7.5 Pa greater for a length of 100 ft or 30.5 m). They are also not as strong as metal sheet duct. They must be handled carefully to prevent damage during installation. Fiberglass ducts are used in duct systems with a pressure dif ferential of  2 in. WC ( 500 Pa) or less. Man y codes restrict the use of f berglass in sensiti ve areas such as operating rooms and maternity wards.

39445 Wang (MCGHP) Chap_17

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pg 17.19

AIR SYSTEMS: AIR DUCT DESIGN

17.19

17.4 DUCT HEAT GAIN, HEAT LOSS, AND DUCT INSULATION Temperature Rise or Drop due to Duct Heat Gain or Loss The temperature rise or drop from duct heat gain or loss is one of the parameters that affect the supply air temperature as well as the supply v olume f ow rate in the air conditioning system design. Heat gain or loss through the duct w all of a rectangular duct section with a constant-v olume f ow rate qd, in Btu / h (W), can be calculated as



qd  UPL Tam 

Ten  T1v 2



(17.28)

For a round duct section



qd  UDdL Tam 

Ten  T1v 2



17.29)

where U  overall heat transfer coeff cient of duct wall, Btu / h ft2 °F (W / m2 °K) P, L  perimeter and length of duct, ft (m) Dd  diameter of round duct, ft (m) Tam  temperature of ambient air, °F (°C) Ten, Tlv  temperature of air entering and leaving duct section, °F (°C) The temperature increase or drop of the air f owing through a duct section is given as T1v  Ten 

qd 60Advscpa

(17.30)

where Ad  cross-sectional area of duct, ft2 (m2). In Eq. (17.30), the mean air v elocity v is expressed in fpm [m / (60 s)]. We substitute Eq. (17.28) into Eq. (17.30). For rectangular duct let y

120 Advscpa UPL

(17.31)

For round duct, let y

30Ddvscpa UL

(17.32)

Then the temperature of air leaving the duct section is T1v 

2Tam  Ten( y  1) y1

(17.33)

Duct Insulation Duct insulation is mounted or inner-lined to reduce heat loss and heat gain as well as to pre vent the condensation on the outer surf ace of the duct. It is usually in the form of duct wrap (outer surf ace), duct inner liner , or f berglass duct boards. Duct liner pro vides both thermal insulation and sound attenuation. The thickness of an insulation layer is based on economical analysis. ASHRAE / IESNA Standard 90.1-1999 mandates that all supply and return ducts and plenums installed as part of an HV AC&R air distrib ution system shall be thermally insulated as listed in Table 17.6. The insulated R-values in h ft2 °F / Btu (m2 °C / W), are for the insulation installed and

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39445 Wang (MCGHP) Chap_17

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CHAPTER SEVENTEEN

TABLE 17.6 Minimum Duct Insulation R-Value,* Cooling and Heating Supply Ducts and Return Ducts Duct location Climate zone Envelop criteria table

HDD65

CDD50

Exterior

Ventilated attic

Unvented Unvented attic with attic with backloaded roof Unconditioned ceiling insulation space†

Indirectly conditioned space§ Buried

Heating ducts only 5-1 to 5-7 5-8 to 5-12 5-13 to 5-15 5-16 to 5-18

0-1800 1801-3600 3601-5400 5401-7200

All All All All

None R-3.5 R-3.5 R-6

None None None R-3.5

None None None None

None None None None

None None None None

None None None None

None None None R-3.5

Cooling only ducts 5-15, 18, 20, 22 to 26 5-12, 14, 17, 19, 21 5-7, 9, 11, 13, 16 5-4, 6, 8, 10

All

0-1800

R-1.9

R-1.9

R-1.9

R-1.9

R-1.9

None

None

All

1801-3600

R-3.5

R-1.9

R-3.5

R-1.9

R-1.9

None

None

All

3601-5400

R-3.5

R-3.5

R-6

R-1.9

R-1.9

None

None

All

5401-7200

R-6

R-6

R-6

R-3.5

R-1.9

None

None

R-1.9 R-3.5 R-3.5 R-1.9 R-3.5 R-1.9 R-1.9

R-1.9 R-3.5 R-3.5 R-1.9 R-3.5 R-3.5 R-1.9

None None None None None None None

None

None

None

Combined heating and cooling ducts 5-9 5-10 5-11 5-12 5-13 5-14 5-15

1801-2700 2701-3600 2701-3600 2701-3600 3601-5400 3601-5400 3601-5400

0-5400 5401  3601-5400 0-3600 3601  1801-3600 0-1800

R-6 R-6 R-6 R-3.5 R-6 R-6 R-3.5

R-3.5 R-6 R-6 R-3.5 R-6 R-3.5 R-3.5

R-3.5

R-3.5

R-6 R-6 R-6 R-3.5 R-6 R-6 R-3.5

Return ducts 5-1 to 5-26

All climates

R-3.5

None

Insulation R-values, measured in (h ft  °F) / Btu, are for the insulation as installed and do not include f lm resistance. The required minimum thicknesses do not consider water vapor transmission and possible surface condensation. Where exterior walls are used as plenum walls, wall insulation shall be as required by the most restrictive condition of 6.2.4.2 or Section 5. Insulation resistance measured on a horizontal plane in accordance with ASTM C518 at a mean temperature of 75°F at the installed thickness. † Includes crawl spaces, both ventilated and nonventilated. § Includes return air plenums with or without exposed roofs above. Source: ASHRAE / IESNA Standard 90.1-1999. Reprinted by permission. *

2

do not include air f lm resistances. The required minimum thickness does not consider w ater vapor transmission and possible surface condensation. The recommended thickness of insulation layer , or duct wrap, for R-3.5 h ft2 °F / Btu (0.62 m 2 °C / W) is 1 to 1.5 in. (25 to 38 mm), and for R-6 h  ft2 °F / Btu (1.06 m2 °C / W) is 2 to 3 in. (50 to 75 mm). Exceptions of duct and plenum insulation include ●

●

●

●

SH__ ST__ LG__ DF

Factory-installed plenums, casings, and ductwork as part of the equipment Ducts or plenums located in heated spaces, semiheated spaces, or cooled spaces For runout less than 10 ft (3 m) in length to air terminals or air outlets, R-value of insulation need not exceed R-3.5. Backs of air outlets and outlet plenums e xposed to unconditioned or indirectly conditioned space with face area e xceeding 5 ft 2 (0.5 m 2) need not e xceed R-2, for those 5 ft 2 (0.5 m 2) or smaller need not be insulated.

39445 Wang (MCGHP) Chap_17

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pg 17.21

AIR SYSTEMS: AIR DUCT DESIGN

17.21

If the temperature of the ambient air is 80°F (26.7°C) with a relative humidity of 50 percent, its dew point is 59°F (18.3°C). Only when the outer surface temperature of the duct Tsd 59°F (15°C) will condensation not occur. Refer to ASHRAE Standard 90.1-1999 for details.

Temperature Rise Curves Duct heat gain or loss and the temperature rise or drop of the air inside the duct depend on air velocity, duct dimensions, and duct insulation. Figure 17.5 sho ws curves for the temperature rise in round ducts. These curves are calculated according to Eqs. (17.30) to (17.33) under these conditions: ●

●

Thickness of the insulation layer of duct wrap is 1.5 in. (38 mm), and the thermal conductivity of the insulating material k  0.30 Btu in / h ft2 °F (0.043 W / m°C). Heat transfer coeff cient of the outer surface of the duct h  1.6 Btu / h ft2 °F (9.1 W / m2 °C)

●

(17.34)

Convective heat-transfer coeff cient of the inside surface hc, in Btu / h  ft °F, can be calculated by 2

h c  0.023

 Dk Re d

0.4 0.8 D Pr

(17.35)

where ReD  Reynolds number based on duct diameter as characteristic length Pr  Prandtl number ●

●

Air temperature inside the air duct is assumed to be 55°F (12.8°C). Temperature difference between the air inside the duct and the ambient air surrounding the duct is 25°F (13.9°C).

FIGURE 17.5 Temperature rise curves from duct heat gain.

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pg 17.22

CHAPTER SEVENTEEN

If the temperature rise or drop has been determined, duct heat gain or heat loss can be either expressed in percentage of supply temperature dif ferential or calculated by Eqs. (17.28) and (17.29).

17.5 FRICTIONAL LOSSES In an air duct system, there are tw o types of resistance against the air f ow: frictional losses and dynamic losses.

Darcey-Weisbach Equation Frictional losses result mainly from the shearing stress between the f uid layers of the laminar sublayer, which is adjacent to the surf ace of the duct w all. Friction also e xists when the f uid particles in the turbulent f ow bump against the protuberances of the duct w all. These lead to the production of eddies and energy loss. Friction losses occur along the entire length of the air duct. For a steady, incompressible f uid f ow in a round duct or a circular pipe, friction head loss Hf, in ft (m) of air column, can be calculated by the Darcy-Weisbach equation in the following form: Hf  f

 DL  2gv  2

(17.36)

If the friction loss is presented in the form of pressure loss pf, in lbf / ft2 (Pa), then pf  f

v 2

L D

  2g 

(17.37)

c

where f  friction factor L  length of duct or pipe, ft (m) D  diameter of duct or pipe, ft (m) v  mean air velocity in duct, ft / s (m / s) If air v elocity v is e xpressed in fpm and pf is e xpressed in in. WC, then a con version f actor Fcv  (1 / 5.19)(1 / 60)2  0.0000535 should be used. That is, pf  0.0000535f

v 2

 DL  2g  c

(17.38)

In Eqs. (17.36), (17.37), and (17.38), strictly speaking, D should be replaced by hydraulic diameter Dh. But for round ducts and pipes, D  Dh. Friction Factor

SH__ ST__ LG__ DF

Friction loss is directly proportional to the friction factor f, which is dimensionless. The relationship between f and the parameters that in f uence the magnitude of the friction f actor is sho wn in Fig. 17.6, which is called a Moody diagram. In Fig. 17.6, the term  represents the absolute roughness of the surf ace protuberances, expressed in ft (m), as shown in Fig. 17.7. The ratio  / D indicates the relative roughness of the duct or pipe. For laminar f ow in an air duct when Re D 2000, f is affected mainly by the viscous force of the f uid f ow and, therefore, is a function of ReD only. That is, f

64 Re D

(17.39)

39445 Wang (MCGHP) Chap_17

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pg 17.23

AIR SYSTEMS: AIR DUCT DESIGN

17.23

FIGURE 17.6 Moody diagram. (Source: Moody, L.F., Transactions A.S.M.E., vol. 66, 1994. Reprinted with permission.)

In an ideal smooth tube or duct, that is,  / D 105, if Re D 4000, the surface roughness is submerged in the laminar sublayer with a thickness of , and the f uid moves smoothly, passing over the protuberances. In this case, f decreases with an increase of Re D. The relationship between f and ReD may be expressed by the Blasius empirical formula 0.316 f Re 0.25 D

(17.40)

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CHAPTER SEVENTEEN

FIGURE 17.7 Modes of air f ow when air passes o ver and around surface protuberances of the duct wall: (a)  ; (b)  .

With a further increase of Re D, the laminar sublayer becomes thinner , even thinner than the height of the irre gularities , that is,  . The protuberances form the separation of f uid f ow, enhance the formation of v ortices, and, therefore, increase the pressure loss as well as the v alue of f at greater ReD. If ReD is beyond a limit called the Rouse limit, f depends mainly on the relative roughness of the duct wall. The Rouse limit line can be determined by Re D  200 / √f(/D) , as shown in Fig. 17.6. Duct Friction Chart In most air ducts, ReD ranges from 1 104 to 2 106, and  / D may vary from 0.005 to 0.00015. This covers a transition zone between hydraulic smooth pipes and the Rouse limit line. Within this region, f is a function of both Re D and  / D. Colebrook (1939) recommended the follo wing empirical formula to relate ReD,  / D, and f for air ducts: 1

√f

 2 log

 2.51   3.7D  Re √f

(17.41)

D

In Eq. (17.41),  and D must be expressed in the same units. Swamee and Jain suggested the use of an explicit expression that can give approximately the same value of f as Colebrook’s formula: f

SH__ ST__ LG__ DF

0.25 {log [ / (3.7D)  5.74 / (0.9 Re D)]}2

(17.42)

In air duct calculations, a rough estimate of f can be determined from the Moody diagram based on knowing the relative roughness  / D and ReD. For practical calculations, a friction chart for round ducts, developed by Wright, in the form shown in Fig. 17.8, is widely used. In this chart, air volume f ow rate V˙ , in cfm, and the frictional

39445 Wang (MCGHP) Chap_17

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pg 17.25

AIR SYSTEMS: AIR DUCT DESIGN

17.25

loss per unit length pf, in in. WC per 100 ft, are used as coordinates. The mean air v elocity v, in fpm, and the duct diameter are shown in inclined lines in this chart. The duct friction chart can be applied to the following conditions without corrections: ●

●

●

●

A temperature from 41 to 95°F (5 to 35°C) Up to an elevation of 1600 ft (488 m) Duct pressure difference of  20 in. WC ( 5000 Pa) with respect to ambient air pressure Duct material of medium-smooth roughness

Roughness and Temperature Corrections The absolute roughness , in ft, given in ASHRAE Handbook 1989, Fundamentals, is listed in Table 17.7. When duct material dif fers from medium-smooth roughness, or when the air duct is installed at an ele vation abo ve 1600 ft (488 m), or when w arm air is supplied at a temperature higher than 95°F (35°C), then corrections should be made to pf as follows: pf  K srK TK e1 pf,c

(17.43)

where pf,c  friction loss found from the duct friction chart, in. WC (Pa). In Eq. (17.43), Ksr indicates the correction f actor for surf ace roughness, which is dimensionless, and can be calculated as K sr 

fa fc

(17.44)

Here fc denotes the friction factor of duct material of surf ace roughness that is speci f ed by the duct friction chart, that is,   0.0005 ft (0.09 mm). The symbol fa represents the actual friction factor of duct material with surf ace roughness dif fering from fc. Both fa and fc can be calculated from Eq. (17.41) or (17.42). TABLE 17.7 Duct Roughness Duct material (roughness, ft)

Roughness category

Absolute roughness , ft

Uncoated carbon steel, clean (0.00015 ft) PVC plastic pipe (0.00003 – 0.00015 ft) Aluminum (0.00015 – 0.0002 ft)

Smooth

0.0001

Galvanized steel, longitudinal seams, 4-ft joints (0.00016 – 0.00032 ft) Galvanized steel, spiral seams, with 1, 2, and 3 ribs, 12-ft joints (0.00018 – 0.00038 ft)

Medium smooth

0.0003

Galvanized steel, longitudinal seams, 2.5-ft joints (0.0005 ft)

Average

0.0005

Fibrous glass duct, rigid Fibrous glass duct liner, air side with facing material (0.0005 ft)

Medium rough

0.003

Fibrous glass duct liner, air side spray-coated (0.015 ft) Flexible duct, metallic (0.004 – 0.007 ft when fully extended) Flexible duct, all types of fabric and wire (0.0035 – 0.015 ft when fully extended) Concrete (0.001 – 0.01 ft)

Rough

0.01

Source: ASHRAE Handbook 1989, Fundamentals. Reprinted with permission.

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CHAPTER SEVENTEEN

FIGURE 17.8 Friction chart for round ducts. (Source: ASHRAE Handbook 1989. Reprinted with permission.)

The term KT indicates the correction factor for air temperature inside the duct, which affects the density of the air; KT is dimensionless and can be calculated as KT 

 T 530  460 

0.825

a

(17.45)

where Ta  actual air temperature inside duct, °F. The term Kel indicates the correction f actor for elevation, which also affects the density of the air; it is dimensionless. When the elevation is greater than 1600 ft (488 m), Kel can be evaluated as K el   SH__ ST__ LG__ DF

pat 29.92 pat,mm 760

where pat  actual atmospheric or barometric pressure, in. Hg pat,mm  actual atmospheric or barometric pressure, mm Hg

(17.46)

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39445 Wang (MCGHP) Chap_17

pg 17.27

AIR SYSTEMS: AIR DUCT DESIGN

17.27

Example 17.1. A fabric and wire f exible duct of 8-in. (200 mm) diameter installed in a commercial building at sea level has a surface roughness   0.12 in. (3 mm). The mean air velocity inside the air duct is 800 fpm (4 m / s). Find the correction factor of surface roughness. The viscosity of air at 60°F is 1.21 105 lb / ft s (2.0 105 Pas). Solution. For the galv anized sheet duct speci f ed in the duct friction chart ,   0.0005 ft, or 0.006 in. (0.15 mm). Then Re D 

vD 



0.075 1000 8  5.5 104 60 12 1.21 105

From Eq. (17.42), fc  

0.25 {log [ / (3.7D)  5.74 / (0.9 Re D)]}2 0.25  0.0204 {log [0.006 / (3.7 8)  5.74 / (0.9 5.5 104)]}2

For the fabric and wire f exible duct, ReD is the same as for the galvanized sheet duct. From the data given,   0.12 in., so the actual friction factor is fa 

0.25  0.044 {log [0.12 / (3.7 8)  5.74 / (0.9 5.5 104)]}2

From Eq. (17.44), the correction factor of surface roughness is K sr 

fa 0.044   2.16 fc 0.0204

Circular Equivalents In Eq. (17.37), if D is replaced by Dh, then pf  f

v 2

 DL  2g  h

c

Apparently, for circular or noncircular air ducts with different cross-sectional shapes and the same hydraulic diameter, the pressure loss is the same for an equal length of air duct at equal mean air v elocities. Circular equi valents are used to con vert the dimension of a noncircular duct into an equi valent diameter De, in in. (mm), of a round duct when their volume f ow rates V˙ and frictional losses per unit length pf,u are equal. A noncircular duct must be con verted to a circular equivalent f rst before determining its pf,u from the duct friction chart. The hydraulic diameter Dh, in in. (mm), is def ned as Dh 

4A P

(17.47)

where A  area, in.2 (mm2) and P  perimeter, in. (mm). Based on e xperimental results, Heubscher (1948) recommended the following formula to calculate De for rectangular duct at equal V˙ and pf,u: De 

1.30(ab)0.625 (a  b)0.25

(17.48)

where a, b  dimensions of tw o sides of rectangular duct, in. (mm). The circular equi valents of rectangular ducts at various dimensions calculated by Eq. (17.48) are listed in Table 17.8.

__SH __ST __LG DF

SH__ ST__ LG__

DF 17.28

TABLE 17.8 Circular Equivalents of Rectangular Ducts for Equal Friction and Capacity

3.0 3.5 4.0 4.5 5.0 5.5 One side, rectangular duct, in.

4.5

5.0

5.5

6.0

6.5

7.0

7.5

8.0

9.0

10.0

11.0

12.0

13.0

14.0

15.0

16.0

3.8 4.1 4.4 4.6 4.9 5.1

4.0 4.3 4.6 4.9 5.2 5.4

4.2 4.6 4.9 5.2 5.5 5.7

4.4 4.8 5.1 5.4 5.7 6.0

4.6 5.0 5.3 5.7 6.0 6.3

4.7 5.2 5.5 5.9 6.2 6.5

4.9 5.3 5.7 6.1 6.4 6.8

5.1 5.5 5.9 6.3 6.7 7.0

5.2 5.7 6.1 6.5 6.9 7.2

5.5 6.0 6.4 6.9 7.3 7.6

5.7 6.3 6.7 7.2 7.6 8.0

6.0 6.5 7.0 7.5 8.0 8.4

6.2 6.8 7.3 7.8 8.3 8.7

6.4 7.0 7.6 8.1 8.6 9.0

6.6 7.2 7.8 8.4 8.9 9.3

6.8 7.5 8.0 8.6 9.1 9.6

7.0 7.7 8.3 8.8 9.4 9.9

Adjacent side, in. 6

7

6.6 7.1 7.6 8.0 8.4

7.7 8.2 8.7 9.1

8

9

10

11

12

13

14

15

16

17

18

19

20

22

24

26

28

30

Side rectangular duct 6 7 8 9 10

8.7 9.3 9.8 9.8 10.4 10.9

8.8 9.5 10.2 10.9 9.1 9.9 10.7 11.3 9.5 10.3 11.1 11.8 9.8 10.7 11.5 12.2 10.1 11.0 11.8 12.6

11.5 12.0 12.4 12.9 13.3

12.0 12.6 13.1 13.5 14.0

13.1 13.7 14.2 14.2 14.7 15.3 14.6 15.3 15.8 16.4

11 12 13 14 15

16 17 18 19 20

10.4 10.7 11.0 11.2 11.5

11.3 11.6 11.9 12.2 12.5

12.2 12.5 12.9 13.2 13.5

13.0 13.4 13.7 14.1 14.4

13.7 14.1 14.5 14.9 15.2

14.4 14.9 15.3 15.7 16.0

15.1 15.6 16.0 16.4 16.8

15.7 16.2 16.7 17.1 17.5

16.4 16.8 17.3 17.8 18.2

16.9 17.4 17.9 18.4 18.9

17.5 18.0 18.5 19.0 19.5

18.6 19.1 19.6 20.1

19.7 20.2 20.8 20.7 21.3 21.9

22 24 26 28 30

12.0 12.4 12.8 13.2 13.6

13.0 13.5 14.0 14.5 14.9

14.1 14.6 15.1 15.6 16.1

15.0 15.6 16.2 16.7 17.2

15.9 16.5 17.1 17.7 18.3

16.8 17.4 18.1 18.7 19.3

17.6 18.3 19.0 19.6 20.2

18.3 19.1 19.8 20.5 21.1

19.1 19.9 20.6 21.3 22.0

19.8 20.6 21.4 22.1 22.9

20.4 21.3 22.1 22.9 23.7

21.1 22.0 22.9 23.7 24.4

21.7 22.7 23.5 24.4 25.2

22.3 23.3 24.2 25.1 25.9

22.9 23.9 24.9 25.8 26.6

24.0 25.1 26.1 27.1 28.0

26.2 27.3 28.3 29.3

28.4 29.5 30.6 30.5 31.7

32.8

22 24 26 28 30

32 34 36 38 40

14.0 14.4 14.7 15.0 15.3

15.3 15.7 16.1 16.5 16.8

16.5 17.0 17.4 17.8 18.2

17.7 18.2 18.6 19.0 19.5

18.8 19.3 19.8 20.2 20.7

19.8 20.4 20.9 21.4 21.8

20.8 21.4 21.9 22.4 22.9

21.8 22.4 22.9 23.5 24.0

22.7 23.3 23.9 24.5 25.0

23.5 24.2 24.8 25.4 26.0

24.4 25.1 25.7 26.4 27.0

25.2 25.9 26.6 27.2 27.9

26.0 26.7 27.4 28.1 28.8

26.7 27.5 28.2 28.9 29.6

27.5 28.3 29.0 29.8 30.5

28.9 29.7 30.5 31.3 32.1

30.2 31.0 32.0 32.8 33.6

31.5 32.4 33.3 34.2 35.1

33.9 34.9 35.9 36.8 37.8

32 34 36 38 40

16 17 18 19 20

32.7 33.7 34.6 35.6 36.4

pg 17.28

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6 7 8 9 10

Adjacent side, in. 4.0

39445 Wang (MCGHP) Chap_17

One side, rectangular duct, in.

17.1 17.5 17.8 18.1 18.4

18.5 18.9 19.3 19.6 19.9

19.9 20.3 20.6 21.0 21.4

21.1 21.5 21.9 22.3 22.7

22.3 22.7 23.2 23.6 24.0

23.4 23.9 24.4 24.8 25.2

24.5 25.0 25.5 26.0 26.4

25.6 26.1 26.6 27.1 27.6

26.6 27.1 27.7 28.2 28.7

27.6 28.1 28.7 29.2 29.8

28.5 29.1 29.7 30.2 30.8

29.4 30.0 30.6 31.2 31.8

30.3 30.9 31.6 32.2 32.8

31.2 31.8 32.5 33.1 33.7

32.8 33.5 34.2 34.9 35.5

34.4 35.1 35.9 36.6 37.2

35.9 36.7 37.4 38.2 38.9

37.3 38.1 38.9 39.7 40.5

38.7 39.5 40.4 41.2 42.0

42 44 46 48 50

52 54 56 58 60

17.1 17.3 17.6 17.8 18.1

18.7 19.0 19.3 19.5 19.8

20.2 20.6 20.9 21.2 21.5

21.7 22.0 22.4 22.7 23.0

23.1 23.5 23.8 24.2 24.5

24.4 24.8 25.2 25.5 25.9

25.7 26.1 26.5 26.9 27.3

26.9 27.3 27.7 28.2 28.6

28.0 28.5 28.9 29.4 29.8

29.2 29.7 30.1 30.6 31.0

30.3 30.8 31.2 31.7 32.2

31.3 31.8 32.3 32.8 33.3

32.3 32.9 33.4 33.9 34.4

33.3 33.9 34.4 35.0 35.5

34.3 34.9 35.4 36.0 36.5

36.2 36.8 37.4 38.0 38.5

37.9 38.6 39.2 39.8 40.4

39.6 40.3 41.0 41.6 42.3

41.2 41.9 42.7 43.3 44.0

42.8 43.5 44.3 45.0 45.7

52 54 56 58 60

20.1 20.3 20.6 20.8 21.0

21.7 22.0 22.3 22.6 22.8

23.3 23.6 23.9 24.2 24.5

24.8 25.1 25.5 25.8 26.1

26.3 26.6 26.9 27.3 27.6

27.6 28.0 28.4 28.7 29.1

28.9 29.3 29.7 30.1 30.4

30.2 30.6 31.0 31.4 31.8

31.5 31.9 32.3 32.7 33.1

32.6 33.1 33.5 33.9 34.4

33.8 34.3 34.7 35.2 35.6

34.9 35.4 35.9 36.3 36.8

36.0 36.5 37.0 37.5 37.9

37.1 37.6 38.1 38.6 39.1

39.1 39.6 40.2 40.7 41.2

41.0 41.6 42.2 42.8 43.3

42.9 43.5 44.1 44.7 45.3

44.7 45.3 46.0 46.6 47.2

46.4 47.1 47.7 48.4 49.0

62 64 66 68 70

23.1 23.3 23.6 23.8 24.1

24.8 25.1 25.3 25.6 25.8

26.4 26.7 27.0 27.3 27.5

27.9 28.2 28.5 28.8 29.1

29.4 29.7 30.0 30.4 30.7

30.8 31.2 31.5 31.8 32.2

32.2 32.5 32.9 33.3 33.6

33.5 33.9 34.3 34.6 35.0

34.8 35.2 35.6 36.0 36.3

36.0 36.4 36.8 37.2 37.6

37.2 37.7 38.1 38.5 38.9

38.4 38.8 39.3 39.7 40.2

39.5 40.0 40.5 40.9 41.4

41.7 42.2 42.7 43.2 43.7

43.8 44.4 44.9 45.4 45.9

45.8 46.4 47.0 47.5 48.0

47.8 48.4 48.9 49.5 50.1

49.6 50.3 50.9 51.4 52.0

72 74 76 78 80

26.1 26.4 26.6 26.9 27.1

27.8 28.1 28.3 28.6 28.9

29.4 29.7 30.0 30.3 30.6

31.0 31.3 31.6 31.9 32.2

32.5 32.8 33.1 33.4 33.8

34.0 34.3 34.6 34.9 35.3

35.4 35.7 36.1 36.4 36.7

36.1 37.1 37.4 37.8 38.2

38.0 38.4 38.8 39.2 39.5

39.3 39.7 40.1 40.5 40.9

40.6 41.0 41.4 41.8 42.2

41.8 42.2 42.6 43.1 43.5

44.1 44.6 45.0 45.5 45.9

46.4 46.9 47.3 47.8 48.3

48.5 49.0 49.6 50.0 50.5

50.6 51.1 51.6 52.2 52.7

52.6 53.2 53.7 54.3 54.8

82 84 86 88 90

51.0 53.2 52.0 54.2

55.3 56.4

92 96

88

Side rectangular duct

62 64 66 68 70 72 74 76 78 80 82 84 86 88 90 92 96

29.1 30.8 32.5 34.1 35.6 37.1 38.5 39.9 29.6 31.4 33.0 34.7 36.2 37.7 39.2 40.6

One side, rectangular duct, in. 32

41.3 42.6 43.9 42.0 43.3 44.7

46.4 48.7 47.2 49.6

Adjacent side, in. 34

36

38

40

35.0 36.1 37.1 38.1 39.0

37.2 38.2 39.4 39.3 40.4 41.5 40.3 41.5 42.6 43.7

42 44 46 48 50

40.0 40.9 41.8 42.6 43.6

41.2 42.2 43.1 44.0 44.9

42.5 43.5 44.4 45.3 46.2

43.7 44.7 45.7 46.6 47.5

44.8 45.8 46.9 47.9 48.8

44

46

48

50

52

56

60

64

68

72

76

80

84

32 34 36 38 40 45.9 47.0 48.0 46.1 50.0

48.1 49.2 50.3 50.2 51.4 52.5 51.2 52.4 53.6 54.7

42 44 46 48 50

pg 17.29

17.29

32 34 36 38 40

42

SECOND PASS bzm 6/28/00

15.6 15.9 16.2 16.5 16.8

39445 Wang (MCGHP) Chap_17

42 44 46 48 50

__SH __ST __LG

DF

SH__ ST__ LG__

DF 17.30

One side, rectangular duct, in. 32

Adjacent side, in. 34

36

38

40

42

44

46

48

50

52

56

60

64

68

72

76

80

45.7 46.5 47.3 48.1 48.9

47.1 48.0 48.8 49.6 50.4

48.4 49.3 50.2 51.0 51.9

49.7 50.7 51.6 52.4 53.3

51.0 52.0 52.9 53.8 54.7

52.2 53.2 54.2 55.1 60.0

53.4 54.4 55.4 56.4 57.3

54.6 55.6 56.6 57.6 58.6

55.7 56.8 57.8 58.8 59.8

56.8 57.9 59.0 61.2 60.0 62.3 61.0 63.4

65.6

62 64 66 68 70

48.0 48.7 49.4 50.1 50.8

49.6 50.4 51.1 51.8 52.5

51.2 51.9 52.7 53.4 54.1

52.7 53.5 54.2 55.0 55.7

54.1 54.9 55.7 56.6 57.3

55.5 56.4 57.2 58.0 58.8

56.9 57.8 58.6 59.4 60.3

58.2 59.1 60.0 60.8 61.7

59.5 60.4 61.3 62.2 63.1

60.8 61.7 62.6 63.6 64.4

62.0 63.0 63.9 64.9 65.8

64.4 65.4 66.4 67.4 68.3

66.7 67.7 68.8 69.8 70.8

70.0 71.0 72.1 74.3 73.2 75.4

72 74 76 78 80

51.4 52.1 52.7 53.3 53.9

53.2 53.8 54.5 55.1 55.8

54.8 55.5 56.2 56.9 57.5

56.6 57.2 57.9 58.6 59.3

58.0 58.8 59.5 60.2 60.9

59.6 60.3 61.1 61.8 62.6

61.1 61.9 62.6 63.4 64.1

62.5 63.3 64.1 64.9 65.7

63.9 64.8 65.6 66.4 67.2

65.3 66.2 67.0 67.9 68.7

66.7 67.5 68.4 69.3 70.1

69.3 70.2 71.1 72.0 72.9

71.8 72.7 73.7 74.6 75.4

74.2 75.2 76.2 77.1 78.1

76.5 77.5 78.6 79.6 80.6

78.7 79.8 80.9 83.1 81.9 84.2 82.9 85.2

87.5

82 84 86 88 90

54.5 55.1 55.7 56.3 56.8

56.4 57.0 57.6 58.2 58.8

58.2 58.8 59.4 60.1 60.7

59.9 60.6 61.2 61.9 62.5

61.6 62.3 63.0 63.6 64.3

63.3 64.0 64.7 65.4 66.0

64.9 65.6 66.3 67.0 67.7

66.5 67.2 67.9 68.7 69.4

68.0 68.7 69.5 70.2 71.0

69.5 70.3 71.0 71.8 72.6

70.9 71.7 72.5 73.3 74.1

73.7 74.6 75.4 76.3 77.1

76.4 77.3 78.2 79.1 79.9

79.0 80.0 80.9 81.8 82.7

81.5 82.5 83.5 84.4 85.3

84.0 85.0 85.9 86.9 87.9

88.5 89.6 90.7 91.7 92.7

92 94 96

57.4 59.3 61.3 63.1 64.9 66.7 68.4 70.1 71.7 73.3 74.9 77.9 57.9 59.9 61.9 63.7 65.6 67.4 69.1 70.8 72.4 74.0 75.6 78.7 58.4 60.5 62.4 64.3 66.2 68.0 69.7 71.5 73.1 74.8 76.3 79.4

Source: ASHRAE Handbook 1989, Fundamentals. Reprinted with permission.

52 54 56 58 60

80.8 83.5 86.2 81.6 84.4 87.1 82.4 85.3 88.0

62 64 66 68 70

86.3 87.3 88.3 89.3 90.3

88.8 91.3 89.7 92.3 90.7 93.2

72 74 76 78 80

96.2 97.3

82 84 86 88 90

93.7 96.1 98.4 94.7 97.1 99.4 95.7 98.1 100.5

92 94 96

91.8 92.9 94.0 95.0

pg 17.30

44.3 45.1 45.8 46.6 47.3

88

SECOND PASS bzm 6/28/00

52 54 56 58 60

84

Side rectangular duct

39445 Wang (MCGHP) Chap_17

TABLE 17.8 (Continued )

39445 Wang (MCGHP) Chap_17

SECOND PASS bzm 6/28/00

pg 17.31

AIR SYSTEMS: AIR DUCT DESIGN

17.31

For galv anized steel f at o val ducts with spiral seams, Heyt and Diaz (1975) proposed the following formula to calculate the circular equivalent for use of the duct friction chart: De 

1.55A0.625 P 0.25

(17.49)

Here A is the cross-sectional area of the f at oval duct, in.2 (mm2), and is given as A

b 2  b(a  b) 4

(17.50)

and the perimeter P, in in. (mm), is calculated as P  b  2(a  b)

(17.51)

The dimensions a and b of the f at oval duct are shown in Fig. 17.4c.

17.6 DYNAMIC LOSSES When air f ows through duct f ttings, such as, elbows, tees, diffusers, contractions, entrances and exits, or certain equipment, a change in velocity or direction of f ow may occur. Such a change leads to f ow separation and the formation of eddies and disturbances in that area. The energy loss resulting from these eddies and disturbances is called dynamic loss pdy, in in. WC (Pa). Although a duct f tting is fairly short, the disturbances it produces may persist over a considerable distance downstream. In addition to the presence of dynamic loss pdy, frictional loss pf occurs when an airstream f ows through a duct f tting. For convenience in calculation, the length of the duct f tting is usually added to the adjacent duct sections connected with this duct f tting of the same mean air velocity. When airstreams of the same Reynolds number f ow through geometrically similar duct f ttings, that is, in dynamic similarity , the dynamic loss is proportional to their v elocity pressure pv. Dynamic loss may be calculated as pdy  Co pv 

Cov 2o 2gccf

vo

2

 4005 

 Co

(17.52)

where Co  local loss coeff cient or dynamic loss coeff cient   air density, lb / ft3 (kg / m3) vo  mean air velocity of airstream at reference cross section o, fpm (m / s) gc  dimensional constant, 32.2 lbm ft / lbf s2, for SI units, gc  1 cf  conversion factor, for SI units, cf  1 Because the mean velocity of the airstream may vary at different ends of a duct f tting, Co is always specif ed with respect to a velocity of a reference cross section o in the duct f tting. Elbows An elbow is a duct f tting in which the air f ow changes direction. Elbo ws are sho wn in Figs. 17.9 and 17.10. Consider an elbo w that mak es a 90 ° turn in a round duct, as sho wn in Fig. 17.9 a. Because of the change of airstream direction, centrifugal force is created and acts to ward the outer wall of the duct. When the airstream f ows from the straight part of the duct to the curv ed part, it is accompanied by an increase in pressure and a decrease in air v elocity at the outer wall. At the same time, a decrease in pressure and an increase in air v elocity take place at the inner w all. Therefore, a diffuser effect occurs near the outer w all, and a bell mouth forms near the inner w all. After turning, the opposite effect takes place as the airstream f ows from the curved part to the straight part of the

__SH __ST __LG DF

39445 Wang (MCGHP) Chap_17

17.32

SECOND PASS bzm 6/28/00

pg 17.32

CHAPTER SEVENTEEN

FIGURE 17.9 Round elbows: (a) region of eddies and turbulences in a round elbo w; ( b) f ve-piece 90 ° round elbow.

duct. The diffuser effects lead to f ow separations from both walls, and eddies and large-scale turbulence form in regions AB and CE, as shown in Fig. 17.9a. In a rectangular elbow, a radial pressure gradient is also formed by the centrifugal force along its centerline NR, as sho wn in Fig. 17.10 c. A secondary circulation is formed along with the main forward airstream. The magnitude of the local loss coeff cient of an elbow is inf uenced by the following factors: ●

●

●

●

●

SH__ ST__ LG__ DF

Turning angle of the elbow Ratio of centerline radius (CLR) to diameter  Rc / D or Rc / W, where Rc represents the throat radius and W the width of the duct, both in in. (mm), as shown in Fig. 17.9a A three-gore (number of pieces), f ve-gore, or pleated seven-gore 90° elbow (Fig. 17.9b) Installation of splitter vanes, which reduce the eddies and turbulence in an elbow Shape of cross-sectional area of the duct

As the CLR becomes greater , the f ow resistance of the airstream becomes smaller . Ho wever, a greater CLR requires more duct material, a higher labor cost, and a larger allocated space. A value of Rc / D (CLR)  1 or 1.5 is often used if the space is available. The installation of splitter v anes, or turning vanes, in rectangular ducts can ef fectively reduce the pressure loss at the elbo w. For a rectangular 90 ° elbow with Rc / W  0.75 and tw o splitter

39445 Wang (MCGHP) Chap_17

SECOND PASS bzm 6/28/00

pg 17.33

AIR SYSTEMS: AIR DUCT DESIGN

17.33

FIGURE 17.10 Rectangular elbo ws: (a) rectangular elbo w, smooth radius, two splitter v anes; ( b) mitered elbo w; ( c) secondary f ow in a mitered elbow.

vanes, as shown in Fig. 17.10 a, the local loss coef ficient Co is only 0.04 to 0.05. The values of Ro, R1, and R2 can be found from Table 17.11. F or details, refer to ASHRAE Handbook 1997 , Fundamentals. For a mitered rectangular 90 ° elbow without turning vanes, Co is about 1.1. If turning v anes are installed, Co drops to only 0.12 to 0.18. Installation of splitter v anes in mitered elbo ws is also

__SH __ST __LG DF

39445 Wang (MCGHP) Chap_17

17.34

SECOND PASS bzm 6/28/00

pg 17.34

CHAPTER SEVENTEEN

FIGURE 17.11 Converging and diverging wyes: (a) converging wye, rectangular,   90°; ( b) di verging wye, rectangular,   90°.

advantageous for noise control because of the smaller f an total pressure and, therefore, lower fan noise as well as lower airf ow noise at the elbow.

Converging and Diverging Tees and Wyes A branch duct that combines with or di verges from the main duct at an angle of 90 ° is called a tee. However, if the angle lies between 15 ° and 75°, it is called a wye (Figs. 17.11 and 17.12). Tees and wyes can be round, f at oval, or rectangular. Various types of converging and diverging tees and wyes for round and f at oval ducts appear in Fig. 17.12. The function of a con verging tee or wye is to combine the airstream from the branch duct with the airstream from the main duct. The function of a di verging tee or wye is to di verge part of the airf ow from the main duct into the branch tak eoff. For airstreams f owing through a converging or diverging tee or wye, the dynamic losses for the main stream can be calculated as pc,s  SH__ ST__ LG__ DF

ps,c 

Cc,sv 2c 2gccf Cs,cv 2c 2gccf

vc

2

vc 4005

2

 4005 

 Cc,s  Cs,c





(17.53)

39445 Wang (MCGHP) Chap_17

SECOND PASS bzm 6/28/00

pg 17.35

__SH __ST __LG

FIGURE 17.12 Round and f at oval tees and wyes: (a) round tees, wyes, and cross; ( b) diverging tees and wyes with elbows; (c) f at oval tees and wyes.

17.35

DF

39445 Wang (MCGHP) Chap_17

17.36

SECOND PASS bzm 6/28/00

pg 17.36

CHAPTER SEVENTEEN

FIGURE 17.13 Airf ows through a rectangular con verging or diverging wye.

For the branch stream, pc,b  pb,c 

Cc,bv 2c 2gccf Cb,cv 2c 2gccf

 Cc,b

 4005 

vc

2

 Cb,c

vc 4005

2





(17.54)

In Eqs. (17.53) and (17.54), subscript c represents the common end, s the straight-through end, and b the branch tak eoff. The subscript c,s indicates the f ow of main stream from the common end to the straight-through end, and s,c the f ow from the straight-through end to the common end. Similarly, c,b denotes the f ow of the branch stream from the common end to the branch tak eoff, and b,c the airf ow from the branch duct to the common end. The airstreams f owing through a rectangular converging or diverging tee are shown in Fig. 17.13. As mentioned in Sec. 17.2, the total pressure of the main stream may increase when it f ows through a con verging or di verging wye or tee. This occurs because ener gy is recei ved from the airstream of higher v elocity or the di verging of the slo wly mo ving boundary layer into branch takeoff from the main airstream. However, the sum of the energies of the main and branch streams leaving the duct f tting is al ways smaller than that entering the duct f tting because of the ener gy losses. The magnitude of local loss coeff cients Cc,s, Cs,c, Cb,c, and Cc,b is affected by the shape and construction of the tee or wye, the v elocity ratios vs / vc and vb / vc, the v olume f ow ratios V˙s /V˙c and V˙b /V˙c, and area ratios As /Ac and Ab /Ac. For rectangular ducts, the converging or diverging wye of the con f guration shown in Fig. 17.13 gives the lower Co and, therefore, less energy loss. For round ducts, the Co values for various types of di verging tees and wyes with nearly the same outlet direction and the same v elocity ratio vb /vc  0.6, as shown in Fig. 17.12b, are quite different.

SH__ ST__ LG__ DF

Tee, diverging, round, with 45° elbow, branch 90° to main Tee, diverging, round, with 90° elbow, branch 90° to main Tee, diverging, round, with 45° elbow, conical branch 90° to main

Co  1.60 Co  1.18 Co  0.84

39445 Wang (MCGHP) Chap_17

SECOND PASS bzm 6/28/00

pg 17.37

AIR SYSTEMS: AIR DUCT DESIGN

FIGURE 17.14

17.37

Openings mounted on a duct or a duct wall: (a) entrances; (b) exits.

Wye, 45°, diverging, round, with 60° elbow, branch 90° to main Wye, 45°, diverging, round, with 60° elbow, conical branch 90° to main

Co  0.68 Co  0.52

The different Co values help in choosing the optimum duct f tting and balancing the total pressure between different paths of airf ow in an air duct system. In Table 17.11 are listed the local loss coef f cients of diverging tees and wyes. F or details, refer to ASHRAE Handbook Fundamentals (1997) and Idelchik (1986).

__SH __ST __LG DF

39445 Wang (MCGHP) Chap_17

17.38

SECOND PASS bzm 6/28/00

pg 17.38

CHAPTER SEVENTEEN

FIGURE 17.15

Enlargements and contractions: (a) abrupt enlargement; (b) sudden contraction.

Entrances, Exits, Enlargements, and Contractions Entrances and exits are the end openings mounted on a duct or a duct w all, as shown in Fig. 17.14. Because of the change in direction of the streamlines at the entrance, eddies and lar ge-scale turbulences develop along the duct w all when the airstream passes through the entrance. Generally , the total pressure drop pt of the airstream before it enters the entrance is ne gligible. A sharp-edge entrance may have a Co  0.9, whereas for an entrance f ush-mounted with the w all, Co reduces to 0.5. An entrance with a conical con verging bellmouth may further reduce Co to 0.4. If an entrance with a smooth converging bell mouth is installed, Co could be as low as 0.1. When the air f ows through a w all outlet or an e xit, f ow separation occurs along the surf ace of the vanes, and wakes are formed do wnstream, so there is a drop in total pressure. The velocity of the airstream reaches its maximum value at the vena contracta, where the cross section of air f ow is minimum and the static pressure is negative. The total pressure loss at the outlet always includes the velocity pressure of the discharge airstream. Various types of return inlets, such as grilles and louv ers, and supply outlets, such as diffusers, are discussed in Chapter 18. When air f ows through an enlar gement or a contraction, f ow separation occurs and produces eddies and large-scale turbulences after the enlargement, or before and after the contraction. Both cause a total pressure loss pt, as shown in Fig. 17.15. To reduce the energy loss, a gradual expansion, often called a dif fuser, or converging transition is preferred. An expansion with an including angle of enlargement   14°, as shown in Fig. 17.15a, is ideal. In actual practice,  may be from 14° to 45° because of limited space. For converging transitions, an including angle of 30° to 60° is usually used.

17.7 FLOW RESISTANCE SH__ ST__ LG__ DF

Flow resistance is a property of f uid f ow that measures the characteristics of a f ow passage which resist the f uid f ow in that passage with a corresponding total pressure loss at a speci f c volume

39445 Wang (MCGHP) Chap_17

SECOND PASS bzm 6/28/00

pg 17.39

AIR SYSTEMS: AIR DUCT DESIGN

FIGURE 17.16

17.39

Total pressure loss pt and f ow resistance R of a round duct section.

f ow rate. Consider a round duct section (Fig. 17.16), between two cross-sectional planes 1 and 2, with a diameter D, in ft (mm), and length L, in ft (m). The f uid f ow has a v elocity of v, in fpm (m / s), and a volume f ow rate of V˙ , in cfm (m 3 / s). From Eq. (17.37), the total pressure loss of the f uid f ow pt, in in. WC (Pa), between planes 1 and 2 can be calculated as pt  Cv f

L D

v 2

8Cv f L

c

gc 2D 5

  2g  

V˙ 2

(17.55)

where f  friction factor   density of f uid, lb / ft3 (kg / m3) gc  dimensional constant, 32.2 lbm ft / lbf s2 (for SI unit, gc  1) Cv  conversion constant Let R  8Cv fL / (gc  2D 5), so that total pressure loss becomes pt  RV˙ 2 R

pt V˙ 2

(17.56)

__SH __ST __LG DF

39445 Wang (MCGHP) Chap_17

17.40

SECOND PASS bzm 6/28/00

pg 17.40

CHAPTER SEVENTEEN

In Eq. (17.56), R is def ned as f ow resistance, which indicates the resistance to f uid f ow of this duct section, and is characterized by its speci f c total pressure loss and v olume f ow rate. Flo w resistance in I-P units is expressed as in. WC / (cfm)2 (in SI units, Pa  s2 / m6). For a given duct section, D and L are constants. In addition, the difference in the mean values of f and  at different volume f ow rates is small, so R can be considered a constant. The relationship between pt and V˙ of this section can be represented by a parabola whose vertex coincides with the origin. Differentiating Eq. (17.56) with respect to V˙ , then, gives dpt  2RV˙ dV˙

(17.57)

This means that the slope of an y segment of the curv e pt  R V˙ 2 is 2 R V˙ . Consequently, the greater the flo w resistance R, the steeper the slope of the characteristic curv e of this duct section. Because the f ow resistance R of a specif c duct section is related to V˙ by means of Eq. (17.56), the total pressure loss of the f uid f ow in this duct section at any V˙ can be calculated. Therefore, the total pressure loss pt curve of this duct section can be plotted at dif ferent V˙ values, and forms the pt-V˙ diagram. On the pt-V˙ diagram, each characteristic curv e represents a unique v alue of f ow resistance R. If a duct section contains additional dynamic losses Cdyv2 / (2gc), as mentioned in Sec. 17.6, the relationship between pt, R, and de V˙ f ned by Eq. (17.56) still holds, but the f ow resistance R is changed to R

8 (Cv fL /D  Cdy) gc 2D 4

(17.58)

where Cdy  local loss coeff cient.

Flow Resistances Connected in Series Consider a duct system that consists of se veral duct sections connected in series with air f owing inside the duct, as shown in Fig. 17.17 a. When these duct sections are connected in series, the volume f ow rate of air that f ows through each section must be the same. The total pressure loss between planes 1 and 4 is the sum of the total pressure losses of each duct section, i.e., pt  p12  p23      pn  (R 12  R 23      R n)V˙ 2  RsV˙ 2

(17.59)

where R12, R23, . . . , Rn  f ow resistances of duct sections 1, 2, . . . , n, in. WC/(cfm)2 (Pas2 /m6) p12, p23, . . . , pn  total pressure losses across f ow resistances R12, R23, . . . , Rn, in. WC (Pa) V˙ 12, V˙ 23, . . . , V˙ n  volume f ow rates of air f owing through resistances R12, R23, . . . , Rn, cfm (m3 / s) The f ow resistance of the duct system Rs is equal to the sum of indi vidual f ow resistances of duct sections connected in series, or R s  R 12  R 23      R n

SH__ ST__ LG__ DF

(17.60)

If the characteristic curv es of each indi vidual duct section with f ow resistances of R12, R23, . . . , Rn are plotted on a pt-V˙ diagram, then the characteristic curv e of the entire duct can be plotted either by using the relationship Rs  pt / V˙ 2 or by graphical method. Plotting by graphical method is done by dra wing a constant V˙ line. The points of pt on the characteristic curv e of the duct can be found by summing the indi vidual total pressure losses p12, p23, . . . , pn of individual duct sections 1, 2, . . . , n, as shown in Fig. 17.17a.

39445 Wang (MCGHP) Chap_17

SECOND PASS bzm 6/28/00

pg 17.41

AIR SYSTEMS: AIR DUCT DESIGN

FIGURE 17.17

17.41

Combination of f ow resistances: (a) connected in series; (b) connected in parallel.

Flow Resistances Connected in Parallel When a duct system has se veral sections connected in parallel, as shown in Fig. 17.17 b, the total volume f ow rate of this duct V˙ that f ows through cross-sectional planes 1 and 2 is the sum of the individual volume f ow rates V˙A, V˙B, . . . , V˙n of aech duct section, i.e., V˙  V˙A  V˙B      V˙n 

√

p12  RA

√

p12    RB

√

p12 Rn

(17.61)

where p12  total pressure loss between planes 1 and 2, in. WC (Pa) RA, RB, . . . , Rn  f ow resistance of duct sections A, B, . . . , n, in. WC/ (cfm)2 As for the entire duct, V˙  √p12 /R p. Here Rp represents the f ow resistance of the entire duct whose sections are in parallel connection. Then 1

√R p



1

√R A



1

√R B

  

1

√R n

(17.62)

__SH __ST __LG DF

39445 Wang (MCGHP) Chap_17

17.42

SECOND PASS bzm 6/28/00

pg 17.42

CHAPTER SEVENTEEN

If duct sections are connected in parallel, the total pressure loss across the common junctions 1 and 2 must be the same, or pt  p12

(17.63)

For two sections with f ow resistances RA and RB connected in parallel, the f ow resistance of the combination is then Rp 

R AR B R A  R B  2√R AR B

(17.64)

As in series connection, the characteristic curve of the duct with se gments connected in parallel can be plotted on the pt-V˙ diagram either by the relationship pt  RpV˙ 2 or by graphical method. In the graphical method, draw a constant- pt line; the total v olume f ow rate on the characteristic curve of the duct can be found by adding the v olume f ow rates of the individual sections, as shown in Fig. 17.17b.

Flow Resistance of a Y Connection When a round duct consists of a main duct section and tw o branches, as shown in Fig. 17.18, a Y-connection f ow circuit is formed. In a Y-connection f ow circuit, the volume f ow of the main duct section 01, denoted by V˙ 01, is the sum of the v olume f ow rates at branches 11  and 12, that is, V˙ 01  V˙ 11  V˙ 12. If the ambient air that surrounds the duct is at atmospheric pressure, then p11  p12 and p01  p02. 2 2 Because p11  R11 V˙ 11 ,  p12  R12 V˙ 12 , V˙11  V˙12 

1 1  √R11/R12 1 1  √R12/R11

V˙01  K 11V˙01 V˙01  K 12V˙01

(17.65)

Therefore, total pressure loss at ductwork 01 and 02 is calculated as p01  p01  p11  R 01V˙ 201  R 11V˙ 211  (R 01  R 11K 211)V˙ 201  R 01V˙ 201 Similarly,

(17.66) p02'  R 02'V˙ 201

(17.67)

For a gi ven duct, Cdy, L01, L11, L12, D01, D11, and D12 are constants. Air density  can be considered a constant. Although the friction factor f is a function of air velocity v and the volume f ow rate V˙ , at high Reynolds numbers f is least affected and can also be tak en as a constant. Therefore, R01 and R02 can be considered constants. Figure 17.18 also sho ws the characteristic curv e of a Y connection plotted on a pt-V˙ diagram.

SH__ ST__ LG__ DF

Flow Resistance of a Duct System Supply and return duct systems are composed of Y connections. If the ambient air is at atmospheric pressure, the total pressure loss of a supply duct system and its f ow characteristics, as shown in

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FIGURE 17.18

17.43

Flow resistance for a Y connection.

Fig. 17.19a, are given as p0n'  p01  p12      p(n1)n  R 01V˙ 201  R 12K 212V˙ 201      R (n1)nK (n1) nV˙ 201

and

 R 0nV˙ 201

(17.68)

R 0n  R 01  R 12K 212      R (n1)nK (n1)n

(17.69)

The characteristic curve of a duct system on a pt-V˙ diagram is called the system curve. It is shown in Fig. 17.19b.

17.8 PRINCIPLES AND CONSIDERATIONS IN AIR DUCT DESIGN Optimal Air Duct Design An optimal air duct system transports the required amount of conditioned, recirculated, or exhaust air to the specif c space and meets the following requirements:

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FIGURE 17.19

Flow characteristics of a supply duct system: (a) schematic diagram; (b) system curve.

pg 17.44

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●

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17.45

An optimal duct system layout within the allocated space A satisfactory system balance, achieved through the pressure balance of v arious paths by changing duct sizes or using different conf gurations of duct f ttings Space sound level lower than the allowable limits Optimum energy loss and initial cost Installation with only necessary balancing devices such as air dampers and nozzle plates National, ASHRAE, and local codes of f re protection, duct construction, and duct insulation met

These requirements result in the design of optimal duct layout, duct size, and total pressure loss of the duct system. Air duct system design often requires comprehensi ve analysis and computer -aided calculating programs. Dif ferent air duct systems ha ve dif ferent transport functions and thus ha ve their o wn characteristics. It is dif f cult to combine the in f uences of cost, system balance, and noise together with duct characteristics into one or two representative indices.

Design Velocity For any air duct system, the nearer the duct section is to the f an outlet or inlet, the greater its v olume f ow rate. The f an outlet or the main duct connected to the f an outlet is often the location where maximum air v elocity occurs. The maximum design air v elocity vd,max is determined mainly according to the space a vailable, noise control, energy use, and cost considerations. Return and branch ducts are nearer to the conditioned space and ha ve comparatively lower volume f ow rates than the supply ducts; therefore, supply main ducts often allo w a higher vd,max than return and branch ducts, except when there is a surplus pressure in a branch duct or a higher branch v elocity is required to produce a negative local loss coeff cient. For supply air duct systems in high-rise commercial b uildings, the maximum design air velocity in the supply duct vd,max is often determined by the space a vailable between the bottom of the beam and the suspended ceiling, as allocated by the architect, where the main duct tra verses under the beam. Because of the impact in recent years of ener gy-eff cient design of HV AC&R systems in commercial buildings (as in f uenced by ASHRAE Standards, local codes, and regulations) the following are true: ●

●

In supply main ducts vd,max usually does not e xceeds 3000 fpm (15 m / s). Airf ow noise must be checked at dampers, elbows, and branch takeoffs to satisfy the indoor NC range. In buildings with more demanding noise control criteria, such as hotels, apartments, and hospital wards, in supply main ducts usually vd,max 2000 to 2500 fpm (10 to 12.5 m / s), in return main ducts vd,max 1600 fpm (8 m / s), and in branch ducts vd,max 1200 fpm (6 m / s).

Higher air velocity results in a higher ener gy cost, and lower air velocity increases the material and labor costs of the installation. If a commercial b uilding has suf f cient headroom in its ceiling plenum, or if an industrial application has enough space at a higher level, an optimization procedure can then be used to reach a compromise between ener gy and installation costs. A check of air f ow noise is still necessary. For a particulate-transporting duct system, the air velocity must be higher than a speci f c value at any section of the duct system to f oat and transport the particulates. Design v elocities for the components in an air duct system are listed in Table 17.9. The face velocity vfc, in fpm (m / s), is def ned as vfc 

V˙ Ag, fc

where V˙  air volume f ow rate f owing through component, cfm (m3 / s) Ag, fc  gross face area of component perpendicular to airf ow, width height, ft2 (m2)

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TABLE 17.9 Design Velocities for Air-Handling System Components Duct element

Face velocity, fpm

Louvers Intake 1. 7000 cfm and greater 2. 2000 to 7000 cfm

400 250 to 400

Exhaust 1. 5000 cfm and greater 2. Less than 5000 cfm

500 300 to 400

Filters Panel f lters 1. Viscous impingement 2. Dry-type extended-surface Flat (low eff ciency) Pleated media (medium eff ciency) HEPA

Duct velocity up to 750 250

Renewable media f lters 1. Moving-curtain viscous impingement 2. Moving-curtain dry-media

500 200

Electronic air cleaners, ionizing-type

150 to 350

Heating coils Steam and hot water

200 to 800

400 to 600 (200 min., 1500 max.)

Dehumidifying coils

500 to 600

Air washers Spray type High-velocity spray type

300 to 700 1200 to 1800

Source: Adapted with permission from ASHRAE Handbook 1989, Fundamentals.

These face velocities are recommended based on the ef fectiveness in operation of the system component and its optimum pressure loss.

System Balancing

SH__ ST__ LG__ DF

For an air duct system, system balancing means that the v olume flow rate that comes from each outlet or flo ws into each inlet should be (1) equal or nearly equal to the design v alue for constant-volume systems and (2) equal to the predetermined values at maximum and minimum flow for v ariable-air-volume systems. System balancing is one of the primary requirements in air duct design. F or supply duct systems installed in commercial b uildings, using dampers only to provide design airflow often causes additional air leakage, as well as an increase in installation cost, and in some cases objectional noise. Therefore, system balancing using dampers only is not recommended. A typical small supply duct system in which pt of the conditioned space is equal to zero is shown in Fig. 17.20. F or such a supply duct system, the pressure and v olume f ow characteristics are as follows:

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FIGURE 17.20 System balancing and critical path of a supply duct system.

●

From node 1, the total pressure loss of the airstreams f owing through v arious paths 1-3, 1-2-4, and 1-2-5 to the conditioned space is always equal, i.e., pt,1-3  pt, 1-2-4  pt, 1-2-5

●

(17.71)

Volume f ow rates V˙1 , V˙2 , and V˙3 supplied from the branch tak eoffs, 1-3, 2-4, and 2-5 depend on the size and the con f guration of the duct f ttings and duct sections in the main duct and in the branch takeoffs as well as the characteritics of the supply outlet.

The relationship between the total pressure loss of an y duct section pt, the volume f ow rate of the duct section V˙ , and the f ow resistance R can be expressed as follows: pt  RV˙ 2

(17.72)

For duct paths 1-3 and 1-2-45, pt,1-3  pt, 1-2-45  R 1-3V˙21  R 1-2-45(V˙2  V˙3)2 Here, path 1-2-45 indicates the airf ow having volume f ow rate V˙2  V˙3 and f owing through nodes 1, 2 and a combined parallel path 2-4 and 2-5. At design conditions, the f ow resistance R1-2-45 is determined in such a manner that the total pressure loss pt,1-2-45 along path 1-2-45 at a v olume f ow rate of V˙2  V˙3 is balanced with the total pressure loss pt,1-3. ●

●

●

If the diverging wye or tee, terminal, duct f ttings, and duct section used in path 1-3 are similar to those of paths 2-4 and 2-5, most probably V˙1 will be greater than the required design v alue as the result of lower R1-3. To have required V˙1 f owing through path 1-3 at design conditions to pro vide a system balance, the following means are needed: (1) Decrease R1-2-45, including an increase in the size of duct section 1-2; (2) increase R1-3, mainly by using a smaller duct in section 1-3 and a diverging wye or tee of greater local loss coeff cient Cc,b. In a duct system, if the f ow resistance of each branch duct Rb is greater, the duct system is more easily adjusted to achie ve an equal amount of supply v olume f ow from each of the branch supply outlets. For the branch tak eoff or connecting duct ha ving a length less than 1 to 2 ft (0.3 to 0.6 m) long, it is often diff cult to increase Rb by reducing its size. Adding a volume damper directly on the supply outlet may alter space air f ow patterns. A better remedy is to v ary the sizes of the successi ve main duct sections with outlets of greater f ow resistance to achieve a better system balance. A variable-air-volume duct system installed with a VAV box in each branch tak eoff does provide system balancing automatically. However, only a part of the modulating capacity of a VAV box is allowed to be used to pro vide system balance of a speci f c branch tak eoff (such as less than 20 percent), so that the quality of its modulation control at part-load operation is not impaired.

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Critical Path For any air duct system there e xists a critical path, or design path of air f ow, whose total f ow resistance Rdgn is maximum compared with other air f ow paths when the v olume f ow rate of the critical path is equal to the design value at design conditions. A critical path is usually a duct path with more duct f ttings and comparati vely higher v olume f ow rate; additionally, it may be the longest one. In Fig. 17.20, path FO-1-2-5 may be the critical path. For an energy-eff cient air duct design, the total pressure loss, including the local loss coef f cients of the duct f ttings along the critical path, should always be minimized, especially at the f an inlet and outlet or their vicinity (see Chap. 21). To reduce the dynamic losses of the critical path, the following are recommended: ●

●

●

●

Maintain an optimum air velocity for airf ow through the duct f ttings. Emphasize the reduction of dynamic losses of duct f ttings that are nearer to the fan outlet or inlet with a high air v elocity, especially the f an system ef fects. F an system ef fects are discussed in Chap. 21. Evans and Tsal (1996) recommend use of 90 ° elbows with an Rc /D or Rc /W of 1.5. If space is not available, use one or tw o splitter v anes in the elbo ws with throat radius or Rc /W  0.1; the throat radius should not be smaller than 4 in. (100 mm). Turning vanes should not be used in a transitional elbow or an elbow other than 90°. Proper turning vane installation is critical to performance, which favors a factory-made unit. Set two duct f ttings as far apart as allowable; if they are too close together , the eddies and largescale turbulences of the f rst duct f tting often af fect the v elocity distribution in the second duct f tting and considerably increase the pressure loss in the second f tting.

For other duct paths, if the total pressure loss at the design f ow rate is smaller than the pressure loss available, a smaller duct size and duct f ttings of greater local loss coef f cients may pro vide a better balance.

Air Leakage Conditioned air leakage from the joints and seams of the air duct to the space, which is not air conditioned, is always a w aste of refrigeration or heating ener gy as well as f an power. Based on their tested results, Swim and Griggs (1995) reported that the joints were the major leakage sites and accounted for 62 to 90 percent of the total air leakage from joints and seams. Air leakage depends mainly on the use of sealant on joints and seams, the quality of f abrication, and the shape of the ducts. Heat-sensiti ve tapes, mastic and glass f abric, and man y other materials are used as the sealants for the joints and seams. Duct leakage classi f cations, based on tests conducted by AISI, SMACNA, ASHRAE, and the Thermal Insulation Manufacturers Association (TIMA), are presented in Table 17.10. The air leakage rate ,V˙L in cfm / 100 ft2 (L / s per m 2) of duct surf ace area can be calculated by the follo wing formula: V˙L  CL p 0.65 sd

(17.73)

The leakage class can then be calculated as CL 

SH__ ST__ LG__ DF

V˙L p 0.65 sd

(17.73a)

where CL  leakage class, cfm per 100-ft2 duct surface area at a static pressure difference of 1 in. WC [based on Eq. (17.73a)] C  constant affected by area characteristics of leakage path psd  static pressure differential between air inside and outside duct, in. WC (Pa)

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TABLE 17.10 Duct Leakage Classif cation Predicted leakage class CL Type of duct

Sealed

Metal (f exible excluded) Round and oval Rectangular 2 in. WC (both positive and negative pressures)

2 and 10 in. WC (both positive and negative pressures) Flexible Metal, aluminum

Unsealed

3

30 (6 to 70)

12

48 (12 to 110)

6

48 (12 to 110)

8

30 (12 to 54) 30 (4 to 54)

Nonmetal

12

Fibrous glass Rectangular Round

6 3

NA NA

The leakage classes listed in this table are averages based on tests conducted by AISA / SMACNA 1972, ASHRAE / SMACNA / TIMA 1985, and ASHRAE 1988. Leakage classes listed are not necessarily recommendations on allowable leakage. The designer should determine allowable leakage and specify acceptable duct leakage classif cations. Source: ASHRAE Handbook 1989, Fundamentals. Reprinted with permission.

ASHRAE / IESNA Standard 90.1-1999 mandates that ductw ork and plenum shall be sealed in accordance with the following requirements: Supply duct Location

2 in. WC

2 in. WC

Exhaust duct

Return duct

Outdoors Unconditioned spaces Conditioned spaces

A B C

A A B

C C B

A B C

The sealing requirement of sealing levels A, B, and C are as follows: A  All transverse joints, longitudinal seams, and duct wall penetrations. Pressure-sensitive tape shall not be used as the primary sealant.

B  All transverse joints, longitudinal seams. Pressure-sensitive tape shall not be used as the primary sealant. C  Transverse joints only.

Longitudinal seams are joints oriented in the direction of air f ow. Transverse joints are connections of two duct sections oriented perpendicular to air f ow. Duct w all penetrations are openings made by any screw fastener, pipe, rode, or wire. Spiral lock seams in round and f at oval duct need not be sealed. All other connections are considered transverse joints.

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ASHRAE / IESNA Standard 90.1-1999 also mandates that: ●

●

●

Metal round and f at oval ducts should have a leaking class C L3, and metal rectangular ducts, rectangular f brous ducts, and round f exible ducts have a CL6. Ducts that are designed to operate at a static pressure e xceeding 3 in. WC (750 Pa) shall be leak tested according to industry accepted test procedures. Representati ve sections totaling no less than 25 percent of total installed duct area for the designed pressure class shall be tested. The maximum permitted duct leakage V˙L, in cfm / 100 ft 2 (L / sm2), shall be calculated from Eq. (17.73).

Shapes and Material of Air Ducts When a designer chooses the shape (round, rectangular, or f at o val duct) or material (galv anized sheet, aluminum, f berglass, or other materials) of an air duct, the choices depend mainly on the space available, noise, cost, local customs and union agreements, experience, quality, and the requirements of the project. In many high-rise commercial buildings, factory-fabricated round ducts and sometimes f at oval ducts with spiral seams are used because the y have fewer sound problems, lower air leakage, and many conf gurations of wyes and tees available for easier pressure balance. Round ducts also have the advantage of high breakout transmission loss at low frequencies (see Chap. 19). For ducts running inside the air conditioned space in industrial applications, metal rectangular ducts are often chosen for their lar ge cross-sectional areas and con venient fabrication. Round ducts are often used for more demanding projects. In projects designed for lo wer cost, adequate duct insulation, and sound attenuation, f berglass ducts may sometimes be the optimum selection. Ductwork Installation Ductwork installation, workmanship, materials, and methods must be monitored at all stages of the design and construction process to ensure that they meet the design intent. Fire Protection The design of air duct systems must meet the requirements of National Fire Codes NFP A 90A, Standards for the Installation of Air Conditioning and Ventilating Systems, Warm Air Heating and Air Conditioning Systems, and Blower and Exhaust Systems as well as local codes. Refer to these standards for details. The following are some of the requirements: ●

●

●

●

●

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The duct material discussed in Sec. 17.3 must be made of class 0 or class 1 material. Also the duct coverings and linings — including adhesive, insulation, banding, coating, and f lm covering the outside surf ace and material lining the inside surf ace of the duct — must have a f ame spread rating not over 25 and a smoke development rating not over 50 except for ducts outside buildings. Supply ducts that are completely encased in a concrete f oor slab not less than 2 in. (50 mm) thick need not meet the class 0 or class 1 requirement. Vertical ducts more than tw o stories high must be constructed of masonry , concrete, or clay tile. When ducts pass through the f oors of buildings, the vertical openings must be enclosed with partitions and walls with a f re protection rating of not less than 1 h in buildings less than four stories high and greater than 2 h in buildings four stories and higher. Clearances between the ducts and comb ustible construction and material must be made as specif ed in NFPA 90A. The opening through a f re wall by the duct system must be protected by (1) a f re damper closing automatically within the f re wall and having a f re protection rating of not less than 3 h or (2) f re doors on the tw o sides of the f re wall. A service opening must be pro vided in ducts adjacent to

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each f re damper. Many regulatory agencies have very rigid requirements for f re dampers, smoke dampers, f re / smoke dampers in combination, and smoke venting. When a duct penetrates through w alls, f oors, and partitions, the gap between the ducts and the walls, f oors, and partitions must be f lled with noncomb ustible material to pre vent the spread of f ames and smoke. Duct systems for transporting products, vapor, or dust in industrial applications must be constructed entirely of metal or noncomb ustible material. Longitudinal seams must be continuously welded, lapped and ri veted, or spot-welded on maximum centers of 3 in. (75 mm). Transitions must be 5 in. (125 mm) long for e very 1-in. (25-mm) change in diameter . Rectangular ducts may be used only when the space is not available and must be made as square as possible.

17.9 AIR DUCT DESIGN PROCEDURE AND DUCT LAYOUT Design Procedure Before an air duct system is designed, the supply v olume f ow rate for each conditioned space, room, or zone should be calculated, and the locations of the supply outlets and return inlets should also be settled according to the requirements of space air dif fusion (see Chap. 18). F or an air duct system, the supply volume f ow rate of cold supply air in summer is usually greater than the w arm volume f ow rate needed in winter. If an air duct system conditions the space with cold air supply in summer, it often also conditions the space with warm air supply in winter. Computer-aided duct design and sizing programs are widely used for more precise calculation and optimum sizing of large and more complicated duct systems. Manual air duct design and sizing are often limited to small and simple duct systems. Computer -aided duct design and sizing programs are discussed in a later section in this chapter. The design procedure for an air duct system is as follows: 1. Designer should verify local customs, local codes, local union agreements, and material availability constraints before proceeding with duct design. 2. The designer proposes a preliminary duct layout to connect the supply outlets and return inlets with the f an(s) and other system components through the main ducts and branch tak eoffs. The shape of the air duct is selected. Space a vailable under the beam often determines the shape of the duct and affects the layout in high-rise buildings. 3. The duct layout is divided into consecutive duct sections, which converge and diverge at nodes or junctions. In a duct layout, a node or junction is represented by a cross-sectional plane perpendicular to air f ow. The volume f ow rate of an y of the cross sections perpendicular to air f ow in a duct section remains constant. A duct section may contain one or more duct se gments (including duct f ttings). A duct system should be divided at a node or junction where the airf ow rate changes. 4. The local loss coeff cients of the duct f ttings along the tentative critical path should be minimized, especially adjacent to fan inlets and outlets. 5. Duct sizing methods should be selected according to the characteristics of the air duct system. The maximum design air v elocity is determined based on the space a vailable, noise, energy use, and initial cost of the duct system. Various duct sections along the tentati ve critical path are sized. 6. The total pressure loss of the tentative critical path as well as the air duct system is calculated. 7. The designer sizes the branch ducts and balances the total pressure at each junction of the duct system by varying the duct and component sizes, and the conf guration of the duct f ttings. 8. The supply volume f ow rates are adjusted according to the duct heat gain at each supply outlet. 9. The designer resizes the duct sections, recalculates the total pressure loss, and balances the parallel paths from each junction.

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10. The airborne and breakout sound level from various paths should be check ed and the necessary attenuation added to meet requirements.

Duct System Characteristics Air duct systems can be classi f ed into the follo wing three cate gories according to their system characteristics: 1. Supply duct, return duct, or exhaust duct systems with a certain pressure loss in branch tak eoffs 2. Supply duct, return duct, or exhaust duct systems in which supply outlets or return grilles either are mounted directly on the duct or ha ve only v ery short connecting duct between the outlet or inlet and the main duct 3. Industrial exhaust duct systems to transport dust particulates or other particulate products

Duct Layout When a designer starts to sk etch a preliminary duct layout using computer -aided design and drafting (CADD) or manually , the size of the air duct system (the conditioned space serv ed by the air duct system) must be decided f rst. The size of an air duct system must be consistent with the size of the air system or e ven the air conditioning system. From the point of vie w of the air duct system itself, a smaller and shorter system requires less power consumption by the fan and shows a smaller duct heat gain or loss. The air duct system is also comparatively easier to balance and to operate. If the designer uses a more symmetric layout, as shown in Fig. 17.21, a more direct and simpler form for the critical path can generally be deri ved. A symmetric layout usually has a smaller main duct and a shorter design path; it is easier to pro vide system balance for a symmetric than a nonsymmetric layout. A more direct and simpler form of critical path usually means a lo wer total pressure loss of the duct system. For variable-air-volume air duct systems, the ends of the main ducts are connected to each other to form one or more duct loops, as shown by the dotted lines in Fig. 17.21. Duct looping(s) allo w some of the duct sections to be fed from the opposition direction. Balance points e xist where the total pressure of the opposite f owing airstreams is zero. The positions of the balance points often follo w the sun’s position and the induced cooling load at various zones during the operating period. Duct looping optimizes transporting capacity and results in a smaller main duct than without looping.

SH__ ST__ LG__ DF

FIGURE 17.21 A typical supply duct system with symmetric duct layout (bold line) and duct loopings (connected by dotted lines) for a typical f oor in a high-rise building.

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The designer then compares v arious alternative layouts and reduces the number of duct f ttings, especially the f ttings with higher velocity and high local loss coef f cients along the critical path, in order to achieve a duct system with lower pressure loss. When duct systems are installed in commercial and public b uildings without suspended ceiling, duct runs should be closely matched with the b uilding structures and give a neat and harmonious appearance.

17.10 DUCT SIZING METHODS Duct sizing determines the dimensions of each duct section in the air duct system. After the duct sections have been sized, the total pressure loss of the air duct system can then be calculated, and the supply, return or relief f an total pressure can be calculated from the total pressure losses of the supply and return duct systems and the pressure loss in the air-handling unit or packaged unit. Four duct-sizing methods are currently used: 1. 2. 3. 4.

Equal-friction method with maximum velocity Constant-velocity method Static regain method T method

Equal-Friction Method This method sizes the air duct so that the duct friction loss per unit length pf,u at various duct sections always remains constant. The f nal dimensions of sized ducts should be rounded to standard size. The total pressure loss of the duct system pt, in in. WC (Pa), equals the sum of the frictional losses and dynamic losses at various duct sections along the critical path: pt  pf,u[(L 1  L 2      L n)  (L e1  L e2      L en)]

(17.74)

where L1, L2, . . . , Ln  length of duct sections 1, 2, . . . , n, ft (m) Le1, Le2, . . . , Len  equivalent length of duct f ttings in duct sections 1, 2, . . . , n, ft (m) If the dynamic loss of a duct f tting is equal to the friction loss of a duct section of length Le, in ft (m), then Cov 2 2gc



f(L e / D)v 2 2gc

and the equivalent length is

Co D (17.75) f The selection of pf,u is usually based on experience, such as 0.1 in. WC per 100 ft (0.82 Pa / m) for low-pressure systems. A maximum velocity is often used as the upper limit. The equal-friction method does not aim at an optimal cost. Dampers are sometimes necessary for a system balance. Because of its simple calculations, the equal-friction method is still used in many low-pressure systems in which airborne noise due to higher air v elocity is not a problem or for small duct systems. Le 

Constant-Velocity Method The constant-velocity method is often used for e xhaust systems that convey dust particles in industrial applications. This method f rst determines the minimum air v elocity at v arious duct sections

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according to the requirement to f oat the particles, either by calculation or by experience. On the basis of the determined air v elocity, the cross-sectional area and, therefore, the dimension of the duct can be estimated and then rounded to a standard size. The total pressure loss of the duct system pt, in in. WC (Pa), along the critical path can be calculated as pt 

K 2gc

fL  D

1 1

  

1



 C1 v 21  fnL n

D

n

f2L 2

D



 Cn v 2n

2



 C2 v 22 (17.76)

where v1, v2, . . . , vn  mean air velocity at duct sections 1, 2, . . . , n, respectively, fpm (m / s) C1, C2, . . . , Cn  local loss coeff cients at duct sections 1, 2, . . . , n, respectively K  5.35 105 for I-P unit (1 for SI unit) Static Regain Method This method sizes the air duct so that the increase of static pressure (static re gain) due to the reduction of air velocity in the supply main duct after each branch tak eoff nearly offsets the pressure loss of the succeeding duct section along the main duct. As a consequence, the static pressure at the common end of the di verging tee or wye of the sized duct section remains approximately the same as that of the preceding section. A rectangular duct section 1-2 between the cross-sectional planes 1 and 2 is illustrated in Fig. 17.22. The size of this duct section is to be determined. Let v1 and v2 be the mean velocities at

SH__ ST__ LG__ DF

FIGURE 17.22

Pressure characteristics of a main duct section.

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planes 1 and 2, V˙1 and V˙2 the volume f ow rates, and A1 and A2 the cross-sectional areas. The total pressure loss in duct section 1-2 consists of the duct friction loss pf1-2 and the dynamic loss of the main airstream f owing through the diverging tee p1c,s. The relationship between the total pressure at planes 1 and 2 can be expressed as pt1  pt2  pf1-2  p1 c,s

(17.77)

because pt  ps  pv. Ignore the dif ference between air densities 1 and 2. Let pf1-2  pf,uL1-2. Here L1-2 represents the length of the duct section 1-2, in ft (m). If the static pressures at planes 1 and 2 are equal, that is, ps1  ps2, then

1(v 21  v 22) 2gc

 pf,uL 1-2 

C1 c,s1v 21

(17.78)

2gc

If v is e xpressed in fpm and pf,u in in. WC per 100 ft, lbm ft / lbf s2, the mean air velocity of the sized duct section is

  0.075 lb / ft3, and gc  32.2

v2  [(1  C1 c,s) v 21  1.6 105pf,uL 1-2]0.5

(17.79)

In SI units, v2 



(1  C1 c,s)v 21  2pf,uL 1-2



0.5



(17.79a)

with v1 and v2 both in m / s, pf,u in Pa / m, L1-2 in m, and  in kg / m3. For any duct section between cross-sectional planes n  1 and n, if the total local loss coef f cient of the duct f ttings is Cn, excluding the local loss coef f cient C(n1)c,s, the mean air v elocity of the sized duct section is vn 



[(1  C(n1) c,s)v 2n1  1.6 105 pf,uL n] 1  Cn

0.5



(17.80)

Because vn1, n, Ln, and C n are known values, by using iteration methods, vn can be determined. The dimension of the duct section and its rounded standard size can also be determined. The static regain method can be applied only to supply duct systems. It tends to produce a more even static pressure at the common end of each di verging tee or wye leading to the corresponding branch takeoff, which is helpful to the system balance. It does not consider cost optimization. The main duct sections remote from the f an discharge often ha ve larger dimensions than those in the equal-friction method. Sound le vel and space required should be check ed against determined air velocity and dimension. When one is using the static regain method to size air ducts, it is not recommended to allow only part of static regain to be used in the calculation.

T Method The T method, f rst introduced by Tsal et al. (1988), is an optimizing procedure to size air ducts by minimizing their life-c ycle cost. It is based on the tree-staging idea and is therefore called the T method. The goal of this method is to optimize the ratio between the v elocities in all sections of the duct system. The T method consists of the following procedures: 1. System condensing — condensing various duct sections of a duct system into a single imaginary duct section ha ving the same hydraulic characteristics and installation costs as the duct system

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39445 Wang (MCGHP) Chap_17

17.56

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pg 17.56

CHAPTER SEVENTEEN

2. Fan selection — selecting a fan that provides the optimum system pressure loss 3. System expansion — expanding the imaginary duct section into the original duct system before condensing with the optimum distrib ution of total pressure loss between v arious duct sections During optimization, local loss coef f cients are considered constant at v arious stages of iteration. For details, refer to Tsal et al. (1988). The T method can be used for sizing duct systems with certain total pressure losses in the branch ducts. Ho wever, the local loss coef f cients are actually v aried at v arious stages of the iteration and should be taken into consideration during optimization.

17.11 DUCT SYSTEMS WITH CERTAIN PRESSURE LOSSES IN BRANCH TAKEOFFS Design Characteristics Supply, return, or exhaust duct systems with certain pressure losses in branch tak eoffs have the following characteristics: ●

●

●

●

●

Duct is sized based on the optimization of the life-c ycle cost of v arious duct sections of the duct system as well as the space available in the building. System balancing is achie ved mainly through pressure balancing of v arious duct paths by changing of duct sizes and the use of v arious con f gurations of duct f ttings and terminals instead of dampers or other devices. Sound level will be check ed and analyzed. Excess pressure at each inlet of VAV box at design conditions is to be avoided. Sound attenuation arrangements are added if necessary. Local loss coeff cients of the duct f ttings and equipment along the critical path are minimized. It may be bene f cial to use the surplus pressure a vailable in the branch tak eoff to produce a higher branch duct velocity and a smaller straight-through local loss coeff cient Cs,c. Supply volume f ow rates are adjusted according to the duct heat gain. For VAV systems, diversity factors are used to determine the v olume f ow rate of various duct sections along the critical path so that the volume f ow rate nearly matches the block load at the fan discharge.

Cost Optimization For any duct section in an air duct system, the total life-c ycle cost Cto, as shown in Fig. 17.23, in dollars, can be calculated as Cto  Ce

1  Cdi CRF

(17.81)

In Eq. (17.81), CRF indicates the capital recovery factor and can be calculated as follows: CRF 

SH__ ST__ LG__ DF

i(1  i)n (1  i)n  1

(17.82)

where i  interest rate and n  number of years under consideration. The f rst-year ener gy cost Ce, in dollars, can be calculated as the product of electric ener gy consumed at the fan and the unit ener gy cost Er, in $ / kWh, times the annual operating hours tan, in

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pg 17.57

AIR SYSTEMS: AIR DUCT DESIGN

17.57

FIGURE 17.23 Cost analysis for a duct system.

h, or in I-P units Ce  pt

0.746 V˙E rt an



 pt

6350fm 1.175 10 4V˙E rt an

fm

  K p e

t

(17.83)

For SI units, Ke 

V˙E rt an 1000fm

(17.84)

where pt  total pressure loss of duct section, in. WC (Pa) V˙  volume f ow rate of duct section, cfm (m3 / s) f, m  total eff ciency of fan and eff ciency of motor Let z1  Ke(1 / CRF). Then we can write 1 C  z 1 pt CRF e

(17.85)

Installation cost for round duct Cdi, in dollars, can be calculated as Cdi  DLCiu where D  diameter of round duct, ft (m) L  length of duct section, ft (m) Ciu  unit cost of duct installation, $ / ft2 ($ / m2)

(17.86) __SH __ST __LG DF

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39445 Wang (MCGHP) Chap_17

17.58

pg 17.58

CHAPTER SEVENTEEN

The surface area of rectangular duct Arec  2(W  H)L. Here W is the width of the duct, and H is the height of the duct, both in ft (m). Then for rectangular duct 2(W  H )

Cdi  2(W  H )LCiu 

D

DLCiu  R recDLCiu

(17.87)

where Rrec  ratio of surf ace area of rectangular duct to surf ace area of round duct. F or any duct section, the total pressure loss pt, in in. WC (Pa), is 5.35 10 5( fL/D  C )v 2

pt 

(17.88)

2gc

where C  sum of local loss coeff cients in duct section, based on air velocity of sized duct section v  mean air velocity in sized duct section, fpm (m / s) For frictional loss D  f (p0.2 ), and for dynamic loss D  f(p0.25 ). However, for simplicity, let t t 0.22 us take D  f(pt ) and consider ( fL / √D)  C√D a constant. Then fL / √D  C√D)

  0.1097  g

D  0.1097

gc

0.22

K

c

0.22



V˙ 0.44 p 0.22 t

p 0.22 t L

(17.89)

Because gc  32.2 lbm ft / lbf  s2, fL  √D  C√D

0.22

D  0.0511

V˙ 0.44 p 0.22 t

(17.90)

where K  ( fL/√D  C√D)0.22V˙ 0.44L. Let z2  0.1097 ( / gc)0.22 Ciu. By substituting into Eq. (17.86), for a round duct, we have Cdi  DLCiu  z 2K p t0.22

(17.91)

Cdi  R recz 2K p 0.22 t

(17.92)

For a rectangular duct

Seeking a minimum cost by taking derivatives of Eq. (17.81) with respect to pt, making it equal to zero, we have dCto  z 1  0.22z 2K p t1.22  0 dpt Because gc  32.2 lbm ft / lbf  s2, then for a round duct LCiufm

fL (1 / CRF)E t   √D  C√D  V˙ 1 

pt  108

0.82

0.18

0.46

r an

(17.93)

For a rectangular duct, SH__ ST__ LG__ DF

LCiufm

fL (1 / CRF)E t   √D  C √D  V˙ 1 

pt  108R rec

r an

0.82

0.18

0.46

(17.94)

39445 Wang (MCGHP) Chap_17

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pg 17.59

AIR SYSTEMS: AIR DUCT DESIGN

17.59

When the total pressure loss of the duct section that has the minimum cost is calculated from Eq. (17.93) or (17.94), the diameter or the circular equi valent of the duct section of minimum cost can then be calculated by Eq. (17.90). Condensing Two Duct Sections According to the T method, if two duct sections 1 and 2 connected in series are condensed into an imaginary section 1-2, as shown in Fig. 17.24a, then the volume f ow rate at each duct section must be equal, i.e., V˙1  V˙2  V˙1-2

(17.95)

Here subscripts 1 and 2 and 1-2 represent the duct sections 1 and 2 and the imaginary duct section 1-2, respectively. The total pressure loss of the condensed duct section must equal the sum of the total pressure losses of duct sections 1 and 2, or pt1  pt2  pt1-2

(17.96)

and the installation cost of the condensed section is Cdi1-2  Cdi1  Cdi2 0.22  z 2(K 1 p t10.22  K 2 p t20.22)  z 2K 1-2 p t1-2

(17.97)

When two duct sections 1 and 2 connected in parallel are condensed into an imaginary duct section 1-2, as shown in Fig. 17.24b, the following relationships hold: V˙1-2  V˙1  V˙2 pt1-2  pt1  pt2

(17.98)

Cdi1-2  Cdi1  Cdi2

FIGURE 17.24 Two duct sections 1 and 2 condensed into an imaginary duct section 1-2: (a) two duct sections 1 and 2 connected in series; (b) two duct sections 1 and 2 connected in parallel.

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39445 Wang (MCGHP) Chap_17

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pg 17.60

CHAPTER SEVENTEEN

If the whole duct system is condensed into one imaginary duct section, its minimum cost can be found by taking derivatives with respect to pt of the imaginary duct section. In using such a procedure, the minimum cost includes both the main and branch ducts. However, if the local loss coefficients of the terminals and con verging and diverging wyes are assumed constant, the benefits of using smaller terminals and dif ferent configuration of wyes to balance the branch duct paths are lost, as demonstrated by Dean et al. (1985). A more flexible and often economical alternative is that only the sizes of duct sections along the critical path of a duct system are determined according to the cost optimization procedure — that is, Eqs. (17.93), (17.94), and (17.90) — and rounded to standard size. The sizes of the duct sections of other duct paths, such as branch tak eoffs, should be determined according to the dif ference between the total pressure at the junction and the space pressure for system or pressure balance using optimum terminals, wyes, and fittings.

Local Loss Coefficients for Diverging Tees and Wyes For a supply duct system, there are often more diverging tees or wyes than elbows along the design path. Selecting di verging tees or wyes with smaller local loss coef f cients Cc,s has a de f nite inf uence on the design of a duct system with minimizing total pressure loss of critical path. Table 17.11 sho ws Cc,s values for di verging wyes and tees for round ducts and di verging wyes for rectangular ducts. The follo wing are recommendations for proper selection of wyes and tees: ●

Select a diverging wye instead of a diverging tee except when there is surplus total pressure at the branch takeoffs. A 30° angle between the branch and the main duct has a smaller Cc,s value than a 45° or a 60° angle.

TABLE 17.11 Local Loss Coeff cients for Elbows, Diverging Tees, and Diverging Wyes (1) Elbow; 3, 4, and 5 pieces, round

Coeff cients for 90° elbows Co

SH__ ST__ LG__ DF

Rc /D

No. of pieces

0.75

1.0

1.5

2.0

5 4 3

0.46 0.50 0.54

0.33 0.37 0.42

0.24 0.27 0.34

0.19 0.24 0.33

Angle correction factors K (Idelchik, 1986; diagram 6-1)

 K

0 0

20 0.31

30 0.45

45 0.60

60 0.78

75 0.90

90 1.00

110 1.13

130 1.20

150 1.28

180 1.40

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39445 Wang (MCGHP) Chap_17

pg 17.61

TABLE 17.11 Local Loss Coeff cients for Elbows, Diverging Tees, and Diverging Wyes (Continued ) (2) Elbow, smooth radius with splitter vanes, rectangular

Coeff cients for elbows with two splitter vanes Co H/W R / W Rc / W 0.05 0.10 0.15 0.20 0.25 0.30

0.55 0.60 0.65 0.70 0.75 0.80

0.25

0.5

1.0

1.5

2.0

0.362 0.26 0.450 0.17 0.507 0.12 0.550 0.09 0.585 0.08 0.613 0.06

CR

0.20 0.13 0.09 0.07 0.05 0.04

0.22 0.11 0.08 0.06 0.04 0.03

0.25 0.12 0.08 0.05 0.04 0.03

0.28 0.33 0.13 0.15 0.08 0.09 0.06 0.06 0.04 0.04 0.03 0.03

3.0

5.0

6.0

7.0

8.0

0.37 0.41 0.16 0.17 0.10 0.10 0.06 0.06 0.05 0.05 0.03 0.03

4.0

0.45 0.19 0.11 0.07 0.05 0.04

0.48 0.20 0.11 0.07 0.05 0.04

0.51 0.21 0.11 0.07 0.05 0.04

(3) Wye, 45°, round, with 60° elbow, branch 90° to main

Branch

b /c Cc,b

0 1.0

0.2 0.88

0.4 0.77

0.6 0.68

0.8 0.65

s /c Cc,s

0 0.40

0.1 0.32

0.2 0.26

0.3 0.20

0.4 0.14

1.0 0.69

1.2 0.73

1.4 0.88

1.6 1.14

0.6 0.06

0.8 0.02

1.0 0

1.8 1.54

2.0 2.2

Main 0.5 0.10

(4) Tee, diverging, round, with 45° elbow, branch 90° to main

__SH __ST __LG 17.61

DF

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39445 Wang (MCGHP) Chap_17

pg 17.62

TABLE 17.11 Local Loss Coeff cients for Elbows, Diverging Tees, and Diverging Wyes (Continued ) Branch Vh /Vc Cc,b

0 1.0

0.2 1.32

0.4 1.51

0.6 1.60

0.8 1.65

1.0 1.74

1.2 1.87

1.4 2.0

1.6 2.2

1.8 2.5

2.0 2.7

For main local loss coeff cient Cc,s, see values in (3) (5) Tee, diverging, rectangular (Idelchik, 1986; diagram 7-21)

Branch, Cc,b V˙b / V˙c Ab /As

Ab /Ac

0.1

0.2

0.3

0.4

0.5

0.6

0.7

0.8

0.9

0.25 0.33 0.5 0.67 1.0 1.0 1.33 2.0

0.25 0.25 0.5 0.5 0.5 1.0 1.0 1.0

0.55 0.35 0.62 0.52 0.44 0.67 0.70 0.60

0.50 0.35 0.48 0.40 0.38 0.55 0.60 0.52

0.60 0.50 0.40 0.32 0.38 0.46 0.51 0.43

0.85 0.80 0.40 0.30 0.41 0.37 0.42 0.33

1.2 1.3 0.48 0.34 0.52 0.32 0.34 0.24

1.8 2.0 0.60 0.44 0.68 0.29 0.28 0.17

3.1 2.8 0.78 0.62 0.92 0.29 0.26 0.15

4.4 3.8 1.1 0.92 1.2 0.30 0.26 0.17

6.0 5.0 1.5 1.4 1.6 0.37 0.29 0.21

Main, Cc,s V˙b / V˙c Ab /As

Ab /Ac

0.1

0.2

0.3

0.4

0.5

0.6

0.7

0.8

0.9

0.25 0.33 0.5 0.67 1.0 1.0 1.33 2.0

0.25 0.25 0.5 0.5 0.5 1.0 1.0 1.0

 0.01 0.08  0.03 0.04 0.72  0.02 0.10 0.62

 0.03 0  0.06  0.02 0.48  0.04 0 0.38

 0.01  0.02  0.05  0.04 0.28  0.04 0.01 0.23

0.05  0.01 0  0.03 0.13  0.01  0.03 0.23

0.13 0.02 0.06  0.01 0.05 0.06  0.01 0.08

0.21 0.08 0.12 0.04 0.04 0.13 0.03 0.05

0.29 0.16 0.19 0.12 0.09 0.22 0.10 0.06

0.38 0.24 0.27 0.23 0.18 0.30 0.20 0.10

0.46 0.34 0.35 0.37 0.30 0.38 0.30 0.20

Source: ASHRAE Handbook 1989, Fundamentals. Reprinted with permission. For details, refer to ASHRAE Handbook.

●

SH__ ST__ LG__ DF

●

17.62

The smaller the dif ference in air v elocity between the common end and the straight-through end, the lower the Cc,s value. At a greater ratio of branch duct velocity to main duct velocity vb /vc, especially when vb /vc 1.2, the Cc,s value for the diverging wye of rectangular duct with smooth elbow takeoff is negative.

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39445 Wang (MCGHP) Chap_17

pg 17.63

AIR SYSTEMS: AIR DUCT DESIGN

17.63

Return or Exhaust Duct Systems Behls (1978) compared four designs of a bunker ventilation duct system with 10 risers connected to a main exhaust duct. The results of these four designs are as follows:

Fixed-diameter riser, tee / diffuser Variable-diameter riser, tee / diffuser Fixed-diameter riser, conical tee Variable-diameter riser conical tee

System total pressure loss, in. WC (Pa)

Riser unbalance, in. WC (Pa)

 7.39 ( 1837)  5.89 ( 1464)  5.29 ( 1315)  4.03 ( 1002)

6.55 (1628) 1.77 (440) 4.06 (1009) 1.62 (403)

Using a greater branch duct v elocity than the main duct, and varying the sizes of the risers, select proper duct f ttings to decrease system total pressure loss and system unbalance. Example 17.2. For a round supply duct system made of galv anized steel with spiral seams, as shown in Fig. 17.20, its operational and constructional characteristics are as follows: Supply air temperature Kinematic viscosity of air Density of supply air Absolute roughness Local loss coeff cients in section FO-1 Fan total eff ciency, average Motor eff ciency Electrical energy cost Installation cost of duct CRF Annual operating hours

60°F (15.6°C) 1.59 104 ft2 / s (1.46 105 m2 / s) 0.075 lb / ft3 (1.20 kg / m3) 0.0003 ft (0.09 mm) 0.5 0.7 0.8 $0.08 / kWh $3.5 /ft2 0.10 3000

The total supply v olume f ow rate at the f an discharge is 3000 cfm (1.415 m 3 / s), and the adjusted volume f ow rates because of duct heat gain for each of the branch tak eoffs are illustrated in Fig. 17.20. First, size this supply duct system. Then if each branch tak eoff is connected to a VAV box that needs a total pressure loss of 0.75 in. WC (186 Pa) at its inlet, calculate the total pressure loss required for this supply duct system e xcluding the VAV box and do wnstream f exible duct and diffuser. Solution: 1. For the f rst iteration, start with the round duct section 2-5, which may be the last section of the design critical path. Assume a diameter of 12 in., or 1 ft, and an air velocity of 1200 fpm, or 20 ft / s. Assume also that the local loss coef f cient of the straight-through stream of the di verging tee Cc,s  0.15, and for the elbow Co  0.22. The Reynolds number based on diameter D is Re D 

vD 20 1  125,786  n 1.59 10 4

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39445 Wang (MCGHP) Chap_17

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pg 17.64

CHAPTER SEVENTEEN

From Eq. (17.42), the friction factor is f 

0.25 {log [ / (3.7D)  5.74 / (0.9Re D)]}2 0.25  0.0166 {log [0.0003 / (3.7 1)  5.74 / (0.9 125,786)]}2

From Eq. (17.93), the optimum pressure loss for the round duct section 2-5 is LCiufm

fL (1/CRF)E t   √D  C√D  V˙ 1  30 3.5 0.7 0.8  108 10 0.08 3000  0.0166 30 1   0.37√10.075    0.13 in. WC 1025 √1.0 pt2-5  108

0.82

0.18

0.46

r an

0.82

0.18

0.46

And from Eq. (17.90), the diameter of the duct section 2-5 when pt  0.13 in. WC is fL  √D  C√D V˙ p 0.0166 30  0.37√1.00.075  0.0511  √1.0

D2-5  0.0511

0.22

0.44

0.22 t

0.22

10250.44

1 0.130.22

 0.927 ft or 11.12 in. (282 mm)

2. Duct section 1-2 is one of the sections of the critical path. F or this section, if we assume that Cc,s  0.05 and the diameter is 1.2 ft, then the calculated friction factor f  0.0155 and the optimal total pressure loss is 0.8  2010 3.50.08 0.7 3000  0.155 20 1   0.05√1.20.075    0.0575 in. WC 2025 1.2 √ 0.82

pt1-2  108

0.18

0.46

and the sized diameter is 20  0.05√1.20.075 0.0155 √1.2

0.22

D1-2  0.051

20250.440.05750.22

 1.229 ft or 14.75 in. (375 mm)

3. Duct section FO-1 is one of the sections of the critical path. F or this section, C  0.5. Let us assume that the diameter is 1.67 ft and the calculated f  0.0147. Then 0.8  2510 3.50.08 0.7 3000  0.147 25   0.5√1.670.075 √1.67

ptFO-1  108 SH__ ST__ LG__ DF

0.82

0.18

1  0.0692 in. WG 30000.46

39445 Wang (MCGHP) Chap_17

SECOND PASS bzm 6/28/00

pg 17.65

AIR SYSTEMS: AIR DUCT DESIGN

17.65

and the sized diameter 25   0.5√1.670.075  0.0147 √1.67

0.22

DFO-1  0.051

30000.440.06920.22

 1.734 ft or 20.81 in. (529 mm) 4. Duct section 2-4 is another le g from junction 2. It must ha ve the same total pressure loss as duct section 2-5. Assume that Cc,b  1 and the diameter is 11 in., or 0.917 ft. Then 12  1√0.9170.075  0.0166 √0.917

0.22

D2-4  0.051

10000.440.130.22

 0.978 ft or 11.78 in. (299 mm)

5. Duct section 1-3 is another leg from junction 1. It must have the sum of the total pressure loss of duct sections 1-2 and 2-5, that is, pt1-3  0.0575  0.13  0.1875 in. WC. Assume that Cc,b  0.8 and the diameter is also 11 in. Then 12  0.8√0.9170.075  0.0166 √0.917

0.22

D1-3  0.051

9750.440.18750.22

 0.858 ft or 10.3 in. (262 mm)

6. The results of the f rst iteration are listed in Table 17.12. These results are rounded to standard sizes and provide the information for the selection of the di verging tee and wye and the determination of the local loss coeff cient. 7. For branch takeoffs 2-4 and 1-3, select proper diverging tees and wyes and, therefore, the Cc,b values based on the air v elocity of the sized duct sections. Vary the size of the duct sections if necessary so that their pt values are approximately equal to the pt of section 2-5 for branch duct 2-4, and the pt of the sum of sections 2-5 and 1-2 for branch 1-3. Use Eq. (17.88) to calculate the total pressure loss of sections 2-4 and 1-3 according to the rounded diameters. 8. After the di verging tee or wye is selected, recalculate the optimum total pressure losses and diameters for duct sections 2-5, 1-2, and FO-1 from Eqs. (17.93) and (17.90). 9. After two iterations, the f nal sizes of the duct sections, as listed in Table 17.12, are the following: Section 2-5, 11 in.; wye, 45°, diverging, round, 30° elbow Section 1-2, 14 in.; tee, diverging, round, 90° elbow Section FO-1, 20 in. Section 2-4, 11 in. Section 1-3, 12 in. For duct section 1-2, at vs /vc  1.2, Cc,s is about 0.07 according to Idelchik (1986), diagram 7-17. TABLE 17.12 Results of Computations of Duct Sizes and Total Pressure Loss in Example 17.2 First iteration Duct section 2-5 1-2 FO-1 2-4 1-3

Volume f ow V, cfm

ft

1025 2025 3000 1000 975

0.927 1.229 1.734 0.978 0.858

Diameter

Final results Air velocity

in.

Rounded in.

fpm

11.12 14.75 20.81 11.78 10.3

11 15 20 11 11

1553 1894 1374 1515 1241

fps

Friction Velocity ratio factor f vs /vc vb /vc

25.88 0.0166 31.97 0.0155 22.92 0.0147 25.25 20.69

0.94 1.20

Cc,s

Cc,b

0.011 0.07 0.80 0.90

0.65 1.70

C

Diameter

0.22

11 14 20 11 12

0.5

pt. in. WC 0.116 0.0645 0.0847

__SH __ST __LG DF

39445 Wang (MCGHP) Chap_17

17.66

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pg 17.66

CHAPTER SEVENTEEN

10. Because the local loss coef f cient Cc,b for the branch stream from the 45 ° diverging wye at junction 2 can be reduced to 0.45 if a conical branch is used, the critical path consists of duct sections FO-1, 1-2, and 2-5. The total pressure loss of this supply duct system e xcluding the VAV box and downstream f exible duct and diffuser, from Table 17.12, is pt  0.116  0.0645  0.0847  0.2652 in. WC (66 Pa)

17.12 DUCT SYSTEMS WITH NEGLIGIBLE PRESSURE LOSS AT BRANCH DUCTS Supply Duct Systems When supply outlets are directly mounted on the main duct without branch takeoffs, or the connecting duct between the supply outlet and the main duct is about 2 ft or less, the total pressure loss of the branch duct, excluding the supply outlet, is often v ery small or ne gligible. In such circumstances, system balancing of the supply duct system depends mainly on the sizes of the successi ve main duct sections. If volume dampers are installed just before the outlet or in the connecting duct, the damper modulation will also vary the space diffusion airf ow pattern.

Pressure Characteristics of Airflow in Supply Ducts The rectangular supply duct with transv ersal slots sho wn in Fig. 17.25 is an e xample of a supply duct system with a supply outlet directly mounted on the main duct. The volume f ow per ft2 (m2) of f oor area of this type of duct system may sometimes e xceed 8 cfm / ft2 (40.6 L / s m2). This type of supply duct system is often used in many industrial applications. Consider two planes n and n  1, with a distance of 1 unit length, say, 1 ft (0.305 m), between them. If the space air is at atmospheric pressure, the total pressure loss of the supply air that f ows from plane n, turns a 90°, and discharges through the slots is (Cc,bnv 2n  Cov 2on)

ptn  ptn  pc,bn  pton 

2gc

(17.99)

where pc,bn, pton  dynamic loss of branch stream when it makes a 90° turn and when it discharges from slots, lb / ft2 or in. WC (Pa) Cc,bn  local loss coeff cient of diverging branch stream with reference to velocity at plane n Co  local loss coeff cient of transversal slots with reference to velocity at slot vn, von  air velocity at plane n and at slot from plane n, ft / s (m / s) Similarly, at plane n  1 the total pressure loss of the branch stream when it dischar ges from the slots is pt(n1) 

(Cc,b(n1)v 2(n1)  Cov 2o(n1)) 2gc

(17.100)

The relationship of the total pressure between planes n and n  1 along the airf ow is given as ptn  pt(n1)  SH__ ST__ LG__ DF

( fnL n/Dn  Cc,sn)v 2n 2gc

(17.101)

In Eq. (17.101), fn, Ln, Dn, and Cs,cn indicate the friction f actor, length, circular equivalent, and Cc,s value between plane n and n  1, respectively. Substituting Eqs. (17.99) and (17.100) into

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FIGURE 17.25 Rectangular supply duct with transversal slots.

Eq. (17.101), Wang et al. (1984) recommended the follo wing formula to calculate vn1, in ft / s (m / s), based on the balanced total pressure before the transversal slots: v(n1) 

(v

2 on



 v 2o(n1))Co  Cc,bn  Cc,sn 

fn L n Dn

1

v  C 2 n

c,b(n1)

0.5



(17.102)

In Eq. (17.102), von and vo(n1), in ft / s (m / s), are the supply air velocities at the slots. For a cold air supply, usually vo(n1) von is desirable because of the effect of the duct heat gain. For a free area ratio of slot area to duct w all area of 0.5, Co for a perforated plate including discharged velocity pressure can be taken as 1.5. The vn, Ln, and Dn are known values during the calculation of vn1. Local loss coeff cients Cc,bn and Cc,sn can be determined from the experimental curves shown in Figs. 17.26 and 17.27. The coeff cient Cc,b(n1) can be assumed to ha ve a value similar to Cc,bn, and f can be calculated from Eq. (17.42). The mean air velocity in the duct at plane n  1, denoted by vn1, can then be calculated from Eq. (17.102). In Figs. 17.26 and 17.27, V˙ on indicates the v olume f ow rate of supply air dischar ged from the slots per ft length of the duct, in cfm / ft (L / s  m); V˙ n represents the volume f ow rate of the supply air inside the duct at plane n, in cfm (m3 / s). Sizing of this kind of supply duct system starts from the duct section immediately after the fan discharge. This section can be sized from Eqs. (17.90) and (17.93) based on life-c ycle cost

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FIGURE 17.26 Local loss coeff cient Cc,s versus volume f ow ratio V˙on / V˙n . (Source: ASHRAE Transactions, 1984, Part II A. Reprinted with permission.)

optimization. The designer should check for space a vailable and noise, if sound control is required. The successive duct sections can then be sized by calculating from Eq. (17.102). If supply duct systems with ne gligible pressure loss at branch ducts are installed with outlets whose Cc,bn and Cc,sn are not known, the static regain method is recommended for sizing the supply main duct for a large duct system and the equal-friction method for a small system. Example 17.3. A galvanized steel rectangular duct section with transversal slots has the following construction and operational characteristics at plane n

SH__ ST__ LG__ DF

Dimension: Absolute roughness  Supply air temperature Volume f ow rate V˙ n

4-ft width 3-ft height (1219 mm 914 mm) 0.0005 ft (0.15 mm) 65°F (18.3°C) 23,760 cfm (11.21 m3 / s)

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FIGURE 17.27 Local loss coef f cient Cc,b versus v olume f ow ratio V˙on / V˙n . ( Source: ASHRAE Transactions, 1984, Part II A. Reprinted with permission.)

Volume f ow discharged from slots related to plane n Discharged velocity at slots related to plane n Required discharged velocity at plane n  1

270 cfm / ft (417.9 L / sm) 7.5 ft / s or 450 fpm (2.29 m / s) 7.6 ft / s or 456 fpm (2.32 m / s)

If the height of the rectangular duct remains the same, n  1, 10 ft (3.05 m) from plane n. Solution

size the dimension of this duct at plane

1. From the information given, the air velocity at plane n is vn 

23,760  33 ft / s (1980 fpm) 60 4 3

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And from Eq. (17.48), the circular equivalent is De 

1.3(ab)0.625 1.3(3 4)0.625   2.9 ft (a  b)0.25 (3  4)0.25

The Reynolds number of the supply air Re D 

33 2.9  596,700 1.62 104

The friction factor f 

0.25 {log [ / (3.7D)  5.74 / (0.9 Re D)]}2 0.25  0.0139 {log [0.0005 / (3.7 2.9)  5.74 / (0.9 596,700)]}2

2. From the information given, the volume f ow ratio is V˙on 270   0.0113 23,760 V˙n and the velocity ratio is von 7.5   0.227 vn 33 From Figs. 17.26 and 17.27, Cc,sn   0.0025 per ft and Cc,bn  1.22. 3. Assume that Cc,b(n1)  1.23. Then from Eq. (17.102), v(n1) 

(v

2 on



 v 2o(n1))Co  Cc,bn  Cc,sn 

 (7.52  7.62)1.5 



fL 2 1 vn D Cc,b(n1)



1.22  (0.0139 10 / 2.9)  0.0025 10)]332 1.23

 29.6 ft / s (9.02 m / s) 4. The required volume f ow rate from the slots related to plane n  1 is 9270 7.6 V˙o(n1)   273.6 cfm / ft 7.5 Then the volume f ow rate of air at plane n  1 is V˙o(n1)  23,760  273.6 10  21,024 cfm The duct area required at plane n  1 is SH__ ST__ LG__ DF

An1 

0.5



21,024  11.83 ft 2 (1.00 m2) 60 29.6

0.5



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FIGURE 17.28

17.71

A return duct system with short connecting ducts for a textile mill.

The width of the duct is W

10.78  3.59 ft (1.095 m) 3

Return or Exhaust Duct Systems Return duct systems with v ery short connecting ducts between the return inlet and the main duct are often used in industrial applications. They are often used in lar ge workshops and need nearly equal return volume f ow rate for each branch intake. To provide a better system balance, it is necessary to reduce the total pressure dif ference between the branch inlets at the tw o ends of the main duct. For a conditioned space where noise is not a major problem, one effective means of reducing the total pressure loss of a long return main duct is to have a higher branch duct velocity than the velocity in the main duct. Because a total pressure dif ference between the branch inlets and the main duct at two ends is inevitable, the area of the return grille can be v aried along the main duct to provide a more even return volume f ow rate. Unlike the supply outlets, the variation of the area of the return grille has a minor effect on the space air diffusion. A return duct system with short connecting ducts is sho wn in Fig. 17.28 for a te xtile mill. Twelve f oor grilles are connected to the return main duct at the lo wer f oor. The connecting duct has a f oor to main return duct difference of 2.6 ft (0.79 m). The sizes of the branch ducts vary from 16 in. 16 in. (406 mm 406 mm) at the remote end to 12 in. 8 in. (305 mm 203 mm) near the fan end. The main duct v aries from 16 in. 32 in. (406 mm 813 mm) to 60 in. 32 in. (1524 mm 813 mm). In Fig. 17.28, the variations in total pressure pt and static pressure ps along the main duct are sho wn by the upper and middle curv es, in in. WG (Pa); and the v elocity in main return duct v, in fpm (m / s), is shown by the lower curve. The velocity in the branches v aries from 700 fpm (3.56 m / s) at the remote end to 2000 fpm (10.16 m / s) at the fan end. Note that pt at grille 6 is  0.111 in. WG (  27.6 P a),  0.105 in. WG (  26.1 P a) at grille 5, and  0.114 in. WG

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( 28.3 Pa) at grille 2. Such a satisf actory pressure balance along the return duct is mainly due to the higher branch velocity in the grilles near the fan end.

17.13 REQUIREMENTS OF EXHAUST DUCT SYSTEMS FOR A MINIMUM VELOCITY To transport dust particles or particulate products contained in the air , exhaust duct systems require a minimum velocity in all duct sections of the system. These systems are used in man y industrial applications and usually have air velocity ranging between 2400 and 4000 fpm (12.2 and 20.3 m / s). In addition to the variation of the sizes of the branches’ ducts, it is essential to select the proper conf guration of duct f ttings to pro vide a better system balance and to reduce the total pressure loss of the system (critical path). The following are recommendations for exhaust duct system designs: ●

●

●

Round ducts usually produce smaller pressure losses; they are more rigid in construction. Air v elocity inside the duct must not e xceed the minimum v elocity too much, in order not to waste energy. Well-sealed joints and seams are important for reducing air leakage at higher pressure dif ferentials.

17.14 COMPUTER-AIDED DUCT DESIGN AND DRAFTING According to Amistadi (1993) and “Scientif c Computing ” (1998), the follo wing computer -aided duct design and drafting (CADD) programs are widely used in air duct design: E20-II Duct Design, Carrier Corp.; Varitrane Duct Designer , The Trane Co.; and Softdesk HV AC. Carrier ’s E20-II requires a disk space of installation of 4 MBytes and a RAM of 1 MByte, Trane’s Varitrane needs a disk space of 13 MBytes and a RAM of 12 MBytes, and Softdesk HV AC needs a disk space of 7 MBytes and a RAM of 8 MBytes. All these three programs will run in Windows 95. An ef fective and intelligent computer -aided duct design and drafting program pro vides the following functions: Drafting The computer program pro vides full, three-dimensional (3D) capabilities and dra ws round, rectangular, and f at oval ducts and f ttings in any cross-sectional plane. The program offers two types of single-line layouts: (1) duct and f tting, which automatically produce elbows, transitions, and reducers from plan or multistory vie ws, and (2) isometric, which allows the designer to produce a 3D drawing in an easy w ay. A single-line layout can be con verted to double-line or 3D ducts and f ttings layouts, or the designer can construct the 2D double-line or 3D ducts and f ttings drawings directly. Duct layout can be produced in 2D or 3D in different layers simultaneously. Schedules and Layering

SH__ ST__ LG__ DF

Schedules of a computer -aided duct design program include dif fusers, ducts, f ttings, and related equipment, such as terminals and air -handling units or packaged units. Duct schedules include the tag number, quantity, size, type, length, gauge, and bill of material. Fitting schedules include the tag number, quantity, type, name, gauge, f tting junction sizes, and cfm (L / s). Values can be totaled and stored for each duct or f tting type. Tag editing capacity includes: creating, moving, and renumbering. The computer program prefers to ha ve a nested-layer hierachy of f ve le vels deep including building information, plan types, layer name and level (elevation), and construction modif er. Build-

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ing information layers are grouped by supply , return, and exhaust ducts and subdi vided into ducts, diffusers, and f ttings. For each of these duct components, layers can be single-line, 2D or 3D construction with tags, labels, dimensions, and accessories. Plan types similar to the dra wing information layers include 2D or 3D supply, return, and exhaust ducts. Besides, there are also f oor, interior, exterior, and bird’s-eye view plans. A level layer indicates its attrib ute characteristic and the v ertical elevation. Each layer can be edited, modif ed, and displayed.

Design Interface Computer-aided duct design softw are either must pro vide interfaces with available duct sizing programs or itself must ha ve the software for duct sizing. It is preferable that the duct design softw are size the ducts by using equal friction, static regain, constant-velocity, and even cost optimiztion T methods. The duct design softw are e xtracts system geometry and f ow rates from the AutoCAD drawing, exports the information to the selected duct sizing program, and uses the sizes calculated to produce a double-line or 3D drawing to replace the centerline drawing.

Running Processes Drawing a Centerline Layout. Designers are required to enter a f an and centerline layout in AutoCAD. They have the option to add terminals, diffusers, necessary dampers, f ttings, duct sections, and related data to the dra wing. Terminal text, line color, and f tting text are interpreted as the airf ow, duct shape, and f tting conf gurations. An ASHRAE Duct Fittings Database with local loss coeff cients of 225 types of f ttings provides default f tting types and coeff cients. Analyzing the Duct System. With a completed centerline layout, designers are ask ed to choose the duct sizing softw are. Default values that are used for the centerline layout, such as duct shape, and airf ow rates are passed to the duct-sizing software. The designer is ask ed to select the f an location. The computer program connects all lines to the the fan in an attempt to connect each line to the end of the preceding line. The computer program tests the continuity of the duct system by checking that all selected components are connected and all junctions can be found. All lines must ha ve a f an, terminal, or diffuser, or another centerline at each end. The computer program generates nodes and inserts a node tag for the f an, main duct section, and branch takeoff endpoints. Node shape is determined according to the duct shape and class of node. Duct section information is stored in the upstream node. Each main duct node and branch duct node are set with default attributes required by the duct-sizing program. Editing Default Attributes and Running Duct-Sizing Program. If the designer is sati f ed with all default values, he or she can proceed to run the duct-sizing computer programs. If the designer needs to change any of the default values, she or he can use AutoCAD to edit duct-sizing input default attributes and visually pick the duct sections from the layout for editing before running the duct-sizing program. When the editing w ork is done, modif ed node data are used to update input data for the ductsizing program. The designer uses the command to run the chosen duct-sizing program. Converting from Centerline to Double-Line or 3D. Centerline layout or centerline 3D layout can be input into sizing calculations, and the computer program will produce a double-line layout or a 3D duct system layout. The conversion begins with the fan and produces the layout in the following order: f ttings, transitions, reducers, duct segments, and dampers. Total Pressure Loss and F an-Sizing Calculations. Some computer-aided duct design computer programs have the ability to determine an e xisting duct layout shape and size in the CAD en vironment and load it into the duct system total pressure loss and fan sizing calculations.

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17.15 DUCT LINER AND DUCT CLEANING Duct Liner Duct liners are lined internally on the inner surf ace of the duct wall. They are mainly used for noise attenuation and form part of the duct insulation layer that helps to reduce the heat transfer between the conditioned air inside the duct and the ambient air surrounding the duct. Fiber glass blanket or boards with a thickness of 12 in. (13 mm), 1 in. (25 mm), 112 in. (38 mm), and 2 in. (50 mm) have been used as duct liner for decades and proved to be cost-effective. In 1994, the U.S. Department of Health and Human Services (DHHS) announced that f berglass will be listed as a material “reasonably anticipated to be a carcinogen.” Although there are many valid arguments concerning mold gro wth and f ber erosion from f berglass duct liners, the use of f berglass in several institutional, educational, and medical projects has been banned or severely limited. Recommendations concerning the use of f berglass duct liner and alternati ves in HVAC&R systems are discussed in Secs. 19.3 and 19.5.

Duct Cleaning The accumulation of dirt and debris as well as the gro wth of mold and fungi inside the ductw ork normally will not occur in a properly designed (such as equipped with medium- or high-ef ficiency air f ilters), installed, operated, and maintained air system including ductw ork. When the cause of indoor air pollution is dirty ductw ork — accumulation of dirt and debris or mold growth, or both inside the ductw ork — source removal or duct cleaning is often the most ef fective treatment. Spielmann (1997) discussed the most widely used duct cleaning equipment and techniques as follows: Planning and System Inspection. Drawings are used as a valuable planning tool. Duct cleaning is usually performed in zones of about 25 ft (7.5 m) or less. Contaminated ducts and uncleanable types of ducts, such as f exible ducts and certain lined ducts, that cannot be cleaned without damaging their duct liner should be identi f ed. Inspection should be performed when the area is unoccupied and the air system is turned off. Starting the J ob. Turn off the air system. The return side is often 5 to 10 times dirtier than the supply side and is al ways cleaned f rst, before the supply side. Lay drop cloths to protect occupied areas. Duct cleaning is started from return grilles and outdoor air intak es. Existing openings should be used for access whene ver possible. Additional access openings often must be cut. Cleaning zones are isolated by using in f atable bladders inserted into the ducts. Duct sections are cleaned along the ductwork toward the air-handling unit or air handler. Agitation Devices. Contaminants are agitated by cleaning devices so that dirt and debris are loosened drom the duct w all and e xtracted by the airstream. There are three currently used agitation methods: ●

●

●

SH__ ST__ LG__ DF

Manual agitation using contact vacuuming requires physical access to all surfaces. Use portable compressed-air skipper nozzles to dislodge debris from the duct surfaces. Adopt a rotary brush or other cleaning head spinning on the end of a motor -driven f exible shaft. When using a rotary brush in a lined duct, do not allow the brush to remain in the same spot for too long.

Most of the duct cleaning work requires a combination of these agitation methods. Duct Vacuum. A duct v acuum (negative air machine) is used to e xtract loosened debris into the airstream of ne gative pressure so that contaminants will not pollute the conditioned space. A duct

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vacuum used in duct cleaning is often portable and consists of a blo wer, a HEPA f lter to collect debris, an outer casing, and connecting f exible duct. A large duct v acuum often has an e xtracting volume f ow rate of 4000 cfm (1.89 m 3 / s) and more. F or each cleaning zone, there are tw o openings : one is the access to the agitation device, and the other opening connects the duct under cleaning directly to the portable duct v acuum. A smaller portable duct v acuum is used to e xtract debris from the fan blades, coils, and condensate drip pans. Sealing of Access Openings. After the duct cleaning is completed, the cleaned ducts should be isolated from the cleaning zone in which duct cleaning is progressing by means of in f atable bladder. The access opening should be patched using reusable doors, sheet metal, duct tape, and sealants.

17.16 PRESSURE AND AIRFLOW MEASUREMENTS The total pressure of the air f ow is the sum of the static pressure, which acts in all directions, and the velocity pressure, which results from the impact and the inertia of the air f ow. The Pitot tube and manometers, shown in Figs. 17.29 and 17.30, are widely used to measure total pressure, static pressure, velocity pressure, and thus airf ow inside air ducts. A standard Pitot tube consists of tw o concentric tubes ha ving an outside diameter D  516 in. (8 mm) and an inner tube of diameter of 18 in. (3.2 mm), as shown in Fig. 17.29. The inner tube opens to the air f ow at the nose, and the other end is used as the total pressure tap. Eight equally spaced holes of 0.04-in. (1-mm) diameter allo w the air to f ow into the hollo w space between the outer and inner tubes. The hollow space is connected to the static pressure tap. Because the small holes are located perpendicular to the centerline of the head, they are able to measure the static pressure when the head of the Pitot tube is placed in a position opposite to the direction of the air f ow. The centerlines of the small holes are located at a distance of 8 D from the nose and 16 D from the stem. The negative pressure produced at the nose is nearly balanced by the positive pressure at the stem. The U tube shown in Fig. 17.30a is the simplest type of manometer used to measure pressure in air ducts. The v ertical dif ference in the liquid column indicates the pressure reading. F or more

FIGURE 17.29 Pitot tube.

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FIGURE 17.30 Manometers: (a) U tube; ( b) inclined manometer.

SH__ ST__ LG__ DF

FIGURE 17.31

Pressure measurements in air ducts.

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accurate measurement, an inclined manometer (see Fig. 17.30 b) is often used. The relationship between the height of the liquid column H and the length of the magnif ed inclined scale is H  L sin 

(17.103)

where L  length of inclined scale with liquid column, ft (m)   angle of inclination of inclined tube, deg Because the cross-sectional area of the inclined tube is v ery small compared with that of the reservoir, the change in the liquid level of the reservoir can be ignored. The following precautions should be taken when one takes pressure measurements in an air duct such as that shown in Fig. 17.31: ●

●

●

The nose of the Pitot tube must al ways be placed opposite to the direction of air f ow whether in the suction side or discharge side of the fan. When the total pressure or static pressure is measured, one leg of the U tube or inclined manometer must be open to the atmospheric air as the reference datum. The smaller pressure is al ways connected to the open end of the inclined tube of the inclined manometer. For a v elocity pressure measurement, the total and static pressure taps must be connected to the two ends of the manometer . The total pressure tap is connected to the reserv oir, and the static pressure tap is connected to the inclined tube. Velocity pressure is always positive in the direction of airf ow.

Equal-Area Versus Log Tchebycheff Rule Because velocity is usually not uniform across the cross-sectional area, a transverse is often used to determine the average velocity. The equal-area method was widely used before. Recently ASHRAE

60

 60

0.500H 0.712H 0.926H H



0.288H



60

0.074H

60



60

60

0.032D 0.135D

0.061W

0.321D

0.235W

0.437W 0.563W 0.765W 0.939W W

(a)

0.679D 0.865D 0.968D D

(b)

FIGURE 17.32 Measuring points in rectangular and round duct transv erse: (a) log Tchebycheff rule for rectangular duct transverse; (b) log-linear rule for round duct transverse.

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Standard 111-88, the Associated Air Balance Council (AABC), and the National En vironmental Balancing Bureau (NEBB) all adopted the log Tchebycheff rule for rectangular duct and round duct transverse, as shown in Fig. 17.32 a and b, because this rule provided greatest accuracy. The reason is that the location of transv erse points of these rules has tak en into account the ef fect of the duct wall friction and the reduction in v elocity near the duct w alls. For round duct transv erse, the log Tchebycheff rule and log-linear transv erse method are similar. Refer to ASHRAE Standard 111-88 for details. The log Tchebycheff rule w as de veloped by a mathematician named Tchebycheff in 1977. In MacFerran (1999), airf ow in ducts measured by the equal-area method w as often higher than correct results, from 7 percent higher maybe up to 25 percent higher. When the velocity pressure is measured, in inches WC (Pa), the air velocity, in fpm, can be calculated from Eq. (17.10). For SI units, the air velocity, in m / s, can be calculated from Eq. (17.9), in which gc  1.

REFERENCES

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Amistadi, H., Design and Drawing Software Review, Engineered Systems, no. 6, 1993, pp. 18 – 29. ASHRAE, ASHRAE Handbook 1996, HVAC Systems and Equipment, ASHRAE Inc., Atlanta, GA, 1996. ASHRAE, ASHRAE Handbook 1997, Fundamentals, Atlanta, 1977a. ASHRAE, Duct Fittings Loss Coeff cient Tables, Atlanta, 1997b. ASHRAE, ASHRAE / IESNA Standard 90.1-1999, Energy Standard for Buildings Except Low-Rise Residential Buildings, ASHRAE Inc., Atlanta, 1999. Behls, H., Balanced Duct Systems, ASHRAE Transactions, 1978, Part I, pp. 624 – 646. Brooks, P. J., Duct Design Fundamentals, ASHRAE Journal, no. 4, 1995, pp. 69 – 76. Coe, P. E., The Economics of VAV Duct Looping, Heating /Piping /Air Conditioning, August 1983, pp. 61 – 64. Colebrook, C. F., Turbulent Flow in Pipes with Particular Reference to the Transition Region Between the Smooth and Rough Pipe Laws, Journal of Institute of Civil Engineers, February 1939, p. 133. Dean, R. H., Dean, F. J., and Ratzenberger, J., Importance of Duct Design for VAV Systems, Heating / Piping / Air Conditioning, August 1985, pp. 91 – 104. Evans, R. A., and Tsal, R. J., Basic Tips for Duct Design, ASHRAE Journal, no. 7, 1996, pp. 37 – 42. Habjan, J., Medium Pressure Duct Sizing and Design, Heating / Piping / Air Conditioning, December 1984, pp. 95 – 100. Heyt, H. W., and Diaz, M. J., Pressure Drop in Flat Oval Spiral Air Duct, ASHRAE Transactions, 1975, Part II, pp. 221 – 232. Heubscher, R. G., Friction Equivalents for Round, Square and Rectangular Ducts, ASHVE Transactions, 1948, pp. 101– 144. Idelchik, I. E., Handbook of Hydraulic Resistance, 2d ed., Hemisphere Publishing Corp., Washington, 1986. MacFerran, E. L., Equal Area vs. Log-Tchebycheff, Heating / Piping / Air Conditioning, no. 12, 1999, pp. 26 – 31. Miller, E. B., and Weaver, R. D., Computer-Aided Duct Design in a Small Engineering Off ce, ASHRAE Transactions, 1978, Part I, pp. 647 – 664. NFPA 90A, Standard for the Installation of Air Conditioning and Ventilating Systems, National Fire Protection Association, Quincy, MA, 1985 ed. North American Insulation Manufacturers Association, New Duct Liner Standard Addresses HVAC System Challenges, Heating / Piping / Air Conditioning, no. 2, 1994, pp. 61 – 67. Scientif c Computing, Software Review: Up for Review (Again), Engineered Systems, no. 1, 1998, pp. 76 – 84. Shitzer, A., and Arkin, H., Study of Economic and Engineering Parameters Related to the Cost of an Optimal Air Supply Duct System, ASHRAE Transactions, 1979, Part II, pp. 363 – 374. Spielmann, S., Check Your Basic Skill on Duct and Ventilation System Cleaning, Air Conditioning, Heating & Refrigeration News, Feb. 24, 1997, pp. 3– 5.

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Stoecker, W. F., and Bertschi, R. I., Design Duct System for Minimum Life-Cycle Cost, Proceedings of Improving Eff ciency and Performance of HVAC Equipment and Systems for Commercial and Industrial Buildings, vol. 1, 1976, pp. 200 – 207. Swim, W. B., and Griggs, E. I., Duct Leakage Measurement and Analysis, ASHRAE Transactions, 1995, Part I, pp. 274 – 291. TIMA / HPAC, Fiber Glass Duct Systems, Heating / Piping / Air Conditioning, October 1986 Supplement. Tsal, R. J., and Behls, H. F., Evaluation of Duct Design Methods, ASHRAE Transactions, 1986, Part I A, pp. 347 – 361. Tsal, R. J., Behls, H. F., and Mangel, R., T-Method Duct Design, Part I: Optimization Theory, ASHRAE Transactions, 1988, Part II, pp. 90 – 111. Tsal, R. J., Behls, H. F., and Mangel, R., T-Method Duct Design, Part II: Calculation Procedure and Economical Analysis, ASHRAE Transactions, 1988, Part II, pp. 112 – 150. Wang, S. K., Air Conditioning, vol. 2, Hong Kong Polytechnic, Hong Kong, 1987. Wang, S. K., Leung, K. L., and Wong, W. K., Sizing a Rectangular Supply Duct with Transversal Slots by Using Optimal Cost and Balanced Total Pressure Principle, ASHRAE Transactions, 1984, Part II A, pp. 414 – 429.

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AIR SYSTEMS: SPACE AIR DIFFUSION H1CC H2CC

T1

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18.1 PRINCIPLES OF SPACE AIR DIFFUSION 18.2 Draft and Effective Draft Temperature 18.2 Air Diffusion Performance Index 18.3 Space Diffusion Effectiveness Factor 18.4 Ventilation Effectiveness and Air Exchange Effectiveness 18.4 18.2 AIR JETS 18.5 Free Isothermal Jets 18.5 Throw, Entrainment Ratio, and Characteristic Length 18.7 Confined Air Jets 18.8 Free Nonisothermal Jets 18.10 18.3 SUPPLY OUTLETS AND RETURN INLETS 18.11 Grilles and Registers 18.11 Ceiling Diffusers 18.12 Slot Diffusers 18.14 Nozzles 18.16 Accessories for Supply Outlets 18.17 Return and Exhaust Inlets 18.17 Light Troffer Diffuser and Troffer-Diffuser Slot 18.19 18.4 MIXING FLOW 18.20 Airflow Pattern and Space Air Diffusion 18.20 Principles and Characteristics of Mixing Flow 18.20 Mixing Flow Using High Side Outlets 18.21 Mixing Flow Using Ceiling Diffusers 18.23 Mixing Flow Using Slot Diffusers 18.24 Mixing Flow Using Sill or Floor Outlets 18.24 Stratified Mixing Flow 18.25 18.5 COLD AIR DISTRIBUTION 18.28 Cold Air Distribution versus Conventional Air Distribution 18.28 High Induction Nozzle Diffusers 18.28 Characteristics of Cold Air Distribution Systems 18.29 Case Study — Florida Elementary School 18.29 Performance of Ceiling and Slot Diffusers 18.29 Cold Air Distribution with Fan-Powered VAV Boxes 18.30 Surface Condensation 18.30

18.6 DESIGN PROCEDURE OF MIXING FLOW AIR DIFFUSION SYSTEM 18.31 Select the Type of Supply Outlet 18.31 Volume Flow Rate per Outlet or per Unit Length 18.32 Choose an Optimum Throw – Characteristic Length Ratio 18.33 Determine the Design Characteristics of Slot Diffusers in Perimeter Zone 18.33 Select the Specific Outlet from Manufacturer’s Catalog 18.34 Determination of Final Layout of Supply Outlets and Return Inlets 18.34 18.7 DISPLACEMENT FLOW AND UNIDIRECTIONAL FLOW 18.38 Displacement Flow 18.38 Unidirectional Flow 18.38 Unidirectional Flow for Clean Rooms 18.39 Ventilating Ceiling 18.40 Ceiling Plenum and Supply Air Velocity 18.41 18.8 STRATIFIED DISPLACEMENT FLOW 18.42 Two-Zone Stratified Model 18.42 Operating Characteristics 18.42 Comparison of Stratified Displacement Flow and Mixing Flow 18.43 18.9 PROJECTING FLOW — SPOT COOLING / HEATING 18.44 Benefits of Projecting Flow 18.44 Industrial Spot Cooling Systems 18.44 Recommendations in Spot Cooling Design 18.46 Desktop Task Conditioning Systems 18.46

Performance of Desktop Task Conditioning Systems 18.47 Application of Desktop Task Conditioning Systems 18.48 18.10 UPWARD FLOW UNDERFLOOR AIR DISTRIBUTION 18.48 Upward Flow from Floor Plenum 18.48 Thermal Storage of Floor Plenum 18.49 Heat Unneutralized 18.50 Maintaining a Consistent Access Plenum Temperature 18.50 Floor Plenum Master Zone Air Temperature Control 18.50

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Design Considerations 18.50 Applications of Upward Flow Underfloor Air Distribution 18.51 18.11 COMPUTATIONAL FLUID DYNAMICS 18.51 CFD Becomes More Popular 18.52

Reynolds-Averaged Navier-Stokes Equations 18.52 Numerical Methods 18.52 Conducting CFD Experiments 18.54 REFERENCES 18.54

18.1 PRINCIPLES OF SPACE AIR DIFFUSION Space air diffusion distributes the conditioned air containing outdoor air to the occupied zone (or a given enclosure) in a conditioned space according to the occupants’ requirements. An occupied zone is a space with the follo wing dimensions: (width of room  1 ft)  (depth of room  1 ft)  (a height of 6 ft). A satisfactory space air dif fusion evenly distributes the conditioned and outdoor air to provide a healthful and comfortable indoor en vironment for the occupants or the appropriate environment for a specific manu acturing process, at optimum cost. Because space air diffusion is the last process of air conditioning and tak es place entirely within the conditioned space, it directly affects the effectiveness of air conditioning. Because the dif fused air and ambient air are transparent, space air diffusion is also difficult to trace

Draft and Effective Draft Temperature One of the most common complaints about air conditioned space is draft. Draft is defined as an un wanted local cooling of the human body caused by air mo vement and lower space air temperature. Recent developments by F anger et al. (1988, 1989) demonstrate that air v elocity fluctuations als create drafts, and they investigated the impact of turbulence intensity on the sensation of draft. Figure 18.1 depicts v elocity fluctuations in a typical conditioned space. A new dimensionless parameter called the turbulence intensity Itur is introduced and expressed as Itur 

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v vm

FIGURE 18.1 Fluctuations of air v elocity in a typical air conditioned space. (Adapted with permission fr om ASHRAE Journal, April 1989, p. 30.)

(18.1)

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where v  standard deviation of air velocity f uctuations, fpm (m / s) vm  mean air velocity, fpm (m / s) Experiments sho wed that if the space temperature Tx  73.4°F (23 °C) and vm  30 fpm (0.15 m / s), the percentage of dissatis f ed occupants may increase from about 10 to 15 percent when Itur increases from 0.1 to 0.5. A parameter called the ef fective draft temperature , in °F (°C), which combines the effects of uneven space air temperature and air mo vement, is often used to assess the deviations of local magnitudes from the mean value, and it is def ned as (18.2)

  Tx  Tr  a(vx  vrm ) where Tx, vx  space air temperature and velocity at specif c location, °F (°C) and fpm (m / s) Tr  mean space air temperature or set point, °F (°C)

In Eq. (18.2), a is a con version constant to combine the ef fects of space air temperature and air movements; its value is 0.07 when T is expressed in °F and v is expressed in fpm, and 8 when T is expressed in °C and v in m / s. The desirable mean space air v elocity vrm, in fpm (m / s), is closely related to the space air temperature Tr to be maintained, the metabolic rate, and the clothing insulation of the occupant. According to Fig. 4.8, for a space relati ve humidity   50 percent, Trad  Tr , a metabolic rate of 400 Btu / h (117.2 W), and a clothing insulation of 0.6 clo (0.093 m 2  °C / W), their relationship may be given as follows: Tr , °F vrm , fpm

70 16

72 20

74 25

76 32

78 40

80 55

Air Diffusion Performance Index To an extent, the magnitude of effective draft temperature  also ref ects the degree of thermal comfort that can be pro vided by comfort air conditioning systems. During cooling mode in commercial and public b uildings, if the space temperature is maintained between 75 and 78 °F (23.9 and 25.6°C), Trad  Tr, space air velocity vr 55 fpm (0.28 m / s), and the space relative humidity is between 30 and 70 percent, then most sedentary occupants feel comfortable when  3°F 

2°F (  1.7°C  1.1°C). The air dif fusion performance inde x (ADPI), in percent, which evaluates the performance of space air diffusion, is calculated as ADPI 

N  100 N

(18.3)

where N  number of points measured in occupied zone in which  3°F  2°F ( 1.7°C  1.1°C) N  total number of points measured in occupied zone The higher the ADPI, the higher the percentage of occupants who feel comfortable. Maximum ADPI approaches 100 percent. Because air v elocity f uctuations create drafts, therefore, they also af fect ADPI. Abu-Ei-Hassan et al. (1996), based on the impact of turb ulence intensity Tu recommended by Fanger et al. (1988), conducted tests to determine the ef fect of turb ulence intensity Tu (v elocity f uctuations) on the ADPI. Abu-Ei-Hassan et al. took the centerline velocity of the test room, which has nearly the same f ow characteristic as the centerline v elocity of the air jet, as the air v elocities of the local points, and the ADPIs thus calculated as Air Diffusion Performance Inde x that tak e into consideration the Tu impact. The results sho wed that for a side wall outlet space air dif fusion system with a room

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cooling load intensity of 10 to 20 Btu / h ft2 (31.5 to 63 W / m2) the Air Diffusion Performance Index considering the turbulence effect were about 10 percent less than those that did not tak e the impact of Tu into account. At higher cooling loads and air f ow rates, the effect of Tu became lar ger. Actually, the drop in ADPI is less than 10 percent for cooling load intensity between 10 and 20 Btu / h ft2 (31.5 and 63 W / m2) due to the impact of Tu because the points of air v elocity located on the centerline of the test room are only a small portion of the total number of measured points. In cooling mode operation, ADPI is an appropriate inde x to use to evaluate a space air diffusion system. For heating mode operation, the temperature gradient between tw o points in the occupied zone may be a better indicator of the thermal comfort of the occupants and the ef fectiveness of a space air dif fusion system. Usually , the temperature gradient is less than 5 °F (2.8 °C) within an occupied zone.

Space Diffusion Effectiveness Factor The effectiveness of a space air diffusion system can also be assessed by using a space dif fusion effectiveness factor for air temperature T or for air contamination C. Both factors are dimensionless. Effectiveness factor T compares temperature differentials and C compares contamination differentials as follows:

T 

Tre  Ts Tex  Ts  Tr  Ts Tr  Ts

C 

Cex  Cs Cr  Cs

(18.4)

where T  temperature, °F (°C) C  concentration of air contamination, g / m3 In Eq. (18.4), subscript re represents the recirculating air, ex the exhaust air, r the space air or air at the measuring point, and s the supply air. When  1, the space air dif fusion is considered ef fective. If  1, a portion of supply air has f ailed to supply the occupied zone and e xhaust through the return or e xhaust inlets directly. Parameters in both the numerator and denominator must be in the same units.

Ventilation Effectiveness and Air Exchange Effectiveness ASHRAE Handbook 1997, Fundamentals, def nes ventilation effectiveness as a description of an air distribution system ’s ability to remo ve internally generated contaminants from a zone, space, or building. Air change ef fectiveness ex is def ned as air system ability to deli ver ventilation air to a zone, space, or building. The age of air age, in minutes or hours, is the time period that outdoor ventilation air has been in a zone, space, or building, and age can be evaluated for an existing building using the trace gas method. Chamberlin et al. (1999) proposed the follo wing when gro wth (step-up) and decay (step-down) tracer methods are used:



age  (end  initial) 1 

Cavg Cend



 (stop  start)(Cavg  Cstart)

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(18.5a)

where Cavg  time-averaged concentration of tracer gas measured in test zone from when tracer gas injections began to when last sample is taken, g / m3 or ppm Cend, Cstart  concentration of tracer gas when last sample was taken at time end and when f rst sample was taken at start, g / m3 or ppm

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end  initial  time elapsed from when tracer gas injection began to when last sample was taken, min or h stop  start  time elapsed from when f rst sample was taken to when tracer gas injection stopped, min or h

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The youngest air is the freshest air through in f ltration. A longer age means a poorer outdoor air delivery. Another measure of ventilation is the nominal time constant N, in min or h, and it can be calculated as V˙s (18.5b) N  Vr where V˙s  supply volume f ow rate to zone, space, or room, cfm (m3 /s) Vr  volume of zone, space, or room, ft3 (m3) Air exchange effectiveness is a relative index to measure how well the outdoor air is diffused to various locations in the occupied zone. Nominal air change effectiveness N can be calculated as

N 

N age,N

(18.6)

and age,N is evaluated through tracer gas measurement. Normally, for a properly designed, installed, operated, and maintained air distribution system, N  1.

18.2 AIR JETS An air jet is an airstream that discharges from an outlet with a signif cantly higher velocity than that of the surrounding air , and moves along its centerline until its terminal v elocity reduces to a v alue that equals or approximately equals the v elocity of the ambient air. Because of the turbulence of air particles, air jets tend to spread. They also rise or f all depending on the b uoyancy of the airstream. The outer boundary of an air jet where air is mo ving at a perceptible v elocity, such as 150 fpm (0.75 m / s), 100 fpm (0.5 m / s), or 50 fpm (0.25 m / s), is called the envelope. In general, air jets can be classif ed as free or conf ned, isothermal or non isothermal, and axial or radial. A free air jet is an ideal air jet whose en velope (outer boundary) is not con f ned by the enclosure of the conditioned space. A conf ned air jet ’s envelope is con f ned by the ceiling, f oor, walls, windows, and furniture of the conditioned space. According to experimental results, the characteristics of an air jet approach those of a free air jet when the relationship √Ar /Do  50 is satis f ed. Here Ar represents the cross-sectional area of the enclosure perpendicular to the centerline of the air jet, in ft 2 (m2), and Do indicates the diameter or the circular equivalent of the supply outlet, in ft (m). An isothermal air jet is one whose temperature is equal or nearly equal to the temperature of the ambient air. A nonisothermal air jet is one whose temperature is dif ferent from that of the ambient air in the conditioned space. An axial air jet projects in one direction, and a radial air jet projects radially in all directions.

Free Isothermal Jets Along the centerline of a free isothermal jet, four zones can be identif ed, as shown in Fig. 18.2: 1. Core zone. In the core zone, the centerline velocity remains unchanged. This zone extends about 4Do from the surface of the outlet.

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FIGURE 18.2 Four zones of a free, isothermal, axial air jet.

2. Transition zone . In the transition zone, the centerline v elocity decreases in versely with the square root of the distance from the surface of the outlet. This zone extends about 8Do. 3. Main zone . In the main zone, turbulent f ow is fully de veloped, and the maximum v elocity decreases in versely with the distance from the surf ace of the outlet. Ev en when the air jet is discharged from a rectangular outlet, the cross section of the airstream becomes circular in the main zone. This zone extends about 25 to 100Do in length. 4. Terminal zone. In the terminal zone, the maximum air v elocity decreases rapidly to a v alue less than 50 fpm (0.25 m / s) within a distance of a few outlet diameters. In core, transition, and main zones, the f uctuating velocity components transport momentum across the boundary. Therefore, ambient air is induced into the air jet, and the airstream di verges to a greater spread either v ertically or horizontally. The angle of di vergence 2  of a free isothermal jet discharged from a nozzle is about 22 °. For an air jet dischar ged from a slot, its angle of divergence 2 perpendicular to the slot is about 33 °. If the guide v anes at the outlet are de f ected at an angle from the straight position, the spread of the air jet is greater. The velocity prof les at different cross-sectional planes perpendicular to the air f ow in the main zone of a free isothermal jet are similar to one another . At any speci f c cross-sectional plane, the velocity prof le can be approximated by the formula

 RR   3.3 log 2

0.5

vc v

(18.7)

where v  air velocity at distance R from centerline of air jet, fpm (m / s) vc  centerline velocity, fpm (m / s) R  radial distance from centerline of air jet, ft (m) R0.5  radial distance from centerline of air jet to a point where velocity equals 0.5vc, ft (m) Research shows that the law of conservation of momentum can be assumed in the main zone of a free air jet. The application of this la w to determine the centerline v elocity of a free air jet gi ves the following results: vc K K K    vo x/Do x/Ho x/√Ao

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where K, K  centerline velocity constants, depending mainly on type of outlets; normally, K  1.13 K Ho  width of radial air jet at outlet or vena contracta, ft (m) vo  mean velocity at vena contracta, fpm (m / s)

(18.8)

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x  distance from surface of outlet to a cross-sectional plane having centerline velocity vc, ft (m)

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In Eq. (18.8), Ao, in ft2 (m2), represents the effective area of the airstream, i.e., the minimum area at the vena contracta. It can be calculated as vo  vcoreCd R fa

(18.9)

Ao  AcCd R fa

where Ac  core area of outlet, i.e., surface area of opening, ft2 (m2) vcore  face velocity at core of outlet, fpm (m / s) Cd  discharge coeff cient, usually between 0.65 and 0.9 Rfa  ratio of free area to gross area (free area is net area of opening through which air can pass) When vo is 500 to 1000 fpm (2.5 to 5 m / s), for round free openings, K  5; for rectangular openings and linear slot diffusers, K  4.9. When vo is 2000 to 5000 fpm (10 to 25 m / s), for round free openings, K  6.2. When multiple air jets are dischar ged into a conditioned space at the same le vel, each air jet behaves independently until the jets meet. From the point where the jets meet, the v elocities between the centerlines of the air jets increase until the y are equal to the centerline v elocities of the air jets.

Throw, Entrainment Ratio, and Characteristic Length Throw in ft (m), is def ned as the horizontal or v ertical axial distance from the outlet to a cross-sectional plane where the maximum v elocity of the airstream at the terminal zone has been reduced to 50 fpm (0.25 m / s), 100 fpm (0.5 m / s), or 150 fpm (0.75 m / s). The throw is indicated by Tv, and the subscript denotes the terminal velocity for which the throw is measured. For instance, T50 (T0.25) indicates the throw with a terminal velocity of 50 fpm (0.25 m / s). From Eqs. (18.8) and (18.9), throw can be calculated as Tv 

KV˙s

(18.10)

vt, max√AcCd R fa

where vt,max  maximum velocity of airstream at terminal zone, fpm (m / s) V˙s  supply volume f ow rate at outlet, cfm (m3 /s) For a specif c conf guration of supply outlet, throw Tv depends on both the supply v olume f ow rate V˙ s and supply outlet velocity vcore. The entrainment ratio Ren is the ratio of the v olume f ow rate of the total air at a speci f c crosssectional plane of the air jet V˙ x to the volume f ow rate of the supply air V˙ o discharged from the outlet ,which is sometimes called primary air. Total air is the sum of supply air and induced air. The entrainment ratio is proportional to the distance or square root of the distance from the outlet. For circular jets in the main zone, the entrainment ratio can be calculated as R en 

V˙x 2vo 2x   vc V˙o K√A o

For a long slot, the entrainment ratio is R en 

V˙x  V˙o

√

2x  KHo

√

2vo vc

(18.11)

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FIGURE 18.3 Surface effect.

The characteristic length L, in ft (m), is the horizontal distance from the surf ace of the outlet to the nearest v ertical opposite w all or the horizontal distance from the surf ace of the outlet to the midplane between tw o outlets in the direction of air f ow, or the distance to the closest intersection of air jets. The ratio of thro w to characteristic length Tv /L is related to the ADPI of various supply outlets and has been used as a parameter in space diffusion design.

Confined Air Jets In actual practice, most air jets are con f ned by the boundary of the room or the conditioned space. For a con f ned air jet, the total momentum of the f uid elements decreases gradually as distance from the outlet increases because of the friction between the airstream and the boundary. Surface Effect. When a primary airstream dischar ged from a supply outlet f ows along a surf ace, the velocity of the primary airstream is signi f cantly higher than that of the ambient air , and a lowerpressure region is formed near the surface along the airf ow, as shown in Fig. 18.3. Consequently, in-

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FIGURE 18.4 Centerline velocities of conf ned air jets.

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duced ambient air at a comparati vely higher pressure presses the air jet against the surf ace, even when it is a curved surface. Such a phenomenon is called the surface effect or the Conda effect. Friction between the air jet and the boundary decreases the centerline v elocity of con f ned air jets, as shown in Fig. 18.4. However, because of the surface effect, the throw of a conf ned air jet is longer, and the drop from the horizontal axis smaller, than that of a free air jet.

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Airfl w Pattern. The f ow pattern and characteristics of a con f ned air jet using a side wall outlet were introduced by Russian scientists during the 1950s. In an air conditioned room whose supply outlet is located above the occupied zone and whose exhaust opening is on the same side of the supply outlet, the supply air jet clings to the surf ace of the ceiling and mix es with the room air. An induced reverse airstream, with more e ven velocity and temperature distrib ution than that of the air jet, covers the occupied zone. Figure. 18.5 shows the airf ow pattern of a typical con f ned isothermal jet. When supply air is discharged from the circular outlet and mo ves along the surf ace of the ceiling, the f uctuating velocity components continue to transport momentum across the boundary of the air jet. Therefore, ambient air is induced into the air jet, and the induced circulating airf ow occupies most of the enclosed space. As the air jet moves forward, its mass f ow rate increases and mean air velocity decreases until it arrives at a cross-sectional plane where dimensionless distance s  0.22. Term s is def ned as ax (18.12) 0.5A √ r where a  turbulence factor; for a circular nozzle a  0.076, and for a rectangular outlet without guide vanes a  0.15. s

FIGURE 18.5 Airf ow pattern of a typical conf ned air jet.

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The air jet terminates be yond s  0.22, and the airstream mak es a 180 ° turn, forming a reverse airstream f owing in the opposite direction. The majority of the re verse airstream turns upw ard and is induced by the air jet again. Only a portion of it is exhausted outside the room. The angle of divergence, the velocity prof le, and the calculation of the entrainment ratio of conf ned jets are similar to those of free jets. Based on the principle of continuity of mass at steady state, other characteristics of conf ned air jets can be summarized as follows: ●

●

●

●

If there is no in f ltration into or exf ltration from the room, the mass f ow rate of supply air is e xactly equal to the exhaust air. If there are no obstructions in the room, the ideal stream lines of the induced air are closed curv es and form recirculating f ow. If the supply outlet and e xhaust inlet are located on the same side of the room, at an y crosssectional plane perpendicular to the horizontal air f ow, the mass f ow rates of airstreams f owing at opposite directions must be equal. The volume f ow rate of induced air is equal to or se veral times greater than that of supply air at the cross-sectional plane where s  0.22. The characteristics of airstreams in the occupied zone depend mainly on the induced re verse airstream. The volume f ow rate and the air v elocity of the reverse airstream are highest at the cross-sectional plane where s  0.22.

Free Nonisothermal Jets When the temperature of the conditioned air dischar ged from a supply outlet is dif ferent from that of the ambient air in a conditioned space, the buoyancy of the f uid particles causes the trajectory of the air jet to deviate from the axis of the free isothermal jet. A cold air jet will descend, and a warm air jet will ascend. Figure 18.6 shows a free cold air jet discharged horizontally from a nozzle.

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FIGURE 18.6 Path of a cold air jet discharged horizontally from a nozzle.

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In Fig. 18.6, x is the horizontal distance between a cross-sectional plane in a cold air jet and the surface of the outlet, and y is the drop of the cold air jet, i.e., the vertical distance between the horizontal axis of the nozzle and the center of the speci f c cross-sectional plane. The velocity prof le at the specif c cross-sectional plane is indicated by the solid line, and the temperature pro f le is indicated by the dotted line. According to the e xperimental results of K ostel (1955), the following empirical formula can be used to determine the v ertical drop of a cold air jet and the v ertical rise of the w arm air jet discharged from a nozzle: y

√Ao



x

√Ao

tan KAr

 √xA

o

cos

3



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(18.13)

where   angle between centerline of nozzle and horizontal axis of nozzle, deg K  constant Also Ar, known as Archimedes’ number, indicates the buoyant forces of the air jet and is given as Ar 

g√Ao To v 2o TrR

(18.14)

where g  gravitational acceleration, ft / s2 (m / s2) To  temperature difference between supply air at outlet and ambient air in conditioned space, °F (°C) TrR  absolute temperature of space air, °R (K) vo  air velocity at supply outlet, f/s (m/s) For free air jets, Kostel (1955) determined that K  0.065. For a nonisothermal jet, the relationship between the decay of the centerline v elocity and the centerline temperature difference Tc, in °F (°C), can be shown as Tc Tc  Tr 0.8vc   To Ts  Tr vo

(18.15)

where subscript c represents centerline, o the outlet, and r the space. The characteristics of isothermal radial air jets are similar to those of axial air jets. When determining the centerline velocity, use x / Ho for radial jets instead of x / √Ao for axial jets.

18.3 SUPPLY OUTLETS AND RETURN INLETS The proper type of supply outlet for a conditioned space is lar gely determined by the architectural setup of the room, the airf ow pattern needed, the indoor environmental requirements, and the load conditions. Five types of supply outlets are currently used: grilles and re gisters, ceiling diffusers, slot diffusers, light troffer diffusers, and nozzles. A window sill outlet is actually a type of grille mounted at the top of a fan coil.

Grilles and Registers A grille or grill, as shown in Fig. 18.7 a, is an outlet for supply air or an inlet for return air or e xhaust air. A register is a grille with a volume control damper. Figure 18.7b shows a single-def ection grille. It consists of a frame and one set of adjustable v anes. Either vertical or horizontal vanes may be used as face vanes to def ect the airstream.

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FIGURE 18.7 Supply grille and re gister: (a) grille, front view; (b) single-def ection, vertical vanes; (c) double-def ection register, with vertical and horizontal vanes.

Figure 18.7c shows a double-de f ection register. It is able to de f ect the airstream both horizontally and v ertically. A volume damper is also used to adjust the v olume f ow through the re gister. Usually extruded aluminum vanes, aluminum or steel frames, and steel dampers are used. A baked enamel f nish gives the grille an attractive appearance. Grilles have a comparatively lower entrainment ratio, greater drop, longer throw, and higher air velocities in the occupied zone than slot and ceiling diffusers. In the manufacturer’s catalog, the performance of a grille or re gister is def ned by the following parameters: ●

●

●

●

●

●

Core size or core area Ac, which indicates the total plane area of an opening, ft2 (m2) Volume f ow rate V˙ , cfm (m3 /s, or L / s) Air velocity v  V˙ / Ak, fpm (m / s) Total pressure loss pt, in. WC (Pa) Throw at terminal v elocities of 50 fpm (0.25 m / s), 100 fpm (0.5 m / s), and 150 fpm (0.75 m / s) with horizontal vanes def ected at 0°, 22.5°, and 45° Noise criteria curve

Here Ak is the area f actor representing the net or corrected area, in ft 2 (m2), which is the unobstructed area of a grille, register, or diffuser measured at the vena contracta. For Air Diffusion Council – certif ed performance data, throw is tested on isothermal jets in order to a void the in f uence of b uoyancy forces at dif ferent supply temperature dif ferential Ts. During testing, the difference between supply and room air temperatures cannot exceed  2 °F (1.11°C). Ceiling Diffusers

SH__ ST__ LG__ DF

A ceiling diffuser consists of a series of concentric rings or inner cones made up of v anes arranged in f xed directions and an outer shell or frame, as shown in Fig. 18.8 a. Ceiling dif fusers can be round, square, or rectangular. Square dif fusers are most widely used. Supply air is dischar ged through the concentric air passages or directional passages in one, two, three, or in all directions by using different

AIR SYSTEMS:

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18.13

..(>

(b)

(c)

FIGURE 18.8 Ceiling diffusers: (a) square and rectangular ceiling diffusers; (b) removable inner-core ceiling diffuser; (c) perforated ceiling diffuser.

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types of inner cone and v anes. The air dischar ge pattern of man y ceiling dif fusers may be changed from horizontal to vertical by means of adjustable inner cones or special de f ecting vanes. Ceiling diffusers are often mounted at the center of the conditioned space and discharge air in all directions. An important characteristic of a ceiling dif fuser is its induction ef fect. Induction is the v olume f ow rate of space air induced into the outlet by the primary airstream. It is also called aspiration. The ratio of the v olume f ow rate of induced air to the v olume f ow rate of supply air V˙ ind /V˙ s, both in cfm (L / s), is called the induction ratio. A ceiling diffuser with a higher induction ratio (see Fig. 18.8b) is bene f cial for a higher supply temperature dif ferential Ts for cold air distrib ution in an ice storage system, which is discussed later in Sec. 18.5 and in Chap. 31. Figure 18.8c shows a perforated ceiling diffuser. A perforated surface increases the induction effect. It is also used to match the appearance of the ceiling tile. The diameter of the holes on the perforated panel is typically 163 in. (4.8 mm). Some ceiling dif fusers are designed to dischar ge air evenly in all directions so that the surrounding ceiling surfaces are less likely to be smudged. The size of a ceiling diffuser is determined by its neck size and is therefore closely related to the neck v elocity of the dif fuser. Ceiling dif fusers are made of an outer shell and a remo vable inner core. The opposed-blade damper with an equalizing de vice of parallel blades mounted within a frame is often used to adjust the volume f ow and to provide a more uniform airf ow. Ceiling diffusers are usually made of aluminum-coated hea vy-gauge steel, extruded aluminum, or teak wood. In addition to high induction ratios, ceiling diffusers have a good surface effect and a shorter throw and can supply air in all directions. Horizontally projected ceiling dif fusers are suitable for conditioned spaces with low headroom. The performance data of a ceiling diffuser are def ned by the following parameters: ●

●

●

●

●

Neck velocity vneck, that is, mean air velocity at the neck of the ceiling diffuser, fpm (m / s) Volume f ow rate V˙ s  Ak vs, in cfm (L / s); area factor Ak, in ft 2 (m 2), is determined at the testing laboratory, and the a verage supply v elocity vs is usually measured at a speci f ed position at the outlet by a def ecting vane anemometer, often called a velometer Total pressure loss pt, in in. WC (Pa) Throw, ft (m) Noise criteria curve

Slot Diffusers

SH__ ST__ LG__ DF

A slot dif fuser consists of a plenum box with single or multiple slots and air de f ecting vanes, as shown in Fig. 18.9. The slots are a vailable in 0.5-in. (13-mm), 0.75-in. (19-mm), 1-in. (25-mm), and 1.5-in. (38-mm) widths. The 0.75- and 1-in. slot widths are most commonly used. Slots are typically 2, 3, 4, 5, or 6 ft (600, 900, 1200, 1500, or 1800 mm) long. Air discharged from a slot diffuser can be projected horizontally or vertically. With a single-slot diffuser, air is always discharged in one direction. With multiple slots, air can be horizontally dischar ged either left or right, or a combination of both, or one slot can discharge vertically while another discharges horizontally. Slot diffusers are often called T-bar diffusers, as they are compatible with standard ceiling T bars. Air enters the plenum box from the round inlet at one side. The function of the plenum box is to distribute the air more e venly at the slot. The plenum box is sometimes insulated internally . After passing the de f ecting vanes, the air is dischar ged from the slot horizontally (left or right) or v ertically according to the required air f ow pattern. Figure 18.9 sho ws the airstreams dischar ged from slot dif fusers in v arious directions. Dif ferent manuf acturers use their o wn patented air de f ecting vanes. Def ecting vanes can also be used to block the airf ow entirely if necessary. Extruded aluminum is widely used as the material for the structural member of the slot, and a galvanized sheet is used for the plenum box. The performance of a slot dif fuser is de f ned by the following parameters: ●

Width of slot, number of slots, and length of slots

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Air deflecting vanes

Slot width (a)

(b) Innerlined material

Plenum box Inlet

Ceiling (c)

(d)

FIGURE 18.9 Slot dif fusers: (a) horizontal left; ( b) horizontal right; (c) vertical; (d ) two-slot, two-way.

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TABLE 18.1 Performance Data for Typical 0.75-in. (19-mn) Width Slot Diffuser with Plenum Slots

CHAPTER EIGHTEEN

Total pressure, in.WC

H

.004

.015

.032

.058

.091

.125

.175

.230

V

.003

.011

.024

.044

.067

.095

.125

.170

6

12

18

24

30

36

42

48

1,1,3

1,3,9

3,6,11

5,9,13

8,10,15

9,11,16

10,12,17

11,13,19

2

6

10

12

14

15

16

17

—

—

—

18

24

30

35

39

Cfm / Ft H

Throw, ft

1

V

NC Cfm / Ft 2

H

Throw, ft

3

pg 18.16

V

12

24

36

48

60

72

84

96

1,1,5

2,5,13

5,10,16

9,13,19

11,15,21

13,16,23

14,17,25

15,19,27 24

4

9

14

17

19

21

23

NC

—

—

13

23

29

35

40

44

Cfm / Ft

18

36

54

72

90

108

126

144

H

1,2,8

4,8,16

7,12,20

11,16,23

14,18,26

16,20,28

18,22,30

19,23,32

V

6

11

17

21

23

26

28

30

—

—

16

26

32

38

43

47

Throw, ft NC

Source: TITUS. Reprinted by permission.

●

●

The volume f ow rate per ft (m) of slot, or slot intensity, in cfm / ft (L / s m) Throw, total pressure loss, and noise criteria curve

Table 18.1 lists the performance data of a typical slot dif fuser of 0.75-in. (19-mm) slot width, taken from a manufacturer’s catalog.

Nozzles A nozzle is a round supply outlet (see Fig. 18.10). The airstream discharged from a nozzle is contracted just before the outlet, which results in a higher f ow velocity and more even distribution. The purpose of using a nozzle instead of other types of supply outlets is to pro vide a longer throw and a smaller spread. Round nozzles (Fig. 18.10 a) are usually used in air dif fusion systems with lar ge

Supply air 40 F (4.4 C)

6 in. (150 mm) Induced air

SH__ ST__ LG__ DF

FIGURE 18.10

Round nozzle and nozzle diffuser (a) Round nozzle and (b) nozzle diffuser.

Ceiling

6 in. (150 mm) (b)

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space volume or in spot cooling / heating, which is co vered in detail in Sec. 18.9. Nozzle dif fusers (Fig. 18.10b) have a high induction ratio and are used in cold air distrib ution, which is discussed in Sec. 18.5.

__RH TX

Accessories for Supply Outlets Grilles, registers, and diffusers may require certain accessories to modulate the v olume f ow rate of each supply outlet and to distrib ute the air more e venly at the outlet. Currently used accessories include ●

●

●

Opposed-blade dampers. Usually the y are attached to the re gister upstream from the v anes, or installed at the outlet collar leading to a ceiling diffuser, as shown in Fig. 18.11a. Split dampers . A piece of sheet metal is hinged at one end to the duct w all, as sho wn in Fig. 18.11b. An equalizing grid is often used at the same time to distrib ute the supply air more evenly to the outlets. A split damper is also often installed at the junction of tw o main ducts, to adjust the volume f ow rates of the supply air between these ducts. Gang-operated turning vanes. These vanes are always installed before the takeoff to the outlet, as shown in Fig. 18.11c. The tips of the turning vanes through which air enters the takeoff remain parallel to the direction of the air f ow. Gang-operated turning v anes are especially suitable for adjusting the air f ow rate for re gisters that are mounted on the air duct with a v ery short outlet collar.

Return and Exhaust Inlets Various types of return inlets are used to return space air to the air-handling unit or packaged unit in the fan room. Exhaust air inlets are usually connected to the e xhaust duct. Before contaminated air is exhausted outside, it is usually cleaned and treated with special air cleaning equipment.

FIGURE 18.11 Accessories for supply outlets: (a) opposed-blade damper; (b) split damper with equalizing device; (c) gang-operated turning vanes.

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FIGURE 18.12 Various types of return grilles and re gisters: (a) return grille with curv ed vanes; ( b) return register; (c) wire-mesh grille; (d ) f oor-type return grille; (e) door louvers.

SH__ ST__ LG__ DF

Return and e xhaust inlets can be classi f ed as return or e xhaust grilles, return or e xhaust registers, return slots, and ventilated light troffers. Return grilles and re gisters (see Fig. 18.12) are similar in shape and construction to supply grilles and registers. In Fig. 18.12a, only one set of vanes is required to direct the air into the return duct, so vision-proofed curved vanes are often used instead of the straight vanes. Vanes can be f xed or adjustable. An opposed-blade damper is sometimes installed behind the curv ed vanes in a return register, as shown in Fig. 18.12 b. Wire-mesh grilles, as shown in Fig. 18.12 c, are simple and ine xpensive. They are often installed in places where the appearance of the grille is not important. In many industrial b uildings, f oor grilles are used when the return duct is belo w the f oor level, as shown in Fig. 18.12d. When room air is required to return to the f an room through a corridor and a return duct, a door louv er is often used to transport return air through the door . Vision-proofed, right-angle louvers are generally used for this purpose, as shown in Fig. 18.12e. Local codes should be consulted before the door louvers are installed. When a ceiling plenum is used as a return plenum (as in man y commercial b uildings), return slots are used to dra w the return air through the ceiling. Return slots are similar to slot dif fusers. They are usually made with single or double slots, as shown in Fig. 18.13. The width of the slot

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FIGURE 18.13 Return slots: (a) single-slot; (b) double-slot.

varies from 0.5 to 1.75 in. (13 to 45 mm). Slots are usually a vailable in lengths of 2, 2.5, 3, 4, and 5 ft (600, 750, 900, 1200, and 1500 mm). The capacity of return slots is also e xpressed in cfm / ft (L / s m) of slot length. If a return slot has the same con f guration, slot width, and slot intensity as the supply slot dif fuser, then its total pressure loss, in in. WC (Pa), is nearly the same as the v ertical projection total pressure loss of the supply slot dif fuser. Return slots are usually inner -lined with sound absorption material. This material pre vents sound transmission between tw o rooms via the return slots and the ceiling plenum. Exhaust inlets are similar to the return inlets but they often have a simpler design. In return and e xhaust inlets, air velocity decreases very sharply as the distance from the surf ace of the inlet increases. This phenomenon is the major dif ference in airf ow characteristics between a return or an exhaust inlet and a supply outlet.

Light Troffer Diffuser and Troffer-Diffuser Slot Light trof fer dif fuser combines a f uorescent light trof fer and slot dif fuser, sometimes in saddle type, as shown in Fig. 18.14 a. The slot can be used as a supply outlet or return inlet. A troffer-diffuser slot is a combination of light trof fer, slot diffuser, and return slot, as shown in Fig. 18.14 b. Light troffer diffusers and troffer-diffuser slots serve two main purposes: 1. To maintain a lo wer air temperature in the light trof fer, which increases the luminous ef f ciency of f uorescent lamps 2. To form an integrated layout of light troffer, diffuser, and return slots on suspended ceilings In addition to these bene f ts, a combination of a light trof fer and return slot also reduces the space cooling load because return air absorbs part of the heat released by the f uorescent lights. In a light troffer diffuser, surface slots are often used. Air can be discharged horizontally or vertically. Troffers and slots are available in 2-, 3-, and 4-ft ( 600-, 900-, and 1200-mm) lengths. To prevent the deposit of dust particles on the surf ace of the f uorescent tube, the light trof fer should be designed so that return air does not come in direct contact with the f uorescent tube. The result is a high ventilation eff ciency and a neat appearance.

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FIGURE 18.14 Light trof fer, slot dif fuser, and return slot combination: (a) light troffer diffuser; (b) troffer-diffuser slot.

18.4 MIXING FLOW Airflow Pattern and Space Air Diffusion Airf ow pattern determines the performance of space air dif fusion in the occupied zone of commercial buildings or in the working area of a factory. The optimum airf ow pattern for an occupied zone depends mainly on the indoor temperature, relative humidity, and indoor air quality requirements; the outdoor air supply; and the characteristics of the building. Four types of airf ow patterns are currently used in air conditioned space: 1. 2. 3. 4.

Mixing f ow Displacement f ow Projecting f ow Upward f ow

Principles and Characteristics of Mixing Flow SH__ ST__ LG__ DF

The airf ow pattern of an air conditioned space is said to be in mixing f ow only when the supply air is thoroughly, or nearly thoroughly, mixed with the ambient air , and the occupied zone or w orking area is dominated by the induced recirculating f ow.

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In many commercial and industrial b uildings, cold supply air at 50 to 60 °F (10 to 15.6 °C) and a velocity of 400 to 800 fpm (2 to 4 m / s) must be nearly thoroughly mix ed with the ambient air f rst. The temperature of the mixture is raised to 72 to 74 °F (22.2 to 23.3 °C), and its v elocity drops to less than 70 fpm (0.35 m / s) before it enters the occupied zone and w orking area, or drafts occur. Mixing f ow is often the best choice for comfort air conditioning systems when the supply temperature differential is greater than 15 °F (8.3°C) and the supply air v elocity exceeds 300 fpm (1.5 m / s). Straub and Chen (1957) tested mixing f ow patterns by using f ve types of supply outlets. These airf ow patterns have the following characteristics:

__RH TX

Induction of Space Air to the Air Jet. Mixing of induced air and the air jet reduces the dif ferences in air velocity and temperature between the air jet and the ambient air to acceptable limits. Reverse Airstream in the Occupied Zone. The induced re verse airstream co vers the occupied zone with a more even temperature and velocity distribution. Minimize the Stagnant Area in the Occupied Zone. A stagnant area is a zone with natural convective currents. Air velocities in most of the stagnant zones are lo wer than 20 fpm (0.1 m / s). Air stratif es into layers, with a signif cant temperature gradient from the bottom to the top of the stagnant zone. If the recirculating f ow f lls the entire occupied zone, no space remains for stagnant air . Types and Locations of Return and Exhaust Inlets. Because the return or e xhaust airf ow rate is only a fraction of the maximum total air f ow of the supply air jet or the maximum air f ow of the reverse airstream, the location of the return or e xhaust inlet does not signi f cantly affect the air f ow pattern of mixing f ow in the occupied zone. Ho wever, the location of the return and e xhaust inlets does af fect the thermal ef fectiveness f actor  T of space air dif fusion. Therefore, for an induced recirculating f ow pattern, the return intake should be located near the end of the re verse airstream, in the vicinity of a heat source, or within the stagnant zone. For a conditioned space in which the space air is contaminated by dust particles, toxic gases, or odors, several e venly distrib uted return or e xhaust grilles should be installed. Collecti ve return grilles, i.e., one or two large return inlets instead of man y inlets, impair indoor air quality. They are only suitable under the following circumstances: ●

●

●

When the conditioned space is small When it is diff cult to provide a return duct When lower initial cost is of prime importance

Types and Locations of Supply Outlets. The type and the location of supply outlets af fect the mixing f ow pattern and its performance. Currently, mixing airf ow incorporates the following types of supply outlets: ●

●

●

●

●

High side outlets Ceiling diffusers Slot diffusers Sill and f oor outlets Outlets from stratif ed mixing f ow

Mixing Flow Using High Side Outlets A high side outlet may be a side wall outlet mounted at the high le vel of a conditioned space, or an outlet mounted directly on the supply duct. Figure 18.15 sho ws a mixing f ow using a high sidewall

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FIGURE 18.15 Fundamentals.)

SH__ ST__ LG__ DF

Mixing f ow using high side outlets. (Adapted with permission from ASHRAE Handbook 1989,

outlet. In Fig. 18.15, the shaded black en velope shows the primary air en velope of the air jet, and the dotted envelope shows the total air envelope. As the air jet discharges from the high sidewall outlet, the surface effect tends to keep the air jet in contact with the ceiling. During cooling, the air jet induces the ambient air and de f ects downward when it strik es on the opposite w all. As soon as the maximum air v elocity of the air jet drops to about 50 fpm (0.25 m / s), the air jet terminates as it f ows along the opposite wall. The induction of space air from the occupied zone into the air jet forms the reverse airstream and f lls the occupied zone. If the thro w of a high side wall outlet is longer than the sum of the length of the room and the height of the opposite wall, then the air jet is def ected by the opposite wall and f oor and enters the occupied zone with excessive air velocity. Both the air jet and the reverse airstream f ll the occupied zone with a higher air velocity and greater temperature difference. If the throw is too short, the air jet drops and may enter the occupied zone directly . This causes higher air velocity and lower temperature (cold draft) in the occupied zone. During heating, warm air tends to rise and results in a shorter thro w. As the induced airstream rises, a stagnant zone may form between the f oor and the induced airstream. A warm air jet with suff cient velocity and a longer throw can reduce or even eliminate the stagnant zone. If the airstream at the terminal zone of the w arm air jet comes in contact with the cold inner surface of the window, it converts to a cold airstream and forms a cold draft when it f ows downward and enters the occupied zone. The installation of a baseboard heater under the windo w sill can prevent the formation of this cold draft and reduce the vertical temperature gradient in the occupied zone. The most suitable location for a return inlet is on the ceiling outside the air jet. Return air f ows through the return inlets and often uses the ceiling plenum as a return plenum. Such an arrangement often results in a higher effectiveness factor T and a lower installation cost for the return system. For a room without a suspended ceiling, a cold or w arm air jet has a shorter thro w because of lack of surface effect. The buoyancy force causes a cold air jet to drop sooner , and it may enter the occupied zone directly . The higher rise of a w arm air jet forms a greater stagnant zone abo ve the f oor level.

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FIGURE 18.16 Fundamentals.)

Mixing f ow using ceiling diffusers. (Adapted with permission from ASHRAE Handbook 1989,

Mixing f ow using high side outlets usually gi ves a longer throw and a higher air v elocity in the occupied zone than other supply outlets. Drop must be check ed to prevent the air jet from entering the occupied zone directly in a large conditioned space.

Mixing Flow Using Ceiling Diffusers The air f ow pattern of mixing f ow using ceiling dif fusers is similar to that of high side outlets. Ceiling diffusers produce a better surface effect, a shorter throw, a lower and more even distribution of air velocity, and a more even temperature in the occupied zone. Figure 18.16 shows mixing f ow using ceiling dif fusers. During cooling, the cold air jet usually maintains contact with the ceiling, producing a better surf ace effect than side wall outlets. The reverse airstream f lls the occupied zone. During heating, thinner vertical spread and shorter thro w create a larger, higher stagnant zone. The best location for return inlets is on the ceiling within the terminal zone of the air jet. When two air jets discharge horizontally from ceiling diffusers or other outlets in opposite directions along the ceiling, with a ceiling height of 10 ft (3.05 m), the air v elocity where tw o air jets collide should be less than 120 fpm (0.6 m / s) in order to maintain an air v elocity below 50 fpm (0.25 m / s) 5 ft (1.5 m) abo ve the f oor. Mixing f ow using ceiling diffusers is widely used for conditioned spaces with limited ceiling height. Man y ceiling dif fusers are designed to ha ve a high

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FIGURE 18.17 heating.

Mixing f ow using slot diffusers: (a) perimeter and interior zone, cooling; (b) perimeter zone,

induction ratio and to produce a v ery good surf ace effect. These ceiling diffusers are often used in variable-air-volume systems.

Mixing Flow Using Slot Diffusers Slot dif fusers are narro wer and longer than ceiling dif fusers and grilles. When supply air is discharged horizontally from a ceiling slot dif fuser, it has a thinner v ertical spread and an e xcellent surface effect. Figure 18.17 shows mixing f ow using slot diffusers discharging vertically downward into the perimeter zone and also discharging horizontally along the ceiling in both the perimeter and interior zones of a large off ce. Because of its superior surf ace effect, the horizontally discharged cold air jet remains in contact with the suspended ceiling e ven when the supply air v olume f ow rate is reduced in a v ariable-airvolume system. The occupied zone is then f lled with the re verse airstream at a more uniform temperature and air velocity. In order to counteract natural con vection along the inner surf ace of a windo w in the perimeter zone, the air jet is often projected do wnward toward the window. A suff cient throw is important to produce a downward air jet that offsets the cold draft at the inner surface of the window during winter heating. The occupied zone f lls with induced air f ow at an air v elocity lower than 50 fpm (0.25 m / s). For a mixing f ow using slot dif fusers, the location of return slot is preferably aligned with the supply slots. In addition to their e xcellent surface effect in mixing f ow, slot diffusers have a linear appearance that can be coordinated easily with electric lights and ceiling modular arrangements. Therefore, they are widely used in b uildings using VAV systems with moderate loads and normal ceiling height.

Mixing Flow Using Sill or Floor Outlets

SH__ ST__ LG__ DF

The purpose of using a sill or f oor outlet is to counteract the cold draft f owing downward along the inner surface of the windo w when the outdoor temperature is belo w 30°F ( 1.1°C). Figure 18.18 shows a mixing f ow using a sill outlet. During cooling, a stagnant zone may form abo ve the eddies be yond the terminal zone of the air jet if the supply air v elocity is not high enough. Meanwhile the space under the cold air jet is f lled

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FIGURE 18.18 Mixing f ow using sill outlet.( ASHRAE Handbook 1989, Fundamentals.)

Adapted with permission fr

om

with reverse airstreams and cooled. During heating, a stagnant zone may form belo w the ascending induced airstream. Mixing f ow with sill or f oor outlets has been widely used in b uildings with large window areas or in raised- f oor off ces. The direction and amount of air f ow from f oor diffusers usually can be adjusted according to the requirements of the occupants. Stratified Mixing Flow In a b uilding with a high ceiling, it is more economical to stratify the b uilding vertically into tw o zones (the strati f ed upper zone and the cooled lo wer zone) or three zones (upper , transition, and lower zones) during cooling. The upper boundary of the lower zone is at the level of the supply outlet where the air jet projects horizontally. Gorton and Sassi (1982) and Bagheri and Gorton (1987) performed a series of model studies and experiments about strati f ed space air dif fusion. Figure 18.19 a shows the elevation view of the nuclear reactor f acility in which the e xperiments were conducted. The ceiling height w as 41 ft (12.5 m). An air -handling unit w as located at f oor le vel, and tw o dif fusers were mounted on the supply riser 16 ft (4.9 m) abo ve the f oor le vel. The supply v elocity at the outlet w as about 1000 fpm (5 m / s), and the supply temperature was around 60°F (15.6°C). Figure 18.19b shows the space air temperature pro f le at 4 p.m. at v arious height le vels during cooling. This temperature prof le could be divided into ●

●

An upper zone, from the ceiling do wn to about 22 ft (6.7 m) from the f oor level, in which air temperature varied from 82.5 to 79°F (28 to 26.1°C) A transition zone, between 16 and 22 ft (4.9 and 6.7 m) from the f oor, in which air temperature varied from 79 to 74.5°F (26.1 to 23.6°C)

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FIGURE 18.19 Stratif ed mixing f ow in a nuclear reactor facility: (a) elevation view of a nuclear reactor f acility; (b) temperature prof le during cooling; (c) temperature prof le during heating.

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FIGURE 18.20

●

Stratif ed mixing f ow in a large indoor stadium using supply nozzles.

A lower zone belo w 16 ft (4.9 m), in which air temperature v aried between 72 and 72.5 °F (22.2 and 22.5°C)

The induced recirculating f ow under the cold air jet formed the lo wer zone. Most of the lo wer zone was f lled with re versed airstreams. The upper zone and the transition zone were formed because of the entered heat gains at the upper le vel and a weaker recirculating induced f ow above the cold air jet. Because of the drop of the cold air jet, a greater amount of recirculating f ow was induced in the cold air jet in the lower zone than in the upper zone. Consequently, most of the cooling occurred in the lo wer zone. The formation of a stagnant layer near the ceiling also retarded air f ow and heat transfer between outdoor and indoor air in the upper zone. Figure 18.19c shows the space temperature prof le at various height levels at 4 p.m. during heating. The supply air temperature w as about 85 °F (29.5°C) and the outside temperature 40 °F (4.4°C). The air temperature v aried from 80 to 81.5 °F (26.7 to 27.5 °C) from a height of 0 to 36 ft (0 to 11.0 m) from f oor level. There was no evidence of signif cant thermal stratif cation. Figure 18.20 shows a stratif ed mixing f ow in a large, high-ceiling indoor stadium using supply nozzles. During summer cooling, the cold air jet creates tw o induced recirculating air f ows: a cooled lower occupied zone and a strati f ed upper zone. If a higher air v elocity is required in the lower occupied zone (the spectator ’s area), the supply nozzles should project at an inclined angle. Otherwise, the cold air jet should project horizontally. During heating, if the throw of the w arm air jet is long enough to arri ve at the lo west spectator seat, the whole spectator area is f lled with the re versed airstream. No stagnant zone is formed in the lower occupied zone. Return inlets should be located in the lo wer cooling zone. They should be evenly distributed under the spectator seats. Characteristics of the strati f ed mixing f ow during summer cooling mode operation are as follows: ●

●

●

●

●

Convective heat transfer from the hot roof is ef fectively block ed by the higher -temperature or stagnant air layer in the upper zone. Cooling loads that occur in the lo wer zone (windo ws, walls, lights, occupants, and equipment) should be offset by cold supply air. Radiant heat from the roof, upper external walls, and electric lights in the upper zone enters the occupied zone and converts to cooling load. Although supply air f ow rate and supply temperature af fect the throw and drop of the air jet, the induced recirculating airf ow pattern and the stratif ed upper and lower zones remain the same. The height of the supply air jet determines the upper boundary of the lower zone.

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The location of return inlets in f uences the cooling load only when the y are located in the upper or transition zone. A portion of exhaust air should be extracted at a higher level.

18.5 COLD AIR DISTRIBUTION Cold Air Distribution Versus Conventional Air Distribution Because of the lo wer chilled w ater dischar ge temperature of 34 to 38 °F (1.1 to 2.2 °C) from ice storage systems, the supply air temperature of the air system can often be reduced to 40 to 45 °F (4.4 to 7.2°F) and results in a cold air distrib ution (low-temperature distribution) system instead of a con ventional system of 55 °F (12.8 °F) supply temperature. Dor gan and Elleson (1993) and Scof eld (1993) summarized that a cold air distrib ution system has the follo wing advantages when compared with a conventional space air diffusion system: ●

●

●

●

●

●

●

Due to the low dew point of supply air, space relative humidity can be maintained between 35 and 45 percent. Cold air distrib ution can increase the space temperature slightly without causing discomfort. For a cold air distribution system using fan-powered VAV boxes, Ts is equal to the air temperature discharged from the f an-powered unit during summer cooling mode operation if the duct heat gain between the fan-powered VAV box and the supply outlet is ignored. If Ts  44°F (6.7°C) and space air temperature Tr  78°F (25.5 °C), then the supply temperature dif ferential for cold air distribution Tr  Ts  78  44  34°F (18.9°C). Compared with a supply temperature dif ferential Tr  Ts  20°F for a con ventional system, cold air distrib ution systems reduce the design supply volume f ow rate by 44 percent. With cold air distrib ution, fans, ducts, and the associated air system components can be do wnsized. Smaller ducts require less space clearance between the ceiling and structural members, and thus a reduction of f oor-to-f oor height in high-rise b uildings. Also there is a signi f cant increase in rental space because of the reduction of the mechanical rooms and duct shafts. Fan energy use can save as much as 40 percent. Savings are seen in construction cost because of the reduction of the structural height and smaller ductwork and air system components (sometimes it is the primary consideration). There are reduced fan sound levels and, therefore, less sound attenuation for a specif ed space NC rating. Cold air distrib ution is often bene f cial for a retro f t project in which the e xisting air distrib ution system may be used for increased space cooling loads.

The major reserv ation of using cold air distrib ution, especially for VAV systems in lar ge buildings, is the justif cation of a reduction of 40 to 45 percent of supply air at a time of increased concern over IAQ and the dumping of the cold air jet at reduced volume f ow.

High Induction Nozzle Diffusers

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Two types of space dif fusion systems are used in cold air distrib ution: (1) Cold supply air from the air-handling unit or packaged unit is supplied directly to the conditioned space by means of high induction diffusers; (2) fan-powered VAV boxes are used to mix low-temperature supply air with return air, and the mixture at a temperature between 55 and 60 °F (13 and 16 °C) is then supplied to the conditioned space.

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Fields and Knebel (1991) and Knebel and John (1993) introduced a high induction nozzle diffuser used in cold air distrib ution, as shown in Fig. 18.10 b. A typical 4-ft- (1.22-m-) long high induction nozzle diffuser has 57 0.75-in. (19-mm) nozzles on each side of the diffuser. At an airf ow rate of 350 cfm (165 L / s), it has a supply air v elocity of about 1000 fpm (5 m / s), a total pressure loss at the supply outlet of 0.213 in. WC (53 Pa), a throw T50  33 ft (T0.25  10 m), and a drop of about 0.8 ft (0.24 m).

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Characteristics of Cold Air Distribution Systems A satisfactory cold air distribution system often has the following characteristics: ●

●

●

●

A higher √Ar /Do value, as def ned in Sec. 18.2, to provide adequate surrounding spaces for the supply air jets to induce suf f cient ambient air and fully mix the ambient air with the 40 to 45 °F (4.4 to 7.2°C) low-temperature supply air A higher supply air v elocity and jet turb ulence (depending on the con f guration of the air dif fusion system) to provide momentum for the induced airstreams An excellent surface effect with adequate throw and small drop to pre vent supply air jet from entering the occupied zone directly An ADPI 80 percent at both design and reduced airf ow

Case Study — Florida Elementary School According to Knebel and John (1993), in a Florida elementary school, the HVAC&R system consists of a 40,000 cfm (18,880 L / s) blow-through AHU with airfoil f an and an ice storage system. The DDC controlled VAV terminals are used to throttle the airf ow and reset the supply air temperature. The classroom that uses cold air distrib ution is 27 ft (8.3 m) wide by 36 ft (11.0 m) long. Three 4-ft nozzle-dif fusers, each with 114 nozzles, were set in one ro w, designed to produce a throw of 18 ft (5.4 m) on each side. Each nozzle-dif fuser had a design air f ow rate of 200 cfm (94 L / s) and a supply velocity of 500 fpm (2.5 m / s) at the nozzle outlet. With this velocity, the 50 fpm (0.25 m / s) air jet envelope was 30 in. (0.76 m) from the nozzle. The air motion in the space was between 0 and 10 fpm (0 and 0.05 m / s). The diffuser supplier modi f ed the number of acti ve nozzles in each dif fuser. The supply velocity at the nozzle outlets w as raised to 2000 fpm (10 m / s) at the design air f ow and to 1000 fpm (5 m / s) at 50 percent part-load operation. For an outlet velocity of 2000 fpm (10 m / s) and a supply air temperature of 40 °F (4.4 °C), space air v elocity had been raised to 15 to 45 fpm (0.075 to 0.23 m / s) and the measured ADPI was 100 percent. At part-load operation, when the outlet velocity dropped to 1000 fpm (5 m / s) with an supply air temperature of 41 °F (5°C), the space air v elocity varied from 10 to 40 fpm (0.05 to 0.20 m / s) and the measured ADPI w as 96.2 percent. The estimated √Ar /Do value for this cold air distribution system in this classroom is about 37. Performance of Ceiling and Slot Diffusers Zhang et al. (1994) conducted v entilation performance tests for tw o square diffusers (three concentric inner cones dischar ged in all directions, which is similar to the ceiling dif fuser shown in Fig. 18.8b) and one-slot dif fusers (horizontally projected) using 38 °F (3.3°C) supply air for cold air distrib ution and 55°F (12.8°C) supply air for conventional air distribution. Test results were as follows: ●

Cold and con ventional air distrib ution systems performed equally well in deli vering the outdoor ventilation air into the occupied zone.

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The average air exchange effectiveness (AEE) of the small square dif fuser using cold air distribution w as 1.01, and using con ventional air distrib ution w as 0.99. The a verage AEE for the slot diffuser using cold air distribution was 0.99, and using conventional air distribution was 1.00. The high magnitudes of AEE (  0.90) showed that almost all the supply air w as delivered into the occupied zone. The ADPI for the lar ge square diffuser using cold air distrib ution was 92.5 percent, for the small square diffuser was 96.7 percent, and for the slot dif fuser was 96.7 percent. The cold air distribution system for the slot dif fuser has a √Ar /Do value of 22. The ADPI for small square diffuser using conventional air distribution was 96.7 percent and for the slot diffuser was 90.0 percent. Cold air distrib ution with these three dif fusers pro vided satisf actory v entilation performance. There w as no appreciable dif ference between the cold and con ventional air distrib ution systems.

Cold Air Distribution with Fan-Powered VAV Boxes There are two types of fan-powered VAV box used in cold air distribution: parallel fan-powered box and series fan-powered box. In a parallel fan-powered box, the cold primary air from the AHU does not f ow through the VAV box fan but instead mixes with the recirculating air after the box f an outlet, whereas in a series f an-powered box, the cold primary air f ows through the box f an. Parallel and series f an-powered box es are discussed in Sec. 21.5. F or 44 °F (6.7 °C) cold primary air and 78°F (25.6°C) ceiling plenum air , 40 percent of 78 °F (25.6°C) return air is often e xtracted to the fan-powered box and is mix ed with 60 percent of 44 °F (6.7°C) cold primary air from the AHU, resulting at a 58 °F (14.4°C) mixture, the same as in the con ventional air distribution system supplied to the space. A fan-powered box has the benef ts of providing space air motion. Elleson (1993) compared the energy use with and without the fan-powered mixing boxes in cold air distribution. In general, the energy required for continuous mixing box f an operation is greater than the supply fan energy savings resulting from the reduced supply temperature. Box fan energy use can be reduced by operating the box f an only when space air motion f alls below a minimum le vel. Supply air temperature reset can further reduce f an energy consumption. Direct supply of cold primary air to the conditioned space entails the lo west fan energy use. Furthermore, compared to the direct supply of cold primary air, cold air distribution using fan-powered boxes has a higher noise le vel and initial cost and requires more site maintenance. Direct supply of low-temperature cold primary air to the conditioned space is recommended.

Surface Condensation

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For a space air temperature of 75°F (23.9°C) and a relative humidity of 50 percent, the dew point of space air is 55 °F (12.8°C). During cold air distrib ution, the space relative humidity can be reduced to 40 percent, and the dew point is then about 49.5°F (9.7°C). For a cold air distribution with a supply air of 40 °F (4.4°C), the temperature differential between the supply air inside the ducts and the ambient air increased to 75  40  35°F (19.4°C) instead of 75  55  20°F (11.1°C) in a conventional air distrib ution system. Harmon and Yu (1993) reported that the actual space relati ve humidity levels measured were 3 to 5 percent abo ve those projected by calculations. In cold air distribution systems, in order to pre vent surf ace condensation, the AHU or PU, ducts, duct access doors, VAV boxes, f exible ducts, and diffusers should be insulated with a layer of f berglass protected by a vapor barrier, or a closed-cell elastomeric or cellular glass insulation, with a thickness of 2 in. (50 mm). It is preferable to select VAV boxes with thermally isolated inlet connections and properly insulated diffusers. Plenum space must be properly sealed to pre vent outside air inf ltration. Gradually cool the supply air and the space air during each early morning in summer to form a soft start, so that the outside surface will not suddenly be cooled below the dew point of the ambient air.

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18.6 DESIGN PROCEDURE OF MIXING FLOW AIR DIFFUSION SYSTEM The supply volume f ow rates of various control zones at summer and winter design conditions are usually determined before a mixing f ow air dif fusion system is designed. Because cooling mode operations are used in both perimeter and interior zones in summer and often in interior zones e ven in winter, heating mode operations often occur only in the perimeter zone during winter . Therefore, cooling mode operation is the basic consideration in space air dif fusion design. In the perimeter zone, heating mode space air dif fusion must be considered, and a compromise may be necessary . The design procedure for such an air diffusion system follows.

Select the Type of Supply Outlet Selection of the supply outlet depends on the following: ●

●

●

●

●

Requirements of indoor en vironmental contr ol. If the conditioned space needs controlled air movements, such as for a badminton and table tennis tournament, or precise air temperature control, a high side outlet is not the right choice. Shape, size, and ceiling height of the b uilding. For buildings with limited ceiling height, ceiling and slot diffusers are often the best choice. F or large buildings with high ceilings, high side outlets mounted at high le vels to form strati f ed induced recirculating f ow patterns are recommended. In a perimeter zone, an overhead two-way slot diffuser projected down toward the window and horizontally projected to the room, a ceiling diffuser with a throw to the inner surface of the window glass, or a sill outlet should be used. Surface effect. A good surface effect is especially important to the VAV system because it allo ws the supply volume f ow rate to be reduced to half or even 30 percent of the design f ow. Volume flow per f 2 of floor a ea. Sidewall outlets are limited to a lo wer volume f ow per ft 2 of f oor area, in cfm / ft2 (L / s  m2), because of the higher air v elocity in the occupied zone. The slot diffuser has a narro wer slot width and can only project in one or tw o directions. Therefore, the volume f ow per ft 2 for a slot diffuser is smaller than that of a ceiling dif fuser. Table 18.2 lists the volume f ow per ft2 for various types of supply outlet recommended by ASHRAE Handbook 1992, HVAC Systems and Equipment. Appearance. The shape and con f guration of outlets and inlets are closely related to the interior appearance of the building and should be coordinated with inlets and lighting troffers.

TABLE 18.2 Volume Flow per ft2 Floor Area

Type of outlet Grille Slot diffuser Perforated panel Ceiling diffuser Ventilating ceiling

Supply air density, cfm / ft2 of f oor space

Approximate maximum air changes per hour for 10-ft (3-m) ceiling

0.6 – 1.2 0.8 – 2.0 0.9 – 3.0 0.9 – 5.0 1.0 – 10.0

7 12 18 30 60

Source: ASHRAE Handbook 1992, Reprinted with permission.

HVAC Systems and Equipment

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Cost. In many commercial buildings, cost is often an important f actor in determining the type of supply outlet.

Volume Flow Rate per Outlet or per Unit Length For a speci f c supply v olume f ow rate in a conditioned space, the v olume f ow rate per supply outlet V˙ s,out, in cfm (L / s), determines the number of outlets in the conditioned space. As def ned in Eq. (18.10), both V˙ s,out and supply outlet v elocity vs,o or vcore, in fpm (m / s), affect the throw of the supply airstream. Ho wever, V˙ s,out has a signi f cantly greater in f uence than vs,o, especially for slot diffusers. The volume f ow rate per supply outlet depends mainly on the thro w required to provide a satisfactory space air dif fusion system design. Load density , space air dif fusion system characteristics, sound control requirements, and cost considerations are also f actors that determine V˙ s,out. In a VAV system in a high-rise of f ce building, avoiding the dischar ge of cold air into the space directly at a reduced volume f ow rate is one of the primary considerations. For a space air diffusion system using slot diffusers, the supply volume f ow per ft (m) length of slot diffuser (the slot intensity), in cfm / ft (L / s m), is often an important inde x, especially in the return slots. It is usually 15 to 40 cfm / ft (23 to 62 L / s m). All return slots mounted on the same

TABLE 18.3 Relationship between ADPI and T50 / L and T100 / L Load density, Btu / h ft2

T50 / L for max. ADPI

Maximum ADPI

For ADPI greater than

Range of T50 / L

High sidewall grilles

80 60 40 20

1.8 1.8 1.6 1.5

68 72 78 85

— 70 70 80

— 1.5 – 2.2 1.2 – 2.3 1.0 – 1.9

Circular ceiling diffusers

80 60 40 20

0.8 0.8 0.8 0.8

76 83 88 93

70 80 80 90

0.7 – 1.3 0.7 – 1.2 0.5 – 1.5 0.7 – 1.3

Sill grille straight vanes

80 60 40 20

1.7 1.7 1.3 0.9

61 72 86 95

60 70 80 90

1.5 – 1.7 1.4 – 1.7 1.2 – 1.8 0.8 – 1.3

Sill grille spread vanes

80 60 40 20

0.7 0.7 0.7 0.7

94 94 94 94

90 80 — —

0.8 – 1.5 0.6 – 1.7 — —

Slot diffusers (for T100 / L)

80 60 40 20

0.3 0.3 0.3 0.3

85 88 91 92

80 80 80 80

0.3 – 0.7 0.3 – 0.8 0.3 – 1.1 0.3 – 1.5

Light troffer diffusers

60 40 20

2.5 1.0 1.0

86 92 95

80 90 90

3.8 3.0 4.5

35 – 160

2.0

96

90 80

1.4 – 2.7 1.0 – 3.4

Terminal device

Perforated and louvered ceiling diffusers

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return ceiling plenum must ha ve the same total pressure loss, or using same type of return slot, the same slot intensity. In a closed off ce of an area of 150 ft2 (14 m2) or less with only one side of external wall, usually one ceiling diffuser is suff cient.

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Choose an Optimum Throw – Characteristic Length Ratio For most types of supply outlets, the selection of an optimum thro w – characteristic length ratio T50 /L (T0.25 /L) or T100 /L (T0.5 /L) determines the layout and grid of the supply outlets. An optimum T / L ratio should meet the following requirements: ●

●

Selected T50 /L or T100 /L values should have an ADPI greater than 70 and 80 percent, as listed in Table 18.3. The spread at the end of the supply air jet co vers or almost co vers the width of the conditioned space before the air jet enters the occupied zone as shown in Fig. 18.15.

Miller (1976) and others conducted man y experiments and determined the relationship between the ADPI and T50 /L or T100 /L. They found that the cooling load density is also a f actor. Table 18.3 lists the relationship between ADPI and T50 /L or T100 /L at various load densities. The characteristic length of a sill grille is de f ned as the length of the room or space in the direction of jet f ow. In Table 18.3, T50 and T100 (T0.25 and T0.5) are throws of isothermal jets gi ven in most of the manuf acturers’ catalogs for selection convenience. Generally, high sidewall outlets have a longer throw and therefore a higher T50 /L (T0.25 /L) range than do ceiling dif fusers. Slot dif fusers have a wider T100 /L (T0.5 /L) range in which ADPI exceeds 80 percent than do high side wall outlets. This characteristic makes them suitable for use at reduced volume f ows. Square ceiling dif fusers have a T50 /L (T0.25 /L) range similar to that of the circular ones. Higher cooling load density usually results in a lower ADPI. For VAV systems, Tv / L must be selected within the satisf actory range, that is, ADPI  80 percent, for both maximum and minimum airf ow.

Determine the Design Characteristics of Slot Diffusers in Perimeter Zone In the perimeter zone, overhead slot diffusers and ceiling diffusers are widely used. The characteristic length of an o verhead slot diffuser in the perimeter zone has not been clearly de f ned. Based on tests performed by Straub and Cooper (1991), Lorch and Straub (1983), and Int-Hout (1998), the design characteristics of overhead slot diffusers in the perimeter zone are as follows: ●

●

●

●

Two-slot, two-way o verhead slot dif fusers should be used. The suitable distance between the overhead slot diffuser and the external wall is about 1 ft (0.3 m). One of the slots near the window is set to blow down, and the other slot blows horizontally into the room. Such an arrangement offsets the cold draft along the window in winter heating. During cooling mode operation, ADPI exceeding 80 percent occurs within a wider range of slot intensities, load densities, and spacing between the dif fuser and external wall than during heating mode. Heating mode operation requires careful consideration. According to ASHRAE Standard 55-1992, the criteria for an acceptable space air dif fusion in heating mode operation are based on the requirement that the v ertical temperature gradient in the occupied zone measured between 4-in. (0.1-m) and 67-in. (1.7-m) le vels not e xceed 5 °F (3°C). According to ASHRAE Handbook 1997, Fundamentals, the supply temperature dif ferential for warm air during heating mode Ts  Tr should not be greater than 15 °F (8.3°C), to prevent excessive buoyancy.

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A terminal velocity of 150 fpm (0.75 m / s) is recommended at the 5-ft (1.5-m) level near the window, to offset cold drafts in winter. An e venly distrib uted air v olume f ow rate between the v ertically and horizontally dischar ged slots is recommended.

Select the Specific Supply Outlet from Manufacturer’s Catalog After an optimum T50 /L (T0.25 /L) or T100 /L (T0.5 /L) and characteristic length L are determined from the preliminary layout, the supply outlet can be selected from the manuf acturer’s catalog from known T50 (T0.25) or T100 (T0.5) and supply v olume f ow rates, after the follo wing parameters ha ve been checked: Sound Level. The combined sound level of terminal and outlet should be at least 3 dB lo wer than the recommended NC criteria in the conditioned space. F or optimum noise control, the recommended air velocities at the supply outlet are as follows: Residences, apartments, churches, hotel guest rooms, theaters, private off ces General off ces

500 to 750 fpm (2.5 to 3.75 m / s) 500 to 1250 fpm (2.5 to 6.25 m / s)

The outlet v elocity for the ceiling dif fuser can be calculated by di viding the v olume f ow b y a rea factor Ak, given in the manufacturer’s catalog. Drop of Cold Air Jet. Drop of a cold air jet should be check ed if the cold jet enters the occupied zone directly. Figure 18.21 a shows the drop and other data of typical side wall outlets mounted within 1 ft (0.3 m) of the ceiling with an adjustable v ane def ection of vertical 5° up and horizontal 0°. Figure 18.21 b shows the drop for side wall outlets without surf ace effects and a v ane def ection of vertical 15° up and horizontal 0 °. Both are cold air jets based on a supply temperature dif ferential of 20°F (11.1°C). Total Pressure Loss of the Supply Outlet. The total pressure loss of supply air when it f ows through a slot dif fuser with a slot width of 0.75 in. (19 mm) is usually between 0.05 and 0.20 in. WC (12 and 50 P a). For a ceiling diffuser, it is between 0.02 and 0.2 in. WC (5 and 50 P a). A total pressure loss higher than 0.20 in. WC (50 Pa) is unsatisfactory.

Determination of Final Layout of Supply Outlets and Return Inlets The determination of the volume f ow rate per outlet, the number of outlets, T50 /L (T0.25 /L), and the selection of speci f c outlets from the manuf acturer’s catalog is often an iteration procedure. After that, f nal layout of supply outlets and return inlets can be determined. In the perimeter zone, the high side wall outlet in recently constructed hotel guest rooms are often located on the partition w all adjacent to the windo w. In the interior zone, ceiling and slot diffusers are usually located directly above the center of the conditioned space. As mentioned in Sec. 18.4, if a ceiling plenum is used as return plenum, the return inlets should be located outside the supply air jet, above the return airstream, or near a concentrated heat source for a better effectiveness factor . The recommended face velocities for return inlets are as follows: SH__ ST__ LG__ DF

Above the occupied zone Within the occupied zone Door louvers

800 to 1000 fpm (4 to 5 m / s) 400 to 600 fpm (2 to 3 m / s) 300 to 500 fpm (1.5 to 2.5 m / s)

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FIGURE 18.21 Performance of cold air jets based on a supply temperature dif ferential of 20 °F. ( a) Performance of cold air jets from typical grilles and re gisters mounted within 1 ft belo w ceiling; de f ection: vertical 5° up, horizontal, 0°. (b) Performance of free cold air jets from typical grilles and re gisters; def ection: vertical 15° up, horizontal, 0°. (Source: Titus Products. Reprinted with permission.)

Example 18.1. A large, open off ce has a perimeter length of e xternal wall of 30 ft (9.1 m) and a depth of 35 ft (10.7 m) of which 15 ft (4.5 m) is the perimeter zone, as shown in Fig. 18.22. The supply air volume f ow rates and load densities for both perimeter and interior zones are as listed belo w: Design supply volume f ow rate, cfm (L / s) Perimeter zone Interior zone

960 (453) 500 (236)

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FIGURE 18.22

Layout of slot diffusers for a large open off ce in Example 18.1. All dimensions in feet.

Supply temperature differential Sound level Cooling load density in perimeter zone Linear density of heating load

15°F (8.3°C) NC 40 50 Btu / h ft2 (158 W/ m2) 300 Btu / h ft (288 W / m)

Design a space air diffusion system for this off ce for year-round operation using a VAV system. Its minimum supply v olume f ow rate is 50 percent of the design v olume f ow at part-load operation. The ceiling plenum is used as a return plenum. Solution

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1. For a lar ge, open off ce using a VAV system, a space air dif fusion system of mixing air f ow pattern using ceiling slot dif fusers is a suitable choice because of its good surf ace effect and wider T100 /L (T0.5 /L) range. 2. For the perimeter zone, the overhead two-way two-slot diffuser should be located on the suspended ceiling parallel to the e xternal w all and windo w glass. One of the slots should dischar ge downward toward the window glass. The space between the slot dif fusers and the windo w glass is 1 ft (0.3 m). Another slot discharges horizontally inward from the window glass. 3. In Table 18.3 the range of T100 /L for slot diffusers having ADPI  80 percent at a load density of 50 Btu / h ft2 is between 0.3 and 0.95. Considering the reduction of supply v olume f ow rate at 50 percent load, select a T100 /L ratio around 0.7. Because the characteristic length L in the perimeter zone is L  15  1  14 ft, the required throw for the slot diffuser in the perimeter zone

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is T100 /L  0.7, or T100  0.7L  0.7  14  9.8 ft From Table 18.1, two-way two-slot diffusers with a length of 4 ft (1.2 m) and a slot width of 0.75 in. (19 mm) are selected. At a slot intensity of 30 cfm / ft (46 L / s m), its performance is as follows: Throw Horizontal projection, ft (m) Vertical projection T50 (T0.25), ft(m) Total pressure loss Horizontal, in. WC (Pa) Vertical, in. WC (Pa) Sound level, NC curve

8-10-15 (2.4-3-4.6) 14 (4.3) 0.091 (23) 0.067 (17) 29

4. If the supply v olume f ow rate in the perimeter zone is e venly split between horizontal and downward airstreams, the number of slot dif fusers is 0.5(960) / (4  30)  4, so four slot dif fusers should be used. F or the do wnward discharge slot, the vertical throw T150  0.5 T50  0.5  14  7 ft, which is greater than the required throw T150 (ceiling height 9  distance from f oor level 5)  4 ft. The terminal v elocity is higher than 150 fpm (0.75 m / s) near the windo w glass at a le vel of 5 ft from the f oor level. For the horizontally discharged slot, T100 is 10 ft. Because the slot is installed 1 ft a way from the window, T100 /L  (10 1) / 15  0.73. From Table 18.3, at a load density of 50 Btu / h ft2, the range of T100 /L ratio for ADPI  80 percent is between 0.3 and 0.95; T100 /L  0.73 is a suitable setup. 5. Heating load in the perimeter zone is Q rh  30  300  9000 Btu / h If the supply volume f ow in heating mode operation is still 960 cfm, and if the density of supply air is 0.072 lb / ft3, then the supply differential can be calculated as Ts  Tr  

Q rh 60V˙s s Cpa 9000  8.9F (5C) 60  960  0.072  0.243

This value is far smaller than 15°F (8.3°C). As the terminal velocity is approximately equal to 150 fpm (0.75 m / s) near the window glass at a level 5 ft (1.5 m) from the f oor level, and a far lower Ts  Ts, cold draft does not occur at the inner glass surface at a heating load linear density of 300 Btu / h ft (288 W / m2). 6. Four 4-ft two-slot, of 0.75-in. slot width, return slots of the same con f guration as the supply slot diffusers with a slot intensity of 900 / (2  4  4)  28 cfm / ft (43 L / s m) are installed in the perimeter zone. Their locations are arranged on tw o sides of the room in a direction perpendicular to the supply outlets shown in Fig. 18.22 for a better effectiveness factor . Because the total pressure TP loss of return slots is nearly equal to the TP of the vertical projection of the same type of supply slot dif fusers, from Table 18.1, for a slot intensity of 28 cfm / ft, TP for selected return slots is 0.06 in. WC (15 Pa). 7. The depth of the interior zone is 20 ft. It is better to select tw o 4-ft two-way two-slot supply slot diffusers with 0.75-in. slot width and located at the midpoint of the interior zone. Each one-way single-slot dif fuser needs a thro w of 20 / 2  10 ft. From Table 18.1, for a slot intensity of 500 / (2  4  2)  31 cfm / ft, horizontal throw T100 is about 10 ft and the ratio T100 /L  10 / 10  1.0. Such a setup is benef cial during part-load operation.

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The return slots in the interior zone must ha ve the same total pressure loss TP  0.06 in. WC, i.e., the same slot intensity , as in the perimeter zone. This is because both perimeter and interior zone inlets are installed on the same ceiling and use the ceiling plenum as the return plenum. For the interior zone, return slots with a slot intensity of 28 cfm / ft  500 / 28  18 ft (5.4 m) must be installed. Select two 5-ft two-way two-slot return slots with a slot width of 0.75 in. The diffusers will be located on the ceiling at a position midw ay from the depth boundaries of the interior zone, as shown in Fig. 18.22. 8. In the perimeter zone, during cooling mode operation of 50 percent part-load operation, the throw – characteristic length ratio may reduce to T100 4.5   0.32 L 14 In the interior zone, T100 4.5   0.45 L 0.5  20 Both are still within 0.3 to 1.4 at a load density 25 Btu / h ft2. 9. During heating mode part-load operation, the supply air temperature dif ferential, the buoyancy effect, and the vertical temperature gradients will be smaller than those at design load.

18.7 DISPLACEMENT FLOW AND UNIDIRECTIONAL FLOW Displacement Flow Displacement f ow is a f ow pattern in an air conditioned space in which cold supply air , at a velocity nearly equal to the required v elocity in the space, enters the occupied zone or w orking area and displaces the original space air with a pistonlike airf ow without mixing the supply air and the original space air. Compared with mixing f ow, displacement f ow provides a better indoor air quality in the occupied zone. If cold air is supplied at a v elocity nearly equal to the mean air v elocity in the occupied zone with a small supply temperature differential for comfort air conditioning systems, there will be low turbulence intensities and fe wer draft problems. On the other hand, displacement f ow usually requires a greater supply volume f ow rate and a higher construction cost. There are three types of displacement f ow: unidirectional f ow, downward uniform f ow, and stratif ed displacement f ow. Unidirectional Flow

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In a unidirectional f ow, the conditioned supply airstream f ows in the same direction as a uniform airf ow and showers the entire w orking area or occupied zone. Supply airstreams occup y the entire working area or occupied zone of the conditioned space. Types of unidirectional f ow currently used in clean rooms include do wnward unidirectional f ow and horizontal unidirectional f ow, as shown in Fig. 18.23. Unidirectional f ow refers to airf ow patterns in which the stream lines of air f ow are uniform and mo ve in the same direction. Because of the unidirectional f ow, contaminants generated in the space cannot mo ve laterally against the downward air f ow, and dust particles will not be carried to higher le vels by recirculating f ow o r large eddies. Such an airf ow pattern in clean rooms was formerly known as laminar f ow. Laminar f ow is often confused with the types of f uid f ow distinguished by Re ynolds number in f uid mechanics.

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FIGURE 18.23 Unidirectional f ow for clean rooms: (a) downward unidirectional f ow for clean room; ( b) horizontal unidirectional f ow for clean room; (c) unidirectional f ow for clean workstation.

Most forced air f ows driven by f ans in air conditioned space are turb ulent f ow except the laminar f ow in the boundary layers adjacent to the surf ace of the building envelope. The Reynolds number of forced air f ow, even at v ery lo w air v elocity, is usually greater than 1  104. As of 1991, ASHRAE Handbook 1991, HVAC Applications, refers to this f ow pattern as unidirectional f ow.

Unidirectional Flow for Clean Rooms Figure 18.23 a shows a do wnward unidirectional f ow for clean rooms. After f owing through the high-eff ciency particulate air (HEP A) f lters or ultralo w penetration (ULP A) f lters, ultraclean air dischar ges uniformly do wnward in parallel stream lines and enters the w orking area. It then

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f ows through the raised f oor grating and is returned to the recirculating air unit at the top of the clean room. Figure 18.23b shows a horizontal unidirectional f ow for a clean room. Instead of f owing downward, the clean airstream discharges horizontally from the HEPA f lters on one side of the room and f ows through the working area. The contamination level near the return inlets of a horizontal unidirectional f ow may be higher than those of a downward unidirectional f ow. Figure 18.23c shows a clean w orkstation that pro vides a small noncontaminated w orking area for a single w orker by means of horizontal or do wnward unidirectional f ow and HEP A f lters. Clean workstations can achieve a high degree of contamination control over a limited area for many practical applications. To provide parallel stream lines, air velocity of 60 to 90 fpm (0.3 to 0.45 m / s) is required. Unidirectional f ow pro vides a direct and predictable path of submicrometer -size dust particles and minimizes the opportunities for these particles to contaminate w orking parts. It also captures internally generated dust particles and carries them a way. Most dust particles in unidirectional f ow reestablish their parallel stream lines after the downstream eddies of an obstruction. The supply temperture differential Tr-Ts for unidirectional f ow for clean rooms depends mainly on the required space velocity and the cooling loads to be remo ved within the working area. A case study of a class 10 clean room is discussed in Chap. 30.

Ventilating Ceiling A ventilating ceiling is sometimes called a perforated ceiling. It creates a do wnward uniform f ow similar to the downward unidirectional f ow for clean rooms. In most cases, ventilating ceilings discharge conditioned air through the entire ceiling to form a do wnward uniform f ow, except in the area occupied by light trof fers. The primary dif ferences between unidirectional f ow for clean rooms and downward uniform f ow from ventilating ceilings are as follows: ●

●

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Unidirectional f ow requires a 60 to 90 fpm (0.3 to 0.45 m / s) air velocity and ultraclean air in the working area, and ventilating ceilings usually have a mean air velocity of less than 15 fpm (0.075 m / s) of conditioned air. There is no mixing of supply and space air in unidirectional f ow; whereas just belo w the perforated ceiling, supply air is mix ed with the ambient air at a v ertical distance of less than 1 ft (0.3 m) in downward uniform f ow from the ventilating ceiling.

FIGURE 18.24

Ventilating ceiling.

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Figure 18.24 shows a downward uniform airf ow. In Fig. 18.24, conditioned air is f rst supplied to the ceiling plenum through supply outlets inside the plenum. It is then squeezed through the holes or slots of the v entilating ceiling and discharged to the conditioned space in a do wnward uniform f ow. When parallel airstreams combine and f ow through the w orking area, they arrive at the f oor level, turn to the side return grilles at lo w levels, and then return to the air -handling unit or packaged unit in the fan room. Downward uniform air f ow patterns from v entilating ceilings ha ve the follo wing characteristics: ●

●

●

__RH TX

There is no mixing or induced recirculating f ow in the w orking area or occupied zone, so dust particles contained in the space air at a lo wer le vel are not carried to a high le vel to enter the working area again. A very low air v elocity can be maintained in the w orking area or the occupied zone e ven when the volume f ow per ft 2 V˙ / Af , in cfm / ft2 (L / s m2) is very high. Here V˙ represents the supply volume f ow rate, in cfm (m3/s), and Af represents the f oor area, in ft2 (m2). When /V˙ Af  3 cfm / ft2 (15 L / s m2) the downward uniform f ow becomes prominent.

Ceiling Plenum and Supply Air Velocity To create a more uniform supply air v elocity vo, in fpm (m / s), at the perforated holes or slots, the maximum velocity of the airstream crossing the v entilating ceiling inside the ceiling plenum vp, in fpm (m / s), should be low. Usually, vp  0.5vo. Moreover, the clearance between the beam and the ventilating ceiling Hip, in ft (m), should be always greater than 8 in. (200 mm ) to pre vent a greater pressure loss against the air f ow crossing the v entilating ceiling inside the plenum. The greater the supply v olume f ow rate, the higher Hip. If the ceiling plenum is of suf f cient height Hp and few small obstructions other than the beams are present, distributing ductwork inside the ceiling plenum is not necessary to provide a uniform air supply. Supply air dischar ged from the outlets into the ceiling plenum al ways mo ves in an upw ard direction so that the v elocity pressure of the dischar ged airstream does not af fect the supply air velocity at the perforated openings. The construction of the ceiling plenum must meet the requirements of the National Fire Protection Association (NFPA) and local f re protection codes. Ventilating ceilings are al ways made of noncomb ustible materials such as metal strips or mineral acoustic tiles. As a supply plenum, the ceiling plenum must be insulated against heat gain and loss, and condensation on its inner surf ace if it is adjacent to an area that is not air conditioned. Supply air v elocity at the perforated openings vo must be optimal. When vo exceeds 1000 fpm (5 m / s), objectionable noise is generated. A higher vo also means higher pressure at the ceiling plenum and, therefore, greater air leakage. On the other hand, a lower vo may cause unevenness in the supply airf ow rate because of a higher vp / vo ratio. Supply air velocity vo is usually between 200 and 700 fpm (1 and 3.5 m / s). When vo has been determined, the perforated area Ao, in ft2, (m2) can be calculated as Ao 

V˙s vo

(18.16)

where V˙ s  supply volume f ow rate, cfm (m 3 /s). Generally, the diameter of the circular holes or the slot width is less than 0.25 in. (6.5 mm). The number of holes or slots can then be determined accordingly. Based on test results, the local loss coef f cient of the perforated holes and slots Co including discharge air velocity pressure at the perforated openings is about 2.75. Ventilated ceilings are often used in industrial applications where v ery low air mo vements or precise control of space temperature up to 68  0.1°F (20  0.056°C) is needed. They are also used in indoor sport stadiums for badminton and table tennis tournaments, where an air v elocity of less than 40 fpm (0.2 m / s) should be maintained in the arena.

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Two-Zone Stratified Model Stratif ed displacement f ow supplies conditioned cold air (in Scandina vian countries usually 100 percent outdoor air at lower velocity) at a low-level supply outlet, as shown in Fig. 18.25. The cold supply air, with a volume f ow rate of V˙ s, in cfm (L / s), f ows in a thin layer along the f oor. Above the heat and contaminant sources, heated air containing contaminants rises upw ard because of its buoyancy effect. Supply air is then entrained into the upw ard convective f ow with a v olume f ow rate of V˙ conv, in cfm (L / s). When upward convective f ow arrives at a height where its v olume f ow rate is equal to the supply v olume f ow rate, that is, V˙ conv  V˙ s, this height is recognized as the stationary level zstat, in ft (m). Above zstat, ambient air is induced into the upw ard convective f ow until it reaches the ceiling. Upw ard convective f ow with induced air spreads laterally along the ceiling. An amount nearly equal to the supply v olume f ow rate exhausts or returns to the f an room through the exhaust or return inlet near the ceiling. The remaining portion containing the contaminant descends to the stationary level to be induced into the upward convective f ow and recirculated. The stratif ed level divides the room v ertically into two zones: an upper zone and a lo wer zone. In the lower zone, only the supply air is induced into the upward convective f ow, as its volume f ow rate is smaller than the supply volume f ow rate V˙ conv V˙ s. In the upper zone, the portion of air and contaminants that is greater than the supply f ow rate V˙ conv  V˙ s recirculates. For cold air supply , the mean temperature of the upper zone is usually 1 to 2 °F (0.56 to 1.1 °C) higher than that of the lower zone.

Operating Characteristics Stratif ed displacement f ow (or displacement ventilation) was introduced to Scandinavian countries in the early 1970s as a means of impro ving general ventilation in industrial applications. F or comfort systems in public and off ce buildings, it was adopted in the early 1980s.

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FIGURE 18.25

Stratif ed displacement f ow in a typical room.

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According to Sv ensson (1989), Mathisen (1989), Sandberg and Blomqvist (1989), and Sepp änen et al. (1989), stratif ed displacement f ow has the following characteristics: ●

●

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Cold air supply of usually 100 percent outdoor air is used to remo ve cooling loads in conditioned space. Heating is usually pro vided by radiant panels or a baseboard heater under the windo ws or on the walls. Air must be supplied at low velocity, usually less than 60 fpm (0.3 m / s) and at a height often less than 1.8 ft (0.54 m) from the f oor level. Cold air is generally supplied at a temperature 5 to 9 °F (2.8 to 5 °C) lower than that of the air in the occupied zone or working area. The height of the lo wer zone, or zstat, generally should be higher than the breath line of a seated occupant (4.5 ft or 1.4 m). Stationary le vel depends mainly on the supply v olume f ow rate V˙ s. In the lower zone, all air is supply air , theoretically, except the downward cold drafts. Supply air is supplied to the occupied zone directly without mixing with the ambient air. Because of the small supply temperature dif ferential, the maximum cooling load that can be removed from the room is considerably less than mixing f ow. Some scientists recommend a maximum cooling load of 13 Btu / h ft2 (41 W / m2) at a ceiling height of 9 ft (2.7 m). F or greater cooling loads, additional cooling panels mounted on the ceiling or part of space air mixing should be used. Return or exhaust inlets are located near the ceiling level. According to the required thermal comfort of the occupant, the vertical temperature difference between the 0.3- and 5.5-ft (0.09- and 1.7-m) le vel as speci f ed in Sec. 4.8 should not e xceed 5°F (2.8°C).

Busweiler (1993) reported the use of a combination of radiant cooling, displacement ventilation, and desiccant cooling during the retrof t of the HVAC&R system of a conference room in a hotel in Bremen, Germany. Because of the limited space in the ceiling plenum, air ducts for a con ventional air conditioning system could not be installed. The cooled ceiling w as used to remo ve the sensible heat and controlled the room temperature. Displacement v entilation supplied 100 percent outdoor air from wall outlets near to the f oor level at a velocity of 40 fpm (0.2 m / s). Desiccant-based cooling was sometimes used to cool and dehumidify the outdoor air and maintain a room temperature of 68°F (20°C) and a relati ve humidity of about 60 percent. The f rst-year operation in the Bremen hotel reveals that these systems are a good and practical combination. Comparison of Stratified Displacement Flow and Mixing Flow Compared with mixing f ow space diffusion systems, stratif ed displacement f ow has the following advantages: ●

●

●

Contamination level in the occupied zone is lower, and therefore indoor air quality is better. Space diffusion effectiveness factors T and C are higher. Turbulence intensities Itur are lower, and therefore fewer draft problems occur even at higher mean air velocities. The disadvantages of the stratif ed displacement f ow are as follows:

●

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Initial cost is signi f cantly higher if cooling load density is greater than the recommended maximum 13 Btu / h ft2 (41 W/ m2) and if cooling panels are added. Energy cost is comparatively higher. Stratif ed displacement f ow is for space cooling only.

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Benefits of Projecting Flow In a projecting f ow air pattern, the cold or warm air jet is deliberately projected into part of the occupied zone or working area, which is often called the target zone. The result is control of the environment in a small or localized area, or microenvironmental control. This projecting airf ow pattern is used in spot cooling / heating, task air conditioning, or controlling personal environments. Spot cooling / heating using projecting f ow has man y adv antages o ver con ventional space air diffusion: ●

●

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●

It provides better control of temperature, air cleanliness, and air mo vements in a localized en vironment. Spot cooling impro ves the occupants ’ thermal conditions signi f cantly, and their heat strain decreases. There is greater direct outdoor air supply. There is direct and eff cient handling of local loads. Occupants can have greater control of their o wn microenvironment or personalized en vironment.

The main disadv antages of spot cooling / heating include draft discomfort or discomfort due to the pressure of the air jet, limited area of environmental control, and a more complicated space air diffusion system design. Air jets in projecting f ow are usually free jets with high entrainment ratios because of their large contact area with ambient air . Long-throat round nozzles are often used as supply outlets for spot cooling in large spaces. Two types of projecting f ow are currently used today: industrial spot cooling systems and desktop task air conditioning systems.

Industrial Spot Cooling Systems In an industrial spot cooling system, as shown in Fig. 18.26 a, the temperature differential between the target zone and the room air (ambient air) is often 5 °F (2.8°C) or greater. Melikov et al. (1994) summarized the following based on their and other research workers’ experimental results: Distance Between Target Zone and Supply Outlet. The core zone of a free air jet (from surf ace of outlet to x / Do  4) and transition zone (x / Do between 4 and 12) had a stronger cooling capacity than full de veloped main zone. This w as due to the intensi ve mixing of the cold jet air with the room air in the main zone. The difference between the target temperature Tj and the outlet temperature To increased as the distance between the tar get zone and the outlet x and the ambient air temperature Ta were increased. For a tar get temperature of 77 °F (25°C) and an ambient (room) temperature of 86 °F (30°C), if x / Do  2, the required outlet temperature To was about 71.6 °F (22°C). If x / Do  8, To was about 64.4°F (18°C). The size of the tar get zone can be estimated since the angle of di vergence of the air jet   22° and the distance from the outlet x are known. High turb ulence intensity caused intensive mixing of the cold air jet and, therefore, decreased the cooling capacity of the air jet.

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Vertical Versus Horizontal J et. Many reseachers preferred a v ertical air jet because a horizontal air jet required a large diameter of the jet’s outlet and the worker’s body was not exposed symmetrically to the cold air sho wer. In Melik ov’s e xperiments, when the room temperature w as 82.4 °F (28°C), the target velocity vj preferred by the subjects participating in the e xperiments was almost identical with v ertical and horizontal cooling jets. Ho wever, when the room temperatures were 91.4°F (33°C) and 100.4°F (38°C), experimental results conf rmed that a vertical jet cooled the subjects more than a horizontal jet.

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Air damper Vertical jet 1m

0.5 m

(a)

Supply nozzles (outlets) Control panel

Mixing box (b)

Access floor panel

FIGURE 18.26 Projecting f ow: (a) an industrial spot cooling system; (b) a desktop task conditioning system.

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Target Velocities. The preferred tar get velocities vj of participating subjects at v arious conditions were as follows: Study

Ta, °F(°C)

Tj, °F(°C)

vj, fpm (m / s)

Thermal sensation

Jet type

Melikov et al. (1994)

91.4 (33)

82.4 (28)

280 (1.4)

Slightly warm

Olesen and Nielsen (1983) Melikov et al. (1994)

86 (30) 86 (30) 86 (30)

75.2 (24) 75.2 (24) 75.2 (24)

320 (1.6) 400 (2.0) 440 (2.2)

Slightly warm Slightly warm

Fanger et al. (1974)

82.4 (28)

82.4 (28)

160 (0.8)

Neutral

Melikov et al. (1994)

82.4 (28) 82.4 (28)

82.4 (28) 82.4 (28)

300 (1.5) 300 (1.5)

Vertical Horizontal Vertical Horizontal Vertical Horizontal Horizotal Whole body Vertical Horizontal

The preferred tar get velocities show substantial dif ferences from dif ferent studies. The amount of the body exposed to the air jet is also an important factor that affects vj. Thermal Sensation. Large dif ferences occurred in the local thermal sensation because of the velocity and temperature distrib ution of the air f ow around the human body . Under these circumstances, an occupant’s acceptance of the thermal en vironment showed a compromise between the thermal sensation of the whole body (general thermal sensation) and the thermal sensation for individual parts of the body (local thermal sensation). When air velocities were above 200 fpm (1 m / s), the airf ow exerted a physical pressure on the skin that produced a discomfort. This pressure is the sum of the static and v elocity pressure, the total pressure which must tak e into account the direction of air f ow. At high local air temperatures when T 82.4°F (28 °C), a subject participating in the tests preferred the slight w armth instead neutral sensation because of the anno ying effect of the pressure at the high-v elocity air jet. At low jet temperatures, local discomfort created by draft was mostly at the head and neck for a v ertical jet and at the neck, face, and left shoulder for a left horizontal jet. Recommendations in Spot Cooling Design Melikov et al. (1994) and Brown (1988) recommend the following for spot cooling design: ●

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Allow the occupant to ha ve individual control of the jet v elocity and temperature. Spot cooling systems that permit the occupant to adjust the distance between the tar get zone and the outlet x and the direction of the air jet are preferrable. The target zone should be located in the core or transition zone of a cooling jet; typically , x / Do  4 to 5 from the nozzle outlet. The maximum target velocity (mean value) should not e xceed 440 fpm (2.2 m / s), and the maximum local velocity should not exceed 840 fpm (4.2 m / s). When ambient air temperature is abo ve 95 °F (35 °C), use of a combination of a v ertical jet, a helmet, and a clothing ensemble to protect the neck from draft is required.

Desktop Task Conditioning Systems SH__ ST__ LG__ DF

System Description. Bauman et al. (1993) reported the results of laboratory and f eld measurements investigating the performance of a desktop task conditioning system. A typical desktop task conditioning system, as sho wn in Fig. 18.26 b, consists of a self-po wered mixing box, two small

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fans, two desktop supply outlets, f exible ducts, and a control panel. Inside the mixing box, there is a variable-speed f an to e xtract primary air (cold conditioned air from the AHUs) either from the low-pressure f oor plenum or from the supply duct in the ceiling plenum by means of connecting f exible ducts. Another fan pulls recirculating air from the knee space under the desk. The primary air and recirculating air are mix ed together according to the required fractions by re gulating the opening of their dampers in each of the transporting ducts. The mixture is then dra wn through an electrostatic air f lter. The primary air damper is ne ver allowed to close completely , to ensure that the minimum amount of outdoor v entilation air is supplied all the time. The mixing box is hung on the back or on the corner of the desk and is connected to the tw o supply nozzles (outlets) on the top of the desk by f exible ducts. Two supply nozzles may be rotated 360 ° on the horizontal plane. They are installed with horizontal guide vanes which are adjustable  30° in the vertical plane.

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Control P anel. A desktop DDC unit controller uses adjustable sliders to control the v ariablespeed fan and modulate the volume f ow of the supply air from the nozzles. The supply air temperature is adjusted by varying the opening of the primary and recirculating dampers, and thus the ratio of primary to recirculating air . A 200-W electric radiant heating panel in the knee space is employed during winter heating. The dimming of the occupant ’s task light controls the lighting luminance and a white noise generator which has a continuous frequenc y spectrum with equal energy/Hz over a specif ed frequency range. A motion detector – based occupanc y sensor sends a signal to the DDC unit which shuts the desktop system off when the workstation is unoccupied for a specif ed time. Operating Characteristic. Each desktop unit can pro vide 40 to 250 cfm (20 to 70 L / s) of supply air from the supply nozzles. Ev en when the tw o fans in the desktop unit are shut of f, 40 cfm (20 L / s) of primary air will still be delivered to the conditioned space for the minimum outdoor ventilation air requirement. The maximum outlet velocity measured at the f ace of the 2.3- by 4-in. (58- by 100-mm) supply outlet during laboratory tests varied between 6.5 and 24.5 fps (2 and 7.5 m / s). For a typical desktop system, the primary air temperature supplied from a v ariable-air-volume (VAV) AHU w as 55 °F (12.8°C). After mixing with recirculating air , the supply air temperature from the desktop outlet was 65°F (18.3°C). The supply nozzles are 5 ft (1.52 m) wide by 30 in. (0.76 m) deep at the back corners of the desk. Air is supplied from these tw o nozzles toward a focus point near the center of the front edge of the desk. F or isothermal air jets of horizontal supply with a v olume f ow of about 90 cfm (43 L / s), the average speed at the work location in front of the desk was 3.3 fps (1 m / s). Performance of Desktop Task Conditioning Systems Based on the laboratory in vestigation and f eld-measured results, the performance of desktop task conditioning systems was summarized in Bauman et al. (1993) as follows: ●

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The desktop unit can be controlled to produce a wide range of thermal conditions, allowing occupants the opportunity to f ne-tune the local en vironment to their indi vidual comfort preferences. Under the inf uence of different load density of the adjacent workstation, or under a warm average room air temperature, a desktop unit is capable of being f ne-tuned to maintain a nearly comfortable environment, or to maintain a temperature of 1 to 3 °F (0.56 to 1.7 °C) lower than that of the workstation without the desktop unit, even at relatively low supply air volume f ow rates. At the same supply v olume f ow rate, a lar ger supply nozzle w as sho wn to deli ver lo wer air velocity than the smaller nozzle. A lar ger nozzle reduces the potential to produce draft while maintaining improved ventilation performance at moderate to high supply volume f ow rates. Turning off the desktop unit whene ver the workstation is unoccupied signi f cantly reduces its energy use.

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The f rst large installation of the desktop system was recently completed in a new off ce building for an insurance company in West Bend, Wisconsin. This building has 370 desktop units and w as fully occupied in July 1991. Reseachers had track ed the productivity of more than 100 emplo yees before and after the y moved to the ne w building. The study has concluded that the desktop system does have a positive impact on worker productivity. The desktop systems pro vide e xcellent indi vidual thermal control for occupants and a direct supply of primary air including outdoor v entilation air to occupants. Ho wever, desktop systems often need a f oor plenum for primary air supply and a ceiling plenum for return air. Desktop units and task air conditioning for microen vironmental (personalized en vironmental) control have great potential because of the smaller controlled en vironment and therefore lo wer energy use. Air conditioning integration with furniture may be important in the future.

18.10 UPWARD FLOW UNDERFLOOR AIR DISTRIBUTION Upward Flow from Floor Plenum Upward f ow from a raised f oor has been successfully used for space air dif fusion in computer rooms and other industrial applications with high cooling load density , such as 60 to 300 Btu / h ft2 (189 to 946 W / m2). Figure 18.27 sho ws an upw ard f ow space air dif fusion from a raised f oor plenum — an underf oor air distribution system (an air distribution system consists of ductwork distribution, terminals, and space air diffusion) with the following characteristics:

Ceiling plenum

Return air 80 F

Floor diffuser Concrete structural slab

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FIGURE 18.27

Floor plenum

Fan diffuser

Desktop supply F 65 outlet Mixing box

55 F

Upward f ow underf oor air distribution system.

Conditioned space

Return air

Ceiling

Access floor panel

Cold primary air supply duct

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Conditioned air is supplied to an access f oor plenum. The f oor is usually raised from the structural f oor 12 to 18 in. (0.30 to 0.45 m) depending on whether f an-coil units (FCUs) or w ater source heat pumps (WSHPs) are installed inside the f oor plenum. Cable services are usually provided below the access f oor. Types of supply outlets include f oor diffusers, fan-driven f oor diffusers, desktop units, and supply outlets from FCUs and WSHPs. The f oor plenum can be pressurized at a positi ve pressure of usually less than 0.08 in. WG (20 P ag) if conditioned air is squeezed out from the f oor plenum through f oor diffusers directly. For fan diffusers, desktop units, grilles from FCUs, and WSHPs, conditioned air is forced to the occupied zone directly by indi vidually installed small fans. Under these conditions, the f oor plenum pressure should be maintained at slightly ne gative pressure between  0.02 and  0.05 in. WG ( 5 and  12.5 Pag). In the occupied zone, the air f ow patterns of the supply air dischar ged from the f oor diffusers, supply outlets of the FCUs, and WSHPs are mixing f ow, and the air discharged from the desktop nozzles is projecting f ow. Ho wever, from the f oor le vel up to the ceiling plenum, the upw ard f ow is the dominating airf ow pattern. Cool primary air (including outdoor v entilation air) from the air -handling unit is supplied at neutral pressure to the f oor plenum at a temperature of 45 °F (7.2 °C) for cold air distrib ution and 55°F (12.8°C) for con ventional air distrib ution. Because all the supply outlets of the under f oor air distribution system are located in the occupied zone, the cool primary air must be mix ed with the recirculating air by using f an diffusers, desktop units, FCUs, and WSHPs. The temperature of the mixture at the supply outlet (supply air temperature) is often between 63 and 65 °F (17 and 18.5°C). After the space cooling load is absorbed by the supply air , the bouyant space air rises. The temperature of the return air entering the light trof fers is usually between 80 and 82 °F (26.7 and 27.8°C). Return air then enters the ceiling plenum. A part of it is returned to the f oor plenum through the return air shafts along the columns, and the remaining part is returned to the fan room through a return system. FCUs or WSHPs are often installed inside the access f oor plenum and are used to of fset the building envelope heat losses and heat gains.

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In recent years, there have been rapid developments in off ce automation. Underf oor air distribution is a new air distribution system that intends to keep in pace with these advances. Shute (1995) summarized the e xperience of an inte grated access f oor plenum prototype project which has been in full operation for 4 years as follows: Thermal Storage of Floor Plenum The distrib ution of cold primary air in the f oor plenum is not fully ducted. If the f oor plenum is composed of a 4-in. (100-mm) concrete structural slab and a concretef lled access f oor, the cold primary air within the f oor plenum is in direct contact with a density of 50 lb / ft2 (244 kg / m2) thermally acti ve concrete and the f oor panel which are potentially thermal storage media. During a winter night, the mass may pro vide after -hours lo w-grade night setback space heating, and creates a cooled reserv oir to assist in of fsetting a part of the cooling load the ne xt day. On a winter day , part of the internal heat gain is stored, and the reco vered heat is used to assist the follo wing night ’s setback space heating. During the summer , from early morning (4.00 a.m.), the system is started to subcool the mass, which helps to reduce the electric peak demand. Observations with a laboratory e xperiment have indicated that a 4-in. (100-mm) structural slab plus the total mass of the access f oor panel could participate in a 6 to 7 °F (3.3 to 3.9 °C) temperature variation cycle within a 24-h period. This allows storage of up to 30 percent of a typical daily cooling load.

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When the supply air from the f oor level is returned at the ceiling level, upward airf ow lifts the heat gains in the upward f ow direction and returns some heat unneutralized to the ceiling. Furthermore, the upw ard air f ow captures most of the lighting heat gain in the return air f ow. Research has demonstrated that for a typical of f ce, 15 percent of space heat gain and most of ceiling-le vel lighting heat gain can be considered unneutralized because of the localized thermal b uoyancy. In off ces where equipment is concentrated, convective plumes, upward w arm e xpanding air f ow a bove t he heat sources, can capture an e ven greater percentage of space load by the f oor-to-ceiling airf ow, and results in greater capability to e xhaust heat through ceiling return than the con ventional space air diffusion system.

Maintaining a Consistent Access Plenum Temperature Because of the smaller supply air temperature dif ferential Tr  Ts in underf oor air distrib ution, if the underf oor plenum has hot or cold areas, the occupants’ comfort will not be satisf ed by pulling air from these localized areas of temperature inconsistency. The following rules are based on test results: ●

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Supply air colder than 63 °F (17.2°C) is perceived as too cold by occupants. Blend air at the f oor outlets and at other supply units. Primary air de gradation becomes unacceptable when it tra vels a distance e xceeding 30 ft (9 m) from the point of air introduction within the access f oor plenum. Pro vide suf f cient insulated ductwork to distribute primary air. A variation of more than 3°F (1.7°C) across a master control zone may e xceed the air transfer fan adjustment range to compensate for this variation. Practical experience limits the master zoning (a zone corresponds to an area served by a VAV box or other terminals) to about 3000 ft 2 (279 m2) if primary air is supplied from a central riser with VAV control, or a maximum of 10,000 ft 2 (929 m2) if primary air is supplied from an on-f oor fan room.

Floor Plenum Master Zone Air Temperature Control For a VAV system, as space load v aries, the supply v olume f ow rate supply to each master control zone through an individually controlled VAV box will be v aried accordingly. Because of the thermal storage effect, the f oor plenum air temperature does not respond as quickly as required. It takes approximately 20 min before the plenum temperature falls suff ciently to call for proper action. As the space load reduces, more recirculating air is pulled to mix with the cold primary air . The primary air rate becomes proportionally greater in the f oor plenum. This effect increases the pressure on the f oor plenum, and its pressure becomes positi ve. The pressurized cold primary air o verpowers the normal downward return airf ow. If a temperature sensor is located in the downward airf ow path, this cold air will be sensed by the sensor , with a resulting rapid reponse to cut the primary air supply f ow rate.

Design Considerations ●

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A structural concrete slab of 4-in. (100-mm) thickness that separates the supply raised f oor plenum and the return ceiling plenum is recommended. An attempt to baf f e supply and return sections and inte grate them in the plenum of the same le vel has inherent dif f culties due to air leakages through plenum baff es, short-circuiting of air in the space, and limiting of the supply air temperature range.

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The diameter of the f oor diffuser is usually 8 to 15 in. (200 to 375 mm). The mean air velocity at the outlet is often 200 to 400 fpm (1 to 2 m / s). For a 12-in. (300-mm) f oor diffuser, the maximum supply volume f ow rate V˙ max  150 cfm (71 L / s). The mean air velocity of a radial air jet of a f oor diffuser in which the supply air is dischar ged in a radial direction along the f oor decays more rapidly than a v ertical air jet of the same size and supply volume f ow rate. According to Sodec and Craig (1990), the attenuation of the ductwork between two f oor diffusers in adjacent rooms is usually greater than 29 dB. Therefore, conversation transmission through f oor diffusers is negligible.

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Applications of Upward Flow Underfloor Air Distribution Compared with a conventional ceiling supply mixing f ow air diffusion system, the primary advantages of an upward f ow underf oor air distribution system are as follows: ●

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It keeps pace with the requirements of the raised f oor plenum in off ce buildings of high automation. Conditioned air (including outdoor v entilation air) is supplied to the occupied zone directly , which results in a better indoor air quality. When air is supplied from the f oor dif fusers and dischar ged and returned from the ceiling plenum, usually there is no stagnant air in the occupied zone. Upward airf ow lifts some unneutralized heat and carries most of the lighting heat gains at high level to the ceiling plenum; these result in a higher return air temperature and, therefore, a higher space diffusion effectiveness factor and a low space cooling load. It utilizes the mass heat storage of the access f oor panels, and part of the structural concrete slab subcools the f oor plenum and reduces the summer peak electric demand charge. Disadvantages of upward f ow underf oor air distribution include the following:

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Valuable f oor space may be occupied by the f oor diffusers and the b uffer zones between the air jets and the workstation. Both f oor and ceiling plenums are required, and the initial investment is higher than that for conventional ceiling supply mixing f ow systems.

For off ces and applications other than computer rooms, upward f ow underf oor air distribution for commercial buildings was f rst developed in Germany, and many underf oor air distribution systems have recently been adopted in South Africa, Hong Kong, and Japan. In the 1990s, some modern upw ard f ow under f oor air distrib ution systems were completed in Canada and the United States. One, an underf oor air distrib ution system using 370 desktop units for an of f ce building in West Bend, Wisconsin, was fully occupied in July 1991 (Bauman et al., 1993). Upward f ow under f oor air distrib ution systems are still in the de velopmental stage. Where a f oor plenum is needed, an underf oor air distribution system might be a suitable choice.

18.11 COMPUTATIONAL FLUID DYNAMICS Computational f uid dynamics (CFD) including heat transfer is a kind of computing technique for the quantitative prediction of air motion and heat transfer , and it presumes that the physical beha vior of air f ow and thermal systems can be predicted by approximated Na vier-Stokes and thermal equations.

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HVAC&R engineers traditionally conducted scale-model physical e xperiments to study air motion and temperature distribution by maintaining similarity through a high Re number and an appropriate Ar corresponding to the adopted scale. CFD w as f rst developed in Europe and Japan. In the late 1980s, ASHRAE organized a research project which thoroughly investigated many issues associated with CFD simulation. In the late 1990s, CFD became more popular in the HVAC&R industry for the determination of locations of supply outlets and return inlets, airf ow patterns, air velocity and temperature distributions, and contaminant concentration and remo val in an HVAC&R system. Baker et al. (1997) and Ladeinde and Nea von (1997) summarized the reasons for CFD ’s popularity: ●

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CFD has the ability to establish f rm quantitative data re garding air motion and can predict f uid characteristics and pressure dif ferentials to v ery low levels that are essentially impossible during experiments. The cost of CFD is lower than that of scale-model experiments. Earlier CFD codes were developed on supercomputers. Today computing software is available for simulations to be performed on personal computers. People are slowly getting over the fear of using CFD.

Reynolds-Averaged Navier-Stokes Equations The mathematical expressions describing the basic space-time relationship between mass, velocity, and temperature are expressed in partial differential equations called Navier-Stokes (NS) equations. These equations may be only directly applied to laminar f ow f elds. Baker et al. (1997) pointed out that CFD modeling based on NS equations requires a lar ge number of assumptions and approximations. These approximations are closely related to the CFD computing results. For predicting turb ulent f ow, an approximation called Re ynolds a veraging (Ra) is used to convert this time-unsteady f ow to a mean v elocity presentation. A principal assumption for CFD is the turb ulence closure model to go vern the Re ynolds-averaged Na vier-Stokes partial dif ferential equations. The selected CFD has been a tw o-equation turbulent kinetic ener gy (TKE) model. The TKE model assumes a fully turb ulent f ow existing everywhere for room air f ow prediction. Actually, at normal air change rates per hour , fully turbulent f ow occurs only in supply ducts, air jets, downstream of the edge of the obstacles. Else where, the f ow is more lik ely weakly turb ulent and actually time-unsteady. A quantitative measurement of the de gree of turbulence of an air f ow is the turbulence Reynolds number (Re t) which is de f ned as the ratio of the turb ulence eddy viscosity to the kinematic viscosity: Ret  vt / v. For laminar f ow, Ret  0, and for fully turb ulent f ow 5 0 Ret 103.

Numerical Methods

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Because these Re ynolds-averaged partial dif ferential equations using the turb ulence Re ynold number Re t are highly nonlinear , they are not solv able by e xplicit, analytical methods. In these partial dif ferential equations, the v elocities u, v, w; and pressure p; temperature T; and some scalar  are dependent v ariables to be calculated. Space displacements x, y, z and time t are independent variables. Initial and boundary conditions must be speci f ed. Numerical solution (approximate methods) such as via the f nite element and f nite volume methods is more popular in CFD simulation. Ladeinde and Nea von (1997) illustrated the f nite element procedure during the calculation of airf ow in a duct section. Figure 18.28 a shows this duct section that contains an inlet and an outlet. The interior of the duct that constitutes the re gion of air f ow (domain of the CFD model)

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Solid wall F

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Flow inlet A

Flow outlet

B C

D

Solid wall (a) cell

node

F

A

E

B C

D (b)

FIGURE 18.28 Calculation of air f ow using CFD in a duct section: (a) a duct section with sudden e xpansion; (b) f nite element model for air f ow in a duct section. ( Source: ASHRAE J ournal, January 1997. Authors Ladeinde and Nearon. Reprinted by permission.)

is discretized into nodes and cells as sho wn in Fig. 18.28 b. The integration of the equations using the f nite element method gi ves the dependent v ariables at each node of the CFD model. The graphic of the v ectors representing the v elocities using the f nite element method is sho wn in Fig. 18.29.

FIGURE 18.29 Velocity v ectors of the air f ow in a duct section. ( Source: ASHRAE J ournal, January 1997. Authors Ladeinde and Nearon. Reprinted by permission.)

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CFD will not replace the physical e xperiments for airf ow and heat transfer in the HVAC&R industry completely; ho wever, it will signi f cantly reduce the physical e xperiments to study the air f ow patterns and the space dif fusion effectiveness factor of v arious space air dif fusion systems, the removal of contaminants in clean rooms, and the ef fectiveness of the introduction of outdoor air to the occupied zone. Baker et al. (1997) asked, How accurate can CFD be? This depends on the form of Reynolds-averaged Navier-Stokes partial differential (ReNSPD) equations selected as well as the boundary conditions. Therefore, a range of CFD simulations or CFD experiments must be conducted to verify solution sensitivity to model parameters and boundary conditions embedded in the selected ReNSPD equations and their numerical method calculation.

REFERENCES

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Abu-Ei-Hassan, M. B., Hosni, M. H., and Miller, P. L., Evaluation of Turbulence Effect on Air Distribution Performance Index (ADPI), ASHRAE Transactions, 1996, Part II, pp. 322 – 331. ASHRAE, ASHRAE Handbook 1996, HVAC Systems and Equipment, ASHRAE Inc., Atlanta, GA, 1996. ASHRAE, ASHRAE Handbook 1997, Fundamentals, Atlanta, 1997. Bagheri, H. M., and Gorton, R. L., Verif cation of Stratif ed Air-Conditioning Design, ASHRAE Transactions, 1987a, Part II, pp. 211 – 227. Bagheri, H. M., and Gorton, R. L., Performance Characteristics of a System Designed for Stratif ed Cooling Operating during the Heating Season, ASHRAE Transactions, 1987b, Part II, pp. 367 – 381. Baker, A. J., Kelso, R. M., Gordon, E. B., Roy, S., and Schaub, E. G., Computational Fluid Dynamics: A TwoEdged Sword, ASHRAE Journal, no. 8, 1997, pp. 51 – 58. Bauman, F. S., Zhang, H., Arens, E. A., and Benton, C. C., Localized Comfort Control with a Desktop Task Conditioning System: Laboratory and Field Measurements, ASHRAE Transactions, 1993, Part II, pp. 733 – 749. Brockmeyer, I. H. P., Air Flow Pattern and Its Inf uence on the Economy of Air Conditioning, ASHRAE Transactions, 1981, Part I, pp. 1127 – 1142. Brown, C. E., Spot Cooling / Heating and Ventilation Effectiveness, ASHRAE Transactions, 1988, Part I, pp. 678 – 684. Busweiler, U., Air Conditioning with a Combination of Radiant Cooling, Displacement Ventilation, and Desiccant Cooling, ASHRAE Transactions, 1993, Part II, pp. 503 – 510. Chamberlin, G. A., Schwenk, D. M., Maki, K. S., Li, Z., and Christianson, L. L., VAV Systems and Outdoor Air, ASHRAE Journal, no. 10, 1999, pp. 39 – 47. Dorgan, C. E., and Elleson, J. S., Cold Air Distribution, ASHRAE Transactions, 1988, Part I, pp. 2008 – 2025. Dorgan, C. E., and Elleson, J. S., Design of Cold Air Distribution System with Ice Storage, ASHRAE Transactions, 1989, Part I, pp. 1317 – 1322. Dorgan, C. E., and Elleson, J. S., Design Guide for Thermal Storgage, ASHRAE Inc., Atlanta, 1993. Elleson, J. S., Energy Use of Fan-Powered Mixing Box with Cold Air Distribution, ASHRAE Transactions, 1993, Part I, pp. 1349 – 1358. Fanger, P. O., Melikov, A. K., Hanzawa, H., and Ring, J., Air Turbulence and Sensation of Draught, Energy and Buildings, vol. 12, 1988, pp. 21 – 29. Fanger, P. O., Melikow, A. K., Hanzawa, H., and Ring, J., Turbulence and Draft, ASHRAE Journal, April 1989, pp. 18 – 25. Fields, W. G., and Knebel, D. E., Cost Effective Thermal Energy Storage, Heating / Piping / Air Conditioning, no. 7, 1991, pp. 59 – 72. Fish, W. J., and Faulkner, D., Air Exchange Effectiveness in Off ce Buildings: Measurement Techniques and Results, Reprints of International Symposium on Room Air Convection and Ventilation Effectiveness, July 22 – 24, 1992, University of Tokyo.

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Genter, R. E., Air Distribution for Raised Floor Off ces, ASHRAE Transactions, 1989, Part II, pp. 141 – 146. Gorton, R. L., and Sassi, M. M., Determination of Temperature Prof les and Loads in a Thermally Stratif ed Air Conditioning System: Part 1-Model Studies, ASHRAE Transactions, 1982, Part II, pp. 14 – 32. Hanzawa, H., Melikov, A. K., and Fanger, P. O., Air Flow Characteristics in the Occupied Zone of Ventilated Spaces, ASHRAE Transactions 1987, Part I, pp. 524 – 539. Harmon, J. L., and Yu, H. C., Cold Air Distribution and Concern about Condensation, ASHRAE Journal, no. 5, 1993, pp. 34 – 40. Hart, G. H., and Int-Hout, D., The Performance of a Continuous Linear Diffuser in the Interior Zone of an Open Off ce Environment, ASHRAE Transactions, 1981, Part II, pp. 311 – 320. Heinemeier, K. E., Schiller, G. E., and Benton, C. C., Task Conditioning for the Workplace: Issues and Challenges, ASHRAE Transactions, 1990, Part II, pp. 678 – 689. Int-Hout, D., Air Distribution for Comfort and IAQ, HPAC, no. 3, 1998, pp. 59 – 70. Int-Hout, D., and Weed, J. B., Throw: The Air Distribution Quantif er, ASHRAE Transactions, 1988, Part I, pp. 667 – 677. Knebel, D. E., and John, D. A., Cold Air Distribution, Application, and Field Evaluation of a Nozzle-Type Diffuser, ASHRAE Transactions, 1993, Part I, pp. 1337 – 1348. Kostel, A., Path of Horizontally Projected Heated and Chilled Air Jets, ASHRAE Transactions, 1955, p. 213. Ladeinde, F., and Nearon, M. D., CFD Applications in the HVAC&R Industry, ASHRAE Journal, no. 1, 1997, pp. 44 – 48. Lorch, F. A., and Straub, H. E., Performance of Overhead Slot Diffusers with Simulated Heating and Cooling Conditions, ASHRAE Transactions, 1983, Part I B, pp. 200 – 211. Mathisen, H. M., Case Studies of Displacement Ventilation in Public Halls, ASHRAE Transactions, 1989, Part II, pp. 1018 – 1027. Mayer, E., Physical Causes for Draft: Some New Findings, ASHRAE Transactions, 1987, Part I, pp. 540 – 548. McCarry, B., Innovative Underf oor System, ASHRAE Journal, no. 3, 1998, pp. 76 – 79. Melikov, A. K., Arakelian, R. S., Halkjaer, L., and Fanger, P. O., Spot Cooling — Part 2: Recommendations for Design of Spot-Cooling Systems, ASHRAE Transactions, 1994, Part II, pp. 500 – 510. Miller, Jr., P. L., Application Criteria for the Air Diffusion Performance Index (ADPI), ASHRAE Transactions, 1976, Part II, pp. 206 – 218. Miller, Jr., P. L., Room Air Diffusion Systems: A Re-evaluation of Design Data, ASHRAE Transactions, 1979, Part II, pp. 375 – 384. Sandberg, M., and Blomqvist, C., Displacement Ventilation Systems in Off ce Rooms, ASHRAE Transactions, 1989, Part II, pp. 1041 – 1049. Seppänen, O. A., Fisk, W. J., Eto, J., and Grimsrud, D. T., Comparison of Conventional Mixing and Displacement Air-Conditioning and Ventilating Systems in U.S. Commercial Buildings, ASHRAE Transactions, 1989, Part II, pp. 1028 – 1040. Scof eld, C. M., Low Temperature Air with High IAQ for Tropical Climates, ASHRAE Journal, no. 3, 1993, pp. 52 – 59. Shute, R. W., Integrated Access Floor HVAC: Lessons Learned, ASHRAE Transactions, 1995, Part II, pp. 877 – 886. Sodec, F., and Craig, R., The Underf oor Air Supply System — The European Experience, ASHRAE Transactions, 1990, Part II, pp. 690 – 695. Straub, H. E., and Chen, M. M., Distribution of Air within a Room, for Year-Round Air Conditioning — Part II, University of Illinois, Engineering Experiment Station Bulletin, No. 442, 1957. Straub, H. E., and Cooper, J. G., Space Heating with Ceiling Diffusers, Heating / Piping / Air Conditioning, May 1991, pp. 49 – 55. Svensson, A. G. L., Nordic Experiences of Displacement Ventilation Systems, ASHRAE Transactions, 1989, Part II, pp. 1013 – 1017. Wang, S. K., Air Conditioning, vol. 2, Hong Kong Polytechnic, Hong Kong, 1987. Wang, S. K., Kwok, K. W., and Watt, S. F., Characteristics of a Space Diffusion System in an Indoor Sport Stadium, ASHRAE Transactions, 1985, Part II B, pp. 416 – 435.

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Wendes, H., Supply Outlets for VAV Systems, Heating / Piping / Air Conditioning, February 1989, pp. 67 – 71. Williams, P. T., Baker, A. K., and Kelso, R. M., Numerical Calculation of Room Air Motion — Part 2: The Continuity Constraint Finite Element Method for Three-Dimensional Incompressible Thermal Flows, ASHRAE Transactions, 1994, Part I, pp. 531 – 548. Yuan, X,. Chen, Q., and Clickman, L. R., Performance Evaluation and Design Guidelines for Displacement Ventilation, ASHRAE Transactions, 1999, Part I, pp. 298 – 309. Zhang, J. S., Zhang, R., Li, Z., Shaw, C. Y., Christianson, L. L., and Sparks, L. H., An Experimental Study of the Ventilation Performance of Cold-Air Distribution Systems, ASHRAE Transactions, 1994, Part II, pp. 360 – 367.

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SOUND CONTROL 19.1 SOUND CONTROL AND SOUND PATHS 19.1 Sound Control 19.1 Sound Paths 19.2 Control at Design Stage 19.3 Recommended Procedure for Noise Control 19.3 19.2 FAN, COMPRESSOR, PUMP, AND AIRFLOW NOISE 19.4 Fan Noise 19.4 Noise from Chillers and Pumps 19.4 Airflow Noise 19.5 19.3 SOUND ATTENUATION ALONG DUCT-BORNE PATH 19.6 Sound Attenuation in Ducts 19.6 Sound Attenuation at Elbows and Branch Takeoffs 19.9 End Reflection Loss 19.10 Duct-Borne Crosstalk 19.11 Attenuation along Duct-Borne Path 19.11 19.4 SILENCERS 19.12 Types of Silencers 19.13 Characteristics of Silencers 19.14 Location of Silencers 19.15 Active Silencers 19.16 Selection of Silencers 19.17 19.5 FIBERGLASS IN HVAC&R SYSTEMS 19.17 Problems 19.17 Recommendations 19.18

19.6 RADIATED NOISE AND TRANSMISSION LOSSES 19.18 Breakout and Break-in 19.18 Duct Rumble 19.19 Transmission Losses 19.19 Breakout and Break-in Sound Power Level 19.19 Transmission Loss for Selected Building Structures 19.23 19.7 RELATIONSHIP BETWEEN ROOM SOUND POWER LEVEL AND ROOM SOUND PRESSURE LEVEL 19.23 Single or Multiple Sound Sources 19.23 Array of Ceiling Diffusers 19.24 19.8 NOISE CONTROL FOR A TYPICAL AIR SYSTEM 19.25 Combination of Supply Fan Noise and Terminal Noise 19.25 Estimate Sound Pressure Level for Spaces Served by Terminal Units 19.25 Environmental Adjustment Factor 19.26 Plenum Ceiling Effect 19.26 19.9 ROOFTOP PACKAGED UNITS 19.29 Basics 19.29 Sound Sources and Paths 19.30 Discharge Side Duct Breakout 19.31 Sound Path on Return Side 19.31 Structure-Borne Noise 19.32 REFERENCES 19.32

19.1 SOUND CONTROL AND SOUND PATHS Sound Control Sound control for an air conditioned space or , more accurately, an occupied zone is pro vided to attenuate the HVAC&R equipment sound to an acceptable background level and to provide a suitable acoustic environment for people in the occupied zone. A suitable acoustic environment is as important as a comfortable thermal environment to the occupants. Noise, or any unwanted sound, is always annoying. The objective of sound control should also be to pro vide an unobtrusive background sound at a level low enough that it does not interfere with human activities. Sound quality depends on the relative intensities of the sound le vels in v arious octa ve bands of the audible spectrum. Unobtrusi ve sound quality means the follo wing: a balanced distrib ution over a wide frequenc y range; no tonal characteristics such as hiss, whistle, or rumble; and a steady sound level. 19.1

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The background sound le vel to be pro vided must be optimized. An NC or RC design criterion that is too lo w always means an unnecessary higher cost. The recommended indoor design RC or NC criteria range is listed in Table 4.8.

Sound Paths When HVAC&R equipment or a component generates sound, it is often called a sound generator or sound source. Sound created at the source is recei ved by the occupant through v arious sound transmission paths. HV AC&R equipment that acts as noise generators includes f ans, compressors, pumps, and dampers; this noise is in addition to other airfl w noise in ducts. Of these, the fan is the major noise source in HVAC&R equipment to be controlled and attenuated because f ans are widely used and have more sound transmission paths than compressors and pumps. Sound transmits from the source to the receiver via the following paths: ●

●

●

●

SH__ ST__ LG__ DF

Duct-borne paths . These are noise transmission paths through ducts, duct fittings and other air system components, at either the supply air side or the return air side, as shown in Fig. 19.1. The noise thus transmitted is often called system noise. Radiated sound path. This is the path along which noise radiates through the duct walls or casings into the ceiling plenum or into an adjacent area and then transmits to the receiver. Airborne path. This path includes the equipment noise transmitted to adjacent areas through the air. Structure-borne path. These are the paths along which equipment vibration and noise are transmitted through the building structure.

FIGURE 19.1 Fan room and sound paths.

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Structure-borne transmission is actually a combination of sound and vibration effects. This problem is usually solved through the combined effort of the structural and HVAC&R designers.

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Control at Design Stage The evaluation and control of all noise transmission paths should tak e place at the design stage, and it should be a cooperati ve effort of the architect and the structural and mechanical engineers. If an excessive noise le vel occurs due to improper design — the designer f ails to analyze the potential noise problem, there are mistak es in analysis, or data are used improperly — remedial measures to reduce the noise level are usually expensive and are often less effective. One ef fective measure to reduce noise in b uildings is to locate the f an room or other sound source away from critical areas, such as conference rooms and e xecutive offices. or purposes of noise control, the fan room may be located near rest rooms, stairwells, or copy rooms. For noise control, theoretical estimation and prediction schemes are mainly guidelines. Pre vious experience and field performance are important. Especially for projects with strict acoustic require ments, field/laboratory tests and chec outs are often necessary. Recommended Procedure for Noise Control Before the analysis and e valuation of sound levels, the indoor design NC or RC criteria are usually determined, and the necessary data to e valuate sound le vels at v arious points of interest are collected and investigated. The recommended procedure for noise control uses the following approach: Source : Path (attenuation) : Receiver The basic procedure is as follows: 1. Determine the sources of the noise from the f an, compressor, and pump, and the sound power level generated. Use certified manu acturer’s data. 2. Carefully analyze all the possible sound paths that can transmit noise from the source to the occupied zone. An overlooked sound path may affect the final results 3. Calculate all the sound attenuation and transmission losses in each sound path during transmission. 4. Determine the airfl w noise Laf, in duct-borne path due to dampers, elbow, or junctions and the attenuated f an noise near the damper Lat, fan, both in dB. If Lat, fan  Laf  8 dB, Laf can be ignored. Only when Lat, fan  Laf  8 dB should Laf be added to Lat, fan for duct-borne transmission calculations. 5. For duct-borne paths, determine the attenuated sound po wer level of f an noise at the supply outlet (system noise or room source Lwr), in dB re 1012 W, that affects the receiver. Convert Lwr to a sound pressure level received by the recei ver at a chosen point in a room Lpr or at a plane 5 ft (1.5 m) from the floor Lpt, both in dB re 20 Pa. 6. For radiated sound transmissions or airborne transmissions, the sound po wer le vel of the sound radiated into the receiving room should be converted to a room sound pressure level. 7. The resultant sound pressure level Lp at the center frequency of each octave band, in dB, that is perceived by the recei ver at a chosen point or at a plane 5 ft (1.5 m) from the floor l vel is the sum of sound pressure le vels from all sound paths. Determine the NC or RC LNC, in dB, from the calculated Lp in each octave band. 8. Compare LNC with the design criteria NC or RC Lp, NC. If LNC  Lp, NC, then one should add a silencer (or an acti ve noise control de vice, which is discussed in later sections) to the duct-borne sound path, change the configuration of the duct system or build structures to bring LNC down to or slightly below Lp, NC.

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9. Check this noise control prediction scheme with the actual field results of projects of simila noise control arrangements, if possible.

19.2 FAN, COMPRESSOR, PUMP, AND AIRFLOW NOISE Fan Noise Use the f an noise data pro vided by the manuf acturer. If the manuf acturer’s data are not a vailable, estimate the sound po wer level for the f an by Eq. (15.21). The specific sound p wer levels at fullload operation are listed in Table 15.1. Fans should be selected for maximum efficiency so as to produce less fan noise. When a fan is operated at of f-peak efficiency conditions, according to ASHRAE Handbook 1999 , HVAC Applications, an additional amount up to 6 dB may be added to a plenum f an sound power level operating at the right-hand side of the peak efficiency along the fan curve. Because the calculated system pressure loss is often less than the actual system pressure loss, a centrifugal f an is often selected to operate just at the right-hand side of its rated peak ef ficiency. When a smaller f an is operated at higher volume flow rates, a noise up to 5 dB will often be produced. F ans should not be operated in surge or stall region at any time because fan noise and vibrations are substantially increased. According to ASHRAE Handbook 1999 , HVAC Applications, backward-inclined (BI) or air foil (AF) centrifugal f ans often ha ve higher sound po wer level in the mid- and upper -frequency range especially at the blade passing frequenc y. They are more sensiti ve to inlet fl w distortion than forward-curved (FC) centrifugal f ans. When inlet v anes are installed inside the f an inlet, the sound level at the blade passing frequenc y increases 2 to 8 dB depending on the percentage of the v olume fl w rate reduced. Externally mounted inlet v anes increase about 2 to 3 dB. BI and AF fans show a narrower operating range and a clearly defined su ging region than FC f ans. Inside the operating range, BI and AF fans are usually percei ved to ha ve lower sound po wer level at lo wer frequency (below 100 Hz) than FC f ans. For centrifugal f ans, air volume fl w control by v arying fan speed with PWM ac inverter is the acoustically preferred method. Forward-curved centrifugal f ans ha ve a wide operating range without a clearly defined su ge region. They are comparatively insensitive to fl w distortion at their inlet. In FC fans, modulation of inlet vanes in a VAV system does not cause an increase in sound le vel at a frequenc y of 63 Hz and above. FC fans are considered having 31.5- and 63-Hz rumbles. Plug and plenum f ans need a slightly greater ener gy use, however, they can have a significantl lower sound power level if the fan plenum is properly sized and acoustically treated. Axial fans are considered to ha ve lower sound po wer level in lo wer frequency than centrifugal fans. They are often used in applications where the high-frequenc y noise can be attenuated by silencers. Axial f ans are most sensiti ve to inlet fl w obstructions. Varying blade angle to control volume fl w rate is predominately used in v ane-axial fans for a VAV system when the system load is reduced. When the air volume fl w reduces from 80 to 40 percent, there is a corresponding noise reduction of 2 to 5 dB between 125 and 4000 Hz.

Noise from Chillers and Pumps If manuf acturer’s data are not a vailable, according to ASHRAE Handbook 1991 , HVAC Applications, the sound pressure level LpA, in dBA, for centrifugal chillers at a distance 3.3 ft (1 m) from the chiller can be calculated as SH__ ST__ LG__ DF

L pA  60  11 log TR

(19.1)

where TR  refrigeration capacity, in tons. F or reciprocating chillers, LpA in dBA at a distance of

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3.3 ft (1 m) is L pA  71  9 log TR

(19.2)

The sound pressure level Lp, in dB, at the center frequency of various octave bands can be obtained by adding the following values at each octave band to the calculated LpA: Center frequency of octave bands, Hz Centrifugal chiller Reciprocating chiller

63

125

250

500

1k

2k

4k

8  19

5  11

6 7

7 1

8 4

5 9

8  14

The centrifugal chiller previously mentioned is constructed hermetically, has internal gears, and operates at medium or full load. At light load operation, Lp may increase 10 to 13 dB for the shaft frequency in the band, 8 to 10 dB in the blade pass frequenc y band, and 5 dB for the remaining octave bands. For circulating pumps, LpA at a distance of 3.3 ft (1 m) is L pA  77  10 log hp

(19.3)

where hp  power input to the pump, hp. Airflow Noise Airf ow noise is an aerodynamic noise generated in f ow passages or at duct f ttings. Airf ow noise is the result of v ortices passing around obstacles, such as elbows, dampers, branch takeoffs, terminals, and dif fusers, that cause local acceleration and deformation and thereby produce local dynamic compression of the air. Airf ow noise can also be generated when v ortices pass through solid discontinuities. Airf ow noise is mainly determined by the v elocity of air f owing through the air passage (duct velocities) and the constrictions due to duct f ttings. The geometry of the duct f ttings also af fects the airf ow turbulence and vortices and, therefore, the airf ow noise. Maximum Duct Velocities. Airf ow noise can be reduced by using lo w duct velocities, preventing abrupt changes, and providing suff cient sound attenuation. According to ASHRAE Handbook 1999, HVAC Applications, the maximum recommended air f ow velocities in ducts or in free openings at various specif ed RC (NC) ratings are as follows: Maximum airf ow velocity, fpm RC (NC) In shaft or above drywall ceiling Above suspended acoustic ceiling Within occupied space Free openings

45 35 25 45 35 25 45 35 25 45 35 25

Rectangular duct

Round duct

3500 2500 1700 2500 1750 1200 2000 1450 950

5000 3500 2500 4500 3000 2000 3900 2600 1700

Supply outlet

Return inlet

625 500 350

750 600 425

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Branch ducts should have air velocities about 80 percent of the listed v alue. Airf ow velocity of the f nal runout to outlets should be 50 percent of the listed v alue. The presence of the elbo w, diffuser, and other f ttings can increase air f ow noise substantially , and therefore the duct v elocity should reduce accordingly. Diffusers and Grilles. The manufacturer’s sound rating of a dif fuser is obtained with a uniform velocity distribution throughout the neck. If a v olume damper is installed immediately before the diffuser, the turbulent airf ow at the neck will be signi f cantly increased compared to the manuf acturer’s laboratory-tested data. This turbulence can be considerably reduced by adding an equalizing grid in the neck of the dif fuser. Volume dampers should not be located closer than 5 ft (1.5 m) from a supply outlet. The sound power level added to the dif fuser sound rating is proportional to the damper pressure ratio (DPR). The DPR is equal to the throttled pressure loss across the damper , divided by the minimum pressure loss across the damper when its blades are fully opened. According to the data provided by ASHRAE Handbook 1999 , HVAC Applications, if a v olume damper is installed in the neck of a linear dif fuser, when the DPR varies from 1.5 to 6, a corresponding 5 to 24 dB should be added to the dif fuser sound rating. If the v olume damper is installed in the inlet of the plenum of a linear diffuser, when DPR varies from 1.5 to 6, a corresponding of 2 to 9 dB should be added. If the volume damper is installed at least 5 ft (1.5 m) from the inlet plenum of a linear diffuser, only a corresponding of 0 to 5 dB should be added. Airf ow noise generated at the dif fusers or grilles at the end of duct-borne paths is dif f cult to attenuate except by reducing their neck or f ace velocities. Imbalance of v olume f ow between various branch runouts may create a greater branch duct v elocity and grille f ace velocity in one of the branch runouts which has the lowest f ow resistance. Poor Fan Entry and Disc harge Conditions. Noise is generated because of the abrupt air passage constrictions and sudden changes in air f ow direction. Both cause f ow turbulence and f ow separation as well as energy losses. Such low-frequency noise results in a duct rumble and is v ery diff cult to attenuate. The designer should carefully design the f an intake and discharge connections. Doing so is the best w ay to control this kind of air f ow noise as well as to pro vide an ener gy-eff cient air system.

19.3 SOUND ATTENUATION ALONG DUCT-BORNE PATH Sound Attenuation in Ducts

SH__ ST__ LG__ DF

Sound attenuation is the reduction in the intensity of sound, expressed in w atts per unit area, as it travels along a sound transmission path from a source to a recei ver. Sound attenuation is achie ved by (1) the absorption of sound ener gy by the absorpti ve material, (2) spherical spreading and scattering, and (3) ref ection of sound waves incident upon a surface. Sound attenuation in duct sections, duct f ttings, silencers, and other sound-reducing elements can be indicated by insertion loss. Insertion loss (IL) at a speci f c frequenc y is the reduction in sound power level, in dB re 10 12 W, measured at the receiver when a sound attenuation element is inserted in the transmission path between the sound source and the recei ver. Noise reduction is the difference in sound pressure levels between any two points along the sound transmission path. Sound absorpti vity is the ability of a material to absorb sound ener gy. When a sound w ave impinges on the surf ace of a porous sound-absorbing material, air vibrates within the small pores. The f ow resistance of air and its vibration con verts a portion of the absorbed sound ener gy to heat. The fraction of the incident sound po wer that is absorbed is called the sound absorption coef f cient . Most sound-absorbing materials have a low at low frequencies and a higher at high frequencies. For a typical sound absorption material, might be equal to 0.15 in the octa ve band whose center frequency is 63 Hz, and  0.9 at 1000 Hz.

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TABLE 19.1 Approximate Natural Attenuation of Sound in Unlined Rectangular Sheet-Metal Ducts Attenuation, dB / ft Octave band center frequency, Hz

Duct size, in.  in.

P / A, 1 / ft

63

125

250

 250

66 12  12 12  24 24  24 48  48 72  72

8.0 4.0 3.0 2.0 1.0 0.7

0.30 0.35 0.40 0.25 0.15 0.10

0.20 0.20 0.20 0.20 0.10 0.10

0.10 0.10 0.10 0.10 0.07 0.05

0.10 0.06 0.05 0.03 0.02 0.02

P / A ratio indicates perimeter / area ratio. Source: ASHRAE Handbook 1995, HVAC Applications. Reprinted with permission.

Ducts can be unlined or inner -lined with foil-coated encapsulated f berglass or other soundabsorptive materials, usually at a thickness of 1 in. (25 mm), sometimes 2 in. (50 mm), to absorb fan noise. Inner -lined ducts ha ve better sound attenuation than unlined ducts, and the inner -lined sound-absorbing layer can also serve as thermal insulation. In unlined round ducts, the sound attenuation, called natural attenuation, is only about one-tenth that in unlined rectangular ducts as follows: Round duct attenuation, dB/ft Octave band center frequency, Hz Diameter, in. D7 7 D  15 15 D  30 30 D  60

125 0.03 0.03 0.02 0.01

250

500

 1000

0.05 0.03 0.02 0.01

0.05 0.05 0.03 0.02

0.10 0.07 0.05 0.02

Source: ASHRAE Handbook 1995, HVAC Applications. Reprinted with permission.

Natural attenuation of unlined rectangular ducts is listed in Table 19.1. For inner-lined ducts, the sound attenuation, or insertion loss, of a gi ven length of duct section depends on the cross-sectional area and the properties of the absorpti ve material. Sound attenuation in inner-lined round duct with a 1-in. (25-mm) thickness of duct liner is listed in Table 19.2, and for 1-in-. (25-mm-) thick inner-lined straight rectangular ducts it is listed in Table 19.3. TABLE 19.2 Round Duct Attenuation (Insertion Loss) Octave band center frequency, Hz Duct diameter, in. 6 12 24 48

63

125

0.38 0.23 0.07 0

0.59 0.46 0.25 0

250

500

1000

2000

4000

8000

Approximate attenuation for 1-in. duct liner, dB / ft 0.93 0.81 0.57 0.18

1.53 1.45 1.28 0.63

2.17 2.18 1.71 0.26

2.31 1.91 1.24 0.34

Source: Abridged with permission from ASHRAE Handbook 1995, HVAC Applications.

2.04 1.48 0.85 0.45

1.26 1.05 0.80 0.44

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TABLE 19.3 Sound Attenuation (Insertion Loss) for Straight Lined Sheet-Metal Rectangular Ducts, Lining Thickness 1 in. Internal cross-sectional dimensions, in. 88 8  18 12  12 12  24 18  18 18  36 24  24 24  48 36  36 36  72 48  48 48  72

Octave and center frequency, Hz 63

125

250

0.77 0.60 0.56 0.50 0.40 0.30 0.30 0.23 0.21 0.16 0.16 0.14

0.90 0.67 0.60 0.48 0.43 0.32 0.32 0.24 0.21 0.16 0.16 0.13

1.29 1.04 0.96 0.79 0.74 0.59 0.59 0.47 0.43 0.34 0.34 0.29

500

1000

2000

4000

8000

5.94 4.72 4.29 3.41 3.10 2.46 2.46 1.95 1.78 1.41 1.41 1.22

3.49 3.06 2.90 2.54 2.41 2.11 2.11 1.85 1.75 1.54 1.54 1.42

2.42 2.26 2.20 2.06 2.00 1.87 1.87 1.76 1.71 1.60 1.60 1.54

Attenuation, dB / ft 2.57 2.22 2.09 1.81 1.70 1.47 1.47 1.27 1.20 1.03 1.03 0.94

6.23 5.10 4.70 3.85 3.54 2.90 2.90 2.37 2.19 1.79 1.79 1.58

1. Based on measurements of surface-coated duct liners of 1.5 to 3 lb / ft3 density. For the specif c materials tested, liner density had a minor effect over the nominal range of 1.5 to 3 lb / ft3. 2. Add natural attenuation (Table 19.1) to obtain total attenuation. Source: Abridged with permission from ASHRAE Handbook 1995, HVAC Applications.

TABLE 19.4 Lined Flexible Duct Insertion Loss Insertion loss dB Diameter, in. 8

9

10

12

14

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Length, ft 12 9 6 3 12 9 6 3 12 9 6 3 12 9 6 3 12 9 6 3

Octave band center frequency, Hz 63

125

250

500

1000

2000

4000

8 6 4 2 8 6 4 2 8 6 4 2 7 5 3 2 5 4 3 1

11 8 6 3 11 8 6 3 10 8 5 3 9 7 5 2 7 5 4 2

21 16 11 5 22 17 11 6 22 17 11 6 20 15 10 5 16 12 8 4

33 25 17 8 33 25 17 8 32 24 16 8 30 23 15 8 27 20 14 7

37 28 19 9 37 28 19 9 36 27 18 9 34 26 17 9 31 23 16 8

37 28 19 9 36 27 18 9 34 26 17 9 31 23 16 8 27 20 14 7

24 18 12 6 22 17 11 6 21 16 11 5 18 14 9 5 14 11 7 4

Source: ASHRAE Handbook 1995, HVAC Applications. Reprinted with permission.

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19.9

Lined f exible ducts reduce duct-borne noise signi f cantly. Insertion losses for lined f exible ducts at various diameters and duct lengths are listed in Table 19.4. The length of the f exible duct should normally be from 3 to 6 ft (0.9 to 1.8 m). An abrupt bend may create unacceptably high noise. The insertion loss (sound attenuation) of lined round, f at oval, rectangular, and f exible ducts is a function of their inside diameter (or the smaller inside dimension of rectangular and f at o val ducts). Generally, as the inside diameter increases, the insertion loss decreases. Surprisingly, the 3 : 1 aspect ratio f at oval duct has approximately the same IL as a round duct with the same diameter as the minor axis. The thickness of the duct liner has a def nite effect on the IL of lined ducts. For duct liner of rectangular duct with a thickness of 2 in. (50 mm), the sound attenuation is nearly double the v alue of a 1-in. (25-mm) duct liner for octave band center frequencies between 250 and 500 Hz. The ILs are nearly the same for both round and rectangular duct liners with a thickness of 1 in. (25 mm) for octave band center frequencies of 1000 Hz and greater. According to Bodle y (1981), the density of the duct liner material has a minor ef fect on the IL of inner-lined ducts. The IL is af fected by the absorption coef f cient of the lining material and its thickness.

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Sound Attenuation at Elbows and Branch Takeoffs Sound is re f ected by elbo ws. Lined or unlined elbo ws pro vide sound attenuation to reduce the duct-borne noise. Sound attenuation depends on the size of duct, whether the elbow has a duct lining before and / or after the elbo w, and whether turning v anes are pro vided. The estimated sound attenuation for unlined round elbows is indicated in Table 19.5. A mitered 90° elbow, often called a square elbow, provides effective noise attenuation by means of re f ection loss in the octa ve bands whose center frequencies equal or e xceed 125 Hz. Rumble noise is diff cult to attenuate even for a square elbow. In Table 19.6 are listed the insertion losses for elbows that are followed or proceeded by duct lining. Square elbows with turning vanes are slightly less effective in sound attenuation than elbows without turning vanes. For any branch takeoff from a diverging wye in a main duct, the dominating sound attenuation is the acoustic energy distribution taking place between the main duct and the branch ducts. That part of sound attenuation because of the branch po wer division Lw,b, in dB, is proportional to the ratio of the area of branch Abr, in ft2 (m2), to the sum of the area of branches Abr, in ft2 (m2), and can be calculated as L w,b  10 log

Abr Abr

(19.4)

TABLE 19.5 Insertion Loss of Unlined Round Elbows Insertion loss, dB fw 1.9 1.9 fw 3.8 3.8 fw 7.5 fw  7.5

0 1 2 3

fw  f  w f  frequency, kHz w  diameter for round duct, in. Source: ASHRAE Handbook 1995, HVAC Applications. Reprinted with permission.

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TABLE 19.6 Insertion Loss of Square Elbows Insertion loss, dB Unlined

Lined

Without turning vanes fw 1.9 1.9 fw 3.8 3.8 fw 7.5 7.5 fw 15 15 fw 30 fw  30

0 1 5 8 4 3

0 1 6 11 10 10

With turning vanes fw 1.9 1.9 fw 3.8 3.8 fw 7.5 7.5 fw 15 fw  15

0 1 4 6 4

0 1 4 7 7

fw  f  w f  frequency, kHz w  duct width or depth of square duct or diameter for round duct, in. Source: ASHRAE Handbook 1995, HVAC Applications. Reprinted with permission.

The sound attenuation of branch po wer division Lw,b should be subtracted from the sound po wer level of duct-borne noise at the common end of the di verging wye in the main duct just before the branch takeoff Lw,c, in dB, to obtain the sound power level of the branch takeoff Lw,b, in dB, L w,b  L w,c  L w,b

(19.5)

If the minor sound attenuations in the di verging wye are ignored, based on the principle of ener gy conservation, the sum of the sound po wer levels of the duct-borne noise at the straight-through end Lw,s, in dB, and the branch tak eoff Lw,b just after the di verging wye Lw,s  Lw,b  Lw,c, from Eq. (19.5); therefore L w,s  L w,b

(19.6)

End Reflection Loss

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When a sound wave propagates through the end opening of a duct section into a room, a signif cant amount of sound is re f ected back to the duct, and thus substantially reduces the lo w-frequency sound that reaches the recei ver in the occupied zone. According to ASHRAE Handbook 1995 , HVAC Applications, the sound attenuation of end re f ection loss for ducts terminated in free space and for ducts terminated f ush with a w all is listed in Table 19.7. Dif fusers that terminate in a suspended acoustic ceiling can be considered as terminating in free space. In Table 19.7, for a diffuser with a rectangular connecting duct, its diameter D  4A / , in in. (mm), where A  area of the rectangular duct, in.2 (mm2). Values listed in Table 19.7 were tested based on straight round connecting ducts leading to the diffusers. In actual installation, there may be dampers and guide v anes. The connecting duct between a dif ffuser and a main duct or terminal may not be straight and smooth. All these f actors may affect the listed values in Table 19.7.

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TABLE 19.7 Duct End Ref ection Loss End ref ection loss, dB Octave band center frequency, Hz

Duct diameter, in.

63

6 8 10 12 16 20 24 28 32

20 18 16 14 12 10 9 8 7

6 8 10 12 16 20 24 28 32

18 16 14 13 10 9 8 7 6

125

250

500

1000

2000

2 1 1 1 0 0 0 0 0

1 0 0 0 0 0 0 0 0

Duct terminated in free space 14 12 11 9 7 6 5 4 3

9 7 6 5 3 2 2 1 1

5 3 2 2 1 1 1 0 0

Duct terminated f ush with wall 13 11 9 8 6 5 4 3 2

8 6 5 4 2 2 1 1 1

4 2 2 1 1 1 0 0 0

1 1 1 0 0 0 0 0 0

Source: ASHRAE Handbook 1995, HVAC Applications. Reprinted by permission.

Duct-Borne Crosstalk Duct-borne crosstalk or plenum crosstalk is the sound transmitted between tw o rooms through the duct systems or ceiling plenum. Usually , the sound attenuation in the duct system or in the ceiling plenum should be 5 dB greater than the transmission loss of the architectural partition between tw o rooms. Transmission loss is discussed in detail in the following sections. Duct lining, devious duct runs, f exible ducts, inner lining in the slot dif fusers, and return slots are often used to attenuate the sound transmitted through duct-borne or plenum-connected paths to avoid crosstalk. A well-sealed architectural partition in the ceiling plenum with an inner -lined return air passage effectively prevents crosstalk through the ceiling plenum for applications with strict NC or RC criteria in the occupied zone. Attenuation along Duct-Borne Path For a typical VAV system with a supply f an installed in the AHU inside a f an room, as shown in Fig. 19.1, the sound attenuation along the duct-borne paths — the supply duct side and the return duct side — are as follows: Supply duct-borne path D1 Supply side silencer Supply duct attenuation Elbows attenuation

Return duct-borne path D2 Return side silencer Return duct attenuation Elbows attenuation

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Supply duct-borne path D1 Branch power division Flexible duct attenuation Room effect

Return duct-borne path D2 Ceiling plenum attenuation or branch power division Return slot attenuation Room effect

On the return air side, an air f lter made of a f brous material may have a certain degree of sound attenuation. This could be considered an added safe f actor in the return duct-borne path. Ceiling plenum attenuation is discussed in the following section. All sound attenuation should be subtracted from the sound po wer level of the supply f an Lw,f, in dB re 10 12 W, and con verted to the sound pressure le vel at the recei ver Lp,r, in dB re 20 Pa, through the room effect.

19.4 SILENCERS The purpose of a silencer is to reduce the sound po wer level of a f an, an air f ow noise, or other sound source transmitted along a duct-borne path or airborne path to a required level.

SH__ ST__ LG__ DF

FIGURE 19.2 Rectangular and cylindrical silencers: (a) rectangular silencer and (b) cylindrical silencer.

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Types of Silencer

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Silencers can be classi f ed into rectanguar, cylindrical, and sound-attenuating plenum according to their conf guration. A typical rectangular splitter silencer is shown in Fig. 19.2a. Inside the retangular casing are a number of f at splitters, depending on the width of the silencer. These splitters direct the airf ow into small sound-attenuating passages. The splitter is made from an envelope containing sound-attenuating material, such as f berglass or mineral w ool, with protected noneroding f acing. The thickness of a splitter is often between 1 and 4 in. (25 and 100 mm). Splitters often ha ve a round instead of a f at nose, to reduce their air f ow resistance. A rectangular silencer is often connected with retangular ducts or sometimes with rectangular fan intakes and discharges. A c ylindrical silencer has an outer c ylindrical jack et and an inner concentric center body , as shown in Fig. 19.2 b. Both the cylindrical jacket and the center body contain sound-attenuating material and noneroding facing. A cylindrical silencer is often used in conjunction with vane-axial fans and in round duct systems. A sound-attenuating plenum consists of se veral splitters in the form of tw o, three, or four successive square elbows, shown in Fig. 19.3, f lled with sound-absorbing material with noneroding

FIGURE 19.3 Sound-attenuating plenum.

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facing. A sound plenum effectively attenuates low-frequency noise and is often located between the collective return grille and the return air intak e of a packaged unit. A low velocity is important in a sound-attenuating plenum because of the high local loss coeff cient of square elbows. Silencers used in HV AC&R systems can again be classi f ed according to the means of sound attenuation into following types: 1. Dissipative silencers. These silencers often use f ace-covered or encapsulated acoustic material, such as f berglass, mineral wool, and acrylic polymers, to attenuate noise o ver a broad range of frequencies. The facing material can be made of Galvanized or aluminum sheet with a perforated area not e xceeding 22 percent of the f ace area Acrylic polymers or polymer sheets 2. Packless silencers. There is no f brous f ll. Noise is attenuated by means of acoustically resisti ve perforations in the splitters. They are often made of sintered aluminum or acrylic plastics. 3. Reflection-dissipative silence s. These silencers use the combined ef fect of sound re f ection and dissipation in airf ow passages of successive square elbows. 4. Active silencers. These silencers produce lo w-frequency inverse sound w aves to cancel the unwanted noise. ●

●

Characteristics of Silencers The acoustic and aerodynamic characteristics of a silencer are mainly indicated by four parameters.

SH__ ST__ LG__ DF

FIGURE 19.4 Difference in insertion loss of a rectangular silencer between forw ard and re verse f ow (at 2000 fpm or 10 m / s f ace v elocity). (Source: Handbook of HVAC Design, 1990. Reprinted by permission.)

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Insertion Loss. This is the capacity of a silencer to reduce the sound po wer level of fan noise or other noise at v arious octave band center frequencies. The IL of a silencer is af fected by the airf ow direction, especially when air v elocity is greater than 200 fpm (1 m / s). A sound w ave that propagates in the same direction as the air f ow is said to be in forw ard f ow, and one that propagates opposite to the air f ow is said to be in re verse f ow. At low frequencies, reverse f ow has a longer contact time and, therefore, a higher IL. At high frequencies, sound waves tend to refract toward the absorptive surface in a silencer under reverse f ow. Therefore, these effects increase the high-frequency attenuation in the forw ard f ow and decrease the attenuation in re verse f ow. T he difference in IL between forw ard and reverse f ow of a rectangular silencer at 2000 fpm (10 m / s) face velocity is presented in Fig. 19.4. Operating Parameters. ●

●

●

●

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The operating parameters for a silencer are as follows:

Volume f ow rate of air through the silencer V˙ sil, in cfm (m3 /s) Cross-sectional area of the silencer, including the free area Afree, ft2 (m2), i,e., area of the passages through which air f ows, and the cross-sectional area of the silencer Asil, ft2 (m2) Free area ratio Rfree, which is def ned as Rfree  Afree /Asil Face velocity vsil, in fpm (m / s), def ned as vsil  V˙ sil /Asil

Many silencers ha ve a free air ratio between 0.3 and 0.8. The actual mean air v elocity inside the air f ow passages in a silencer vfree, in fpm (m / s), may be 1.25 to 3.3 times the f ace v elocity vsil. The f ace v elocity of a silencer and sound-attenuating plenum is 500 to 2000 fpm (2.5 to 10 m / s). Self-Noise. This is the lo wer limit of sound po wer level, in dB, that a speci f c silencer can approach at various octave band center frequencies. Pressure Drop psil. This is the total pressure drop of airstream when it f ows through a silencer. Pressure drop psil is a function of its f ace velocity, free area ratio, length, and the conf guration of the splitter or the center body . In general, ASHRAE Handbook 1995 , HVAC Applications, recommends that if psil  0.35 in. WC (87 P a), both vfree and air f ow noise should be investigated.

Location of Silencers The location of a silencer to attenuate fan noise should meet the following two requirements: 1. There must be a minimum distance Lsil between the upstream f an discharge outlet or other duct f ttings and the silencer, to ensure a uniform approach velocity at silencer inlet or an undisturbed discharging velocity at the silencer e xit. From the f an discharge, Lsil must be equal to or greater than the distance of one duct diameter for e very 1000 fpm (5 m / s) average duct velocity. From the f an intak e, Lsil should be equal to or greater than 0.75 duct diameter for e very 1000 fpm (5 m / s) average duct velocity. 2. The silencer should straddle or be adjacent to the solid w all of the f an room (or mechanical room), as shown in Figs. 16.7 a and 16.8a. Such a setup can pre vent or reduce the follo wing fan noise transmissions: The fan noise that has not been attenuated by the silencer will not break out into the room adjacent to the fan room. The noise in the f an room may ha ve the chance to break in to the supply main duct after the silencer. ●

●

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TX Fan noise

Input microphone

Error microphone

DDC controller Loudspeaker

FIGURE 19.5 An active silencer.

Active Silencers Active silencers use ducted enclosure to cancel duct-borne, low-frequency f an noise (including rumbles) by producing sound w aves of equal amplitude and opposite phase. The primary sound source is the unw anted fan noise. The secondary sound source which cancels the unw anted source comprises the inverse sound waves from a loudspeaker. Operating-Characteristics. An acti ve silencer consists of tw o or more microphones, a microprocessor-based controller, a loudspeaker, and a ducted enclosure. When a f an noise is propagated down along a duct, as shown in Fig. 19.5, an input microphone located in the airstream measures the noise and sends an electric signal proportional to the sound w ave to the controller . The microprocessor-based controller calculates the amplitude, frequency, and phase of the propagating sound and sends a cancel signal to the loudspeak er. The loudspeaker broadcasts sound w aves of the same amplitude and frequency as the unwanted noise, but 180° out of phase. The destructive interference between these tw o sound sources results in the cancellation of the incident f an noise by the secondary source broadcasted by the loudspeak er. The amplitude of the f an noise is reduced do wnstream of the loudspeak er. An error microphone measures the residual noise. This feedback is used to optimize the performance of the active silencer. Frequency Limits. Gelin (1997) noted that in an acti ve silencer, the ducted enclosure acts as a waveguide for the sound w aves. At lower frequencies, when the wavelength of the sound is at least 2 times longer than the greatest dimension of the w aveguide, sound waves may propagate in plain waves in which sound pressure and the phase of the sound w ave are uniform across an y duct cross section. At intermediate and high frequencies, the wavelength is smaller than a w aveguide dimension, and the sound waves may propagate as plain waves as well as more complicated sound waves. A low-frequency fan noise is easily measured, def ned, and reproduced 180° out of phase, whereas for a high-frequenc y noise it is v ery diff cult to do these things. Therefore, active silencers w ork best at an octave band center frequency range of 31 to 250 Hz. System Characteristics. ●

SH__ ST__ LG__ DF

●

The following are the system characteristics of active silencers:

In an active silencer, sound energy is added to the f an noise to cancel it, whereas in a traditional passive silencer (such as dissipative and packless silencers), sound energy is converted to heat and removed from the system. Because the microphone in an active silencer cannot distinguish the fan noise and the turbulent airf ow noise, it is acceptable practice that the air velocity in the ducted enclosure of an active silencer

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●

●

19.17

not exceed 1500 fpm (7.5 m / s). When the duct v elocity exceeds 2500 fpm (12.5 m / s), the turbulent noise may completely mask fan noise, and the effect of an active silencer is reduced to nil. Pressure drop of a passi ve silencer is often 0.2 to 0.35 in. WC (50 to 63 P a). The electric energy required to produce the canceling sound is only 40 W, a substantial saving compared to a passi ve silencer. Due to the de velopment of the microprocessor -based technologies, in 1997, the price of the controller used in an active silencer dropped substantially compared to that in 1987.

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Performance. Gelin (1997) compared the insertion loss, in dB, of an active silencer, a passive silencer, and a duct liner for a large fan of 5000 cfm (2360 L / s) volume f ow rate as follows: Octave bands, Hz Active silencer 1-in. Lined, 84-in.- (2.1-m-) long duct 84-in. (2.1-m) Prefabricated silencer

31 5–8

63 10 – 15 2 8

125 10 – 15 2 17

250 2– 4 4 26

500 — 10 43

1000 — 20 43

2000 — 17 26

4000 — 15 19

The pressure drop of the 84-in. (2.1-m) prefabricated silencer is 0.24 in. WC (60 Pa). The active silencer has a good sound attenuation in frequencies between 31 and 125 Hz, the duct liner provides effective sound attenuation in frequencies of 500 Hz and more, and the 84-in. (2.1-m) prefabricated silencer is effective in frequencies between 63 and 4000 Hz.

Selection of Silencers During the selection of silencers, the following points should be considered: ●

●

●

Dominating frequencies of noise. For a low-frequency fan “rumble” noise, an active silencer or a ref ective dissipative silencer should be selected. F or a broadband noise, a combination of acti ve and dissipative silencers or a dissipati ve silencer with thick splitters, of 2- to 4-in. (50- to 100mm) thickness, is preferable. Design requirements. For clean rooms, f ber-free all-metal silencers are often selected. F or health care facilities, f ber-free packless silencers should be selected. Wet surfaces. If there are wet surfaces in the air system, f ber-free all-metal silencers should be selected.

For a specif c type of silencer, variations in its length and free area ratios often meet most of the requirements for IL and pressure drops. Silencers are usually made in lengths of 3, 5, 7, and 10 ft (0.9, 1.5, 2.1, and 3 m). The IL and psil of a silencer are proportional to its length. A lower face velocity and greater free air ratio for a specif c silencer at a given length always result in a lo wer pressure drop and therefore lo wer energy consumption. If space is allo wed, a lower vsil is benef cial for both IL and psil.

19.5 FIBERGLASS IN HVAC&R SYSTEMS Problems Fiberglass-made duct liners, duct boards, silencers, and insulation layers ha ve been widely used in HVAC&R systems for decades. On June 24, 1994, the U.S. Department of Health and Human Services (DHHS) announced that f berglass will be listed as a material “reasonably anticipated to be a

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carcinogen.” According to ASHRAE, recently the use of f berglass in se veral institutional, eduational, and medical projects has been banned or se verely limited because of the concerns that the f bers may be carcinogenic and f berglass may promote microbial growth. The International Agency for Research on Cancer (IARC) has performed e xtensive research concerning the carcinogenicity of f berglass materials and found inadequate evidence to link insulation of f berglass with cancer in humans. IARC does indicate a possible link to cancer from glass wool based on hea vy dosage in animals. The Canadian En vironment Protection Act indicated that studies have failed to show any evidence that f berglass is carcinogenic and classif ed it as “unlikely to be carcinogenic to humans.” Concerning the f ber erosion, more recent studies reported in the 6th International Conference on Indoor Air Quality and Climate in Helsinki, Finland, 1993 that airborne f ber levels in buildings with these products range from nondetectable up to only 0.006 f ber per cm3 (1 in.3  16.4 cm3). Both liquid water and dirt are required for mold and microbial growth, and microbial growth can occur on any surface in an HVAC&R system. Tinsley (1998) concludes that “. . . the presence of high humidity air , by itself, is not suf f cient to cause fungal gro wth. . . . fungal growth occurred only where both standing w ater and nutrients (in the form of dirt, pollen, or other or ganics) could be found.” Control of the wetted surf ace and dirt in the duct system is the most ef fective method to reduce potential microbial growth. Recommendations 1. For health care facilities, schools, semiconductor fabricating, food processing, pharmaceutical manufacturing, clean rooms, and many more demanding projects, f berglass duct liners and f berglass silencers should be replaced by any one or a combination of the following alternatives: ●

●

●

●

Foil-coated products such as acrylic polymer-coated f berglass duct liners or silencers Fibrous acoustic inf ll (f berglass or mineral wool) encapsulated inside the polymer sheets Fiber-free sound absorber using sintered aluminum or acrylic polymer with microperforations Active silencers

2. For air systems that ha ve wetted surf aces (standing w ater) from outdoor air intak es, cooling coils, coil condensate droplet carryo ver, drain pans, and humidif cation processes, especially atomizing devices, direct evaporative coolers, and air washers, all-metal f ber-free silencers or acti ve silencers should be used to attenuate duct-borne f an and airf ow noise, if required. Interior surface of ducts that contact with supply air must be galv anized metal sheet. Duct insulation should be wrapped externally. 3. For air systems that emplo yed draw-through fans and medium- or higher -eff ciency air f lters and had no wetted surf aces (standing water) in contact with the air in the air system (e xcept in the coil section and condensate pan) ha ving a f ace v elocity of the cooling or DX coil of 500 fpm (2.5 m / s) or less, and properly designed, installed, operated, and maintained, f berglass duct liners and f berglass silencers with perforated metal facing sheet can be used. Periodic monitoring of the moisture and dirt conditions of the f berglass duct liner and silencers must be provided.

19.6 RADIATED NOISE AND TRANSMISSION LOSSES Break-out and Break-in SH__ ST__ LG__ DF

When a duct section enters the ceiling plenum or directly passes o ver occupied spaces carrying duct-borne fan noise that is not well attenuated, the noise radiates through the duct w all, raises the

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sound power level in the ceiling plenum or the occupied zone, and is transmitted to the recei ver. Noise that radiates through the duct w all and causes the duct wall to vibrate, caused either by internal sound w aves or by air f ow turbulence, is called break out. Noise can also be transmitted into a duct and then tra vel along the duct-borne path, either dischar ging into a space through the duct opening or breaking out into ambient air, where it is transmitted to the receiver. The transmission of external noise into a duct section through the duct wall is called break-in. Breakout fan noise from the dischar ge duct directly under a rooftop packaged unit into the ceiling plenum may cause a serious acoustic problem in the underlying conditioned space.

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Duct Rumble Duct rumble is the result of air pressure f uctuations caused by v ariations in the speed of the f an, fan motor, or f an belt or due to the air f ow instabilities transmitted to the f an housing or nearby ductwork. When the air pressure f uctuations are e xerted on lar ge, f at, and unreenforced duct surfaces that have resonance frequencies near or equal to the f uctuating frequencies, the duct surfaces vibrate. Poor f an outlet connections often generate duct rumbles. Duct rumble can produce sound pressure levels of 65 to 95 dB at frequencies from 10 to 100 Hz. Changing the fan speed as well as the frequencies of the air pressure f uctuations is one method to reduce duct rumble. Another method is to increase the rigidity of the duct w all and thus change the resonance frequencies. Lo w-frequency duct rumble noise is v ery diff cult to attenuate by duct liners. Transmission Loss Transmission loss (TL) is the reduction of sound power when sound is transmitted through a wall, a partition, or a barrier. The relationship between transmission loss and mass of the material indicates that the sound transmission loss of a homogeneous solid partition is a function of its mass and the frequency of sound f transmitted through it. This relationship can be e xpressed in the follo wing form: TL  F(log m s, log f )

(19.7)

where ms  surface density of homogeneous partition, lb / ft (kg / m ). Concrete walls and galvanized sheet ducts show that for each doubling of the mass partition, TL increases about 2 to 3 dB for low-frequency noise (less than 800 Hz), and TL increases about 5 to 6 dB for high-frequency sound (800 Hz and over). 2

2

The TL-mass relationship is v alid only when sound is incident on the surf ace of the partition in a normal direction and is an approximation. Actual measured data may show considerable deviation from the predicted v alues. Factors that cause de viations are nonhomogeneity, cracks, stiffness, and resonance. Cracks and gaps around doors, windows, duct and piping slee ves, or other openings on the partition or w all may considerably reduce TL. The TL of a well-sealed partition of 100 ft 2 (9.3 m2) surface area may drop from 40 to 20 dB because of a total of 1 ft 2 (0.093 m 2) of openings on that partition. Breakout and Break-in Sound Power Level According to ASHRAE Handbook 1999 , HVAC Applications, the breakout sound power level radiated from the outer surface of duct walls Lw,out, in dB, can be calculated as L w,out  L w,in  10 log

S  TL out Ai

(19.8)

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FIGURE 19.6 Breakout transmission loss TLout.

where Lw,in  sound power level of sound inside duct, dB S  surface area of outer sound-radiating duct surface, in.2 (mm2) Ai  inner cross-sectional area of duct, in.2 (mm2) TLout  duct breakout transmission loss, dB Equation (19.8) assumes no sound attenuation along unlined interior duct surf ace. Such a break out transmission model is shown in Fig. 19.6.

TABLE 19.8 Duct Breakout Transmission Loss TLout for Round Duct at Various Octave Band Center Frequencies TLout, dB Diameter, in.

Length, ft

Octave band center frequency, Hz Gauge

63

125

250

500

1000

2000

4000

55 54 37 26

52 36 33 26

44 34 33 24

35 31 27 22

34 25 25 38

 75 55 26 36 28

72 33 26 32 25

56 34 25 32 26

56 35 22 28 24

46 25 36 41 40

Long seam ducts 8 14 22 32

15 15 15 15

26 24 22 22

 45  50  47 (51)

(53) 60 53 46 Spiral-wound ducts

8 14 26 26 32

SH__ ST__ LG__ DF

10 10 10 10 10

26 26 24 16 22

 48  43  45  48  43

 64  53 50 53 42

Note: In cases where background sound swamped the sound radiated from the duct walls, a lower limit on TLout is indicated by a  sign. Parentheses indicate measurements in which background sound has produced a greater uncertainty than usual. Source: ASHRAE Handbook 1995, HVAC Applications. Reprinted by permission.

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TABLE 19.9 TLout for Rectangular Ducts at Various Octave Bands, dB Duct size

__RH

Octave band center frequency, Hz

in.

Gauge

63

125

250

500

1000

2000

4000

8000

12  12 12  24 12  48 24  24 24  48 48  48 48  96

24 24 22 22 20 18 18

21 19 19 20 20 21 19

24 22 22 23 23 24 22

27 25 25 26 26 27 25

30 28 28 29 29 30 29

33 31 31 32 31 35 35

36 35 37 37 39 41 41

41 41 43 43 45 45 45

45 45 45 45 45 45 45

The data are tests on 20-ft-long ducts, but the TL values are for ducts of the cross section shown regardless of length. Source: Adapted with permission from ASHRAE Handbook 1991, HVAC Applications.

As a result of duct sound break out, the sound pressure level in an occupied space Lp, in dB, can be calculated as L p  L w,out  10 log rL  10

(19.9)

where r  distance between duct and receiver’s position where Lp is calculated, ft (m) L  length of duct sound-radiating and breakout surface, ft (m) The break-in sound power level transmitted into duct and then transmitted upstream or do wnstream Lw,in, in dB, can be calculated as L w,in  L w,out  TL in  3

(19.10)

where TLin  duct break-in transmission loss, dB. Table 19.8 lists the duct break out transmission losses TLout for round ducts at v arious octa ve band center frequencies, Table 19.9 lists TLout for rectangular ducts, and Table 19.10 lists TLout for f at oval ducts. Also Table 19.11 lists the duct break-in transmission losses TLin for round ducts at various octave band center frequencies, Table 19.12 lists TLin for rectangular ducts, and Table 19.13 lists TLin for f at oval ducts. In lower frequencies, round ducts have the highest TLout and TLin, and rectangular ducts have the lowest TLout and TLin among these three kinds of ducts.

TABLE 19.10 TLout for Flat Oval Ducts at Various Octave Bands, dB TLout, dB

Duct size (a  b)

Octave band center frequency, Hz

in.

Gauge

63

125

250

500

1000

2000

4000

8000

12  6 24  6 24  12 48  12 48  24 96  24 96  48

24 24 24 22 22 20 18

31 24 28 23 27 22 28

34 27 31 26 30 25 31

37 30 34 29 33 28 —

40 33 37 32 — — —

43 36 — — — — —

— — — — — — —

— — — — — — —

— — — — — — —

The data are tests on 20-ft-long ducts, but the TL values are for ducts of the cross section shown regardless of length. Source: Adapted with permission from ASHRAE Handbook 1991, HVAC Applications.

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TABLE 19.11 TLin for Round Ducts at Various Octave Bands, dB TLin, dB Octave band center frequency, Hz

Diameter, in.

Length, ft

Gauge

63

8 14 22 32

15 15 15 15

26 24 22 22

17 27 28 (35)

8 14 26 26 32

10 10 10 10 10

26 26 24 16 22

20 20 27 30 27

125

250

500

1000

2000

4000

Long seam ducts (31) 43 40 36

39 43 30 23

42 31 30 23

41 31 30 21

32 28 24 19

31 22 22 35

59 44 20 30 25

62 28 23 29 22

53 31 22 29 23

43 32 19 25 21

26 22 33 38 37

Spiral-wound ducts 42 36 38 41 32

Note: In cases where background sound swamped the sound radiated from the duct walls, a lower limit on TLin is indicated by a  sign. Parentheses indicate measurements in which background sound has produced a greater uncertainty than usual. Source: ASHRAE Handbook 1995, HVAC Applications. Reprinted by permission.

TABLE 19.12 TLin for Rectangular Ducts at Various Octave Bands, dB Duct size

Octave band center frequency, Hz

in.

Gauge

63

125

250

500

1000

2000

4000

8000

12  12 12  24 12  48 24  24 24  48 48  48 48  96

24 24 22 22 20 18 18

16 15 14 13 12 10 11

16 15 14 13 15 19 19

16 17 22 21 23 24 22

25 25 25 26 26 27 26

30 28 28 29 28 32 32

33 32 34 34 36 38 38

38 38 40 40 42 42 42

42 42 42 42 42 42 42

The data are tests on 20-ft-long ducts, but the TL values are for ducts of the cross section shown regardless of length. Source: Adapted with permission from ASHRAE Handbook 1991, HVAC Applications.

TABLE 19.13 TLin for Flat Oval Ducts at Various Octave Bands, dB TLin, dB

SH__ ST__ LG__ DF

Octave band center frequency, Hz

Duct size in. in.

Gauge

63

125

250

500

1000

2000

4000

12  6 24  6 24  12 48  12 48  24 96  24 96  48

24 24 24 22 22 20 18

18 17 15 14 12 11 19

18 17 16 14 21 22 28

22 18 25 26 30 25 —

31 30 34 29 — — —

40 33 — — — — —

— — — — — — —

— — — — — — —

Note: The data are for duct lengths of 20 ft, but the values may be used for the cross section shown regardless of length. Source: ASHRAE Handbook 1995, HVAC Applications. Reprinted by permission.

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TABLE 19.14 Transmission Loss of Some Building Structures Octave band center frequency, Hz Building structures

63

125

250

500

1000

2000

4000

4-in. dense concrete or solid concrete block, 48 lb / ft2 4-in. hollow-core dense aggregate concrete block, 28 lb / ft2 8-in. hollow-core dense aggregate concrete block Standard drywall partition, 58 -in. gypsum board on both sides of 2-in.  4-in. wood studs Standard drywall partition, two layers of 58 -in. gypsum board on each side of 3 58 -in. metal studs

32 29 31

34 32 33

35 33 35

37 34 36

42 37 41

49 42 48

55 49 54

12

17

34

35

42

38

44

25

36

43

50

50

44

55

2 -in. plate glass Double glazing, two 12 -in. panes, 12 -in. airspace

11 12

16 16

23 23

27 27

32 32

28 30

32 35

9

15

20

25

31

33

27

25

41

47

56

65

68

69

10

5 15

9 21

10 25

12 27

14 26

15 27

15

18

21

39

38

49

55

23

27

29

27

26

29

1

5 8 -in. Gypsum board ceiling Roof construction, 6 in. thick, 20 gauge (0.0396-in). steel deck with 4-in. lightweight concrete topping, 58 -in. gypsum board ceiling on resilient hangers *Plenum / ceiling *Plenum / ceiling cavity effect: lay-in mineral f ber tile 5 8 in., 35 lb / ft3 *Plenum / ceiling cavity effect: f nished sheetrock, 58 in.

Acoustic equipment housing, 20-gauge steel outer shell, 2-in.-thick acoustic insulation, 22-gauge (0.0336-in.) perforated inner shell Solid-core wood door, normally closed† * ASHRAE Transactions, 1989, Part I. † Handbook of HVAC Design, 1990. Source: ASHRAE Handbook 1991, HVAC Applications.

Transmission Loss for Selected Building Structures The transmission losses for selected building structures — including walls, partitions, window glass, ceiling and plenum, and acoustic equipment housing — are listed in Table 19.14. These are mainly abridged from ASHRAE Handbook 1991 , HVAC Applications . In Table 19.14, the combined plenum / ceiling cavity effect indicates the combined ef fect of the plenum sound absorption and the transmission through the ceiling material. These values are based on data from se veral manufacturers’ laboratory and mock-up spaces.

19.7 RELATIONSHIP BETWEEN ROOM SOUND POWER LEVEL AND ROOM SOUND PRESSURE LEVEL The sound pressure level at a given location in a room corresponding with a particular point source or an array of multiple sound sources is a function of the sound po wer level and transmission characteristic of the sound source, the room’s acoustical properties, and the distance between the sound source and the receiver.

Single or Multiple Sound Sources Based on f eld measurements, Schultz (1985) recommended the follo wing empirical formula to estimate the room sound pressure le vel Lpr, in dB re 20 Pa, from the room sound po wer level of

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single or multiple sound sources at a chosen point in a room: L pr  L wr  5 log V  3 log f  10 log r  25

(19.11)

where Lwr  sound power level of room source, dB re 1012 W V  volume of room, ft3 (m3) f  octave band center frequency, Hz r  distance from source to receiver or reference point, ft (m) For a single sound source in the room, Eq. (19.11) can be applied directly . If there are multiple sound sources, estimate Lwr for each sound source and sum the contrib utions to obtain the total sound pressure level at the receiver due to multiple sound sources. The difference between the room sound pressure level and the room sound power level Lpr  Lwr is called the room effect.

Array of Ceiling Diffusers Off ces and lar ge rooms in commercial b uildings often ha ve several ceiling dif fusers mounted in a distributed array on the ceiling. Although the sound pressure le vels decrease as the v ertical distance from the ceiling increases, the sound pressure le vel along an y selected horizontal plane, such as a reference plane of 5 ft (1.5 m) from the f oor le vel, is nearly constant. According to ASHRAE Handbook 1995, HVAC Applications, the calculation of sound pressure le vel for a distributed array of ceiling dif fusers at a reference plane of 5 ft (1.5 m) from the f oor level can be greatly simplif ed as L p5  L w,s  D

(19.12)

where Lw,s  sound po wer le vel of single outlet, dB re 10 12 W. V alues of D are listed in Table 19.15. In Eqs. (19.11) and (19.12), the calculation of sound pressure le vel Lp is based on the assumption that the acoustic characteristics of a room range from a verage to medium dead, which is generally true for most rooms. F or rooms that are acoustically medium-li ve to li ve (ha ve little sound absorption), Eqs. (19.11) and (19.12) can overestimate the decrease in Lp.

TABLE 19.15 Values for D in Eq. (19.12) Value for D, dB Floor area per diffuser, ft2

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Octave band center frequency, Hz 63

125

250

500

1000

2000

4000

Ceiling height 8 to 9 ft 100 to 150 200 to 250

2 3

3 4

4 5

5 6

6 7

7 8

8 9

Ceiling height 10 to 12 ft 150 to 200 250 to 300

4 5

5 6

6 7

7 8

8 9

9 10

10 11

Ceiling height 14 to 16 ft 250 to 300 350 to 400

7 8

8 9

9 10

10 11

11 12

12 13

13 14

Source: ASHRAE Handbook 1995, HVAC Applications. Reprinted by permission.

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19.8 NOISE CONTROL FOR A TYPICAL AIR SYSTEM

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In many commercial buildings using a v ariable-air-volume air system emplo ying VAV boxes, there are two noise sources, as shown in Fig. 19.1: ●

●

Fan noise due to the supply fan of the air-handling unit usually installed in a fan room Airf ow noise due to the VAV box with a single-blade damper installed inside the ceiling plenum

Sound is transmitted from these two sources to the receiver in the conditioned space via the follo wing paths: ●

●

●

●

●

Duct-borne path, supply side Duct-borne path, return side Radiated noise breakout from the supply duct in the ceiling plenum Radiated sound breakout from the VAV box casing in the ceiling plenum Airborne path through fan room walls to adjacent area

Combination of Supply Fan Noise and Terminal Noise Variable-air-volume box es are tested for radiated sound po wer le vel and dischar ge-side sound power level in accordance with ADC / ARI Industry Standard 880-83, and the results are listed in the manufacturers’ catalogs. The attenuated duct-borne sound po wer level of the supply f an from the air-handling unit at terminal unit LAHUi should be added to the discharge sound power level from the VAV box Lter to form a resultant sound po wer level if their dif ference Lter  LAHUi  8 dB. If their difference is greater than 8 dB, the smaller of these two sound power levels can be ignored.

Estimate Sound Pressure Level for Spaces Served by Terminal Units Recently, the ADC and ARI jointly de veloped Industry Standard 885 to estimate the space sound pressure levels when terminal units are installed inside the ceiling plenum. Consider a terminal unit connected to tw o slot dif fusers through f exible ducts to serv e a conditioned space, as sho wn in Fig. 19.7, where C represents the terminal casing, D the duct-borne or discharge side, D1 and D2 the discharge outlets 1 and 2, and O1 the supply outlet 1. Another outlet, O2, is not shown in Fig. 19.7. Along with the arro ws, B represents the transmission loss of the break out from duct and casing, P the sound attenuation due to the plenum ceiling, and S the room effect. There are altogether se ven sound paths by which the recei ver can hear f an, terminal, and slot diffuser noises. Path 1: Radiated sound power level of the terminal unit : plenum ceiling : room effect : receiver Path 2: Discharge sound po wer level : breakout : plenum ceiling : room effect : receiver Path 4: Duct-borne sound po wer level at the outlet of slot dif fuser : room effect : receiver Path 5: Sound power level of the slot diffuser : room effect : receiver Sound path 3 transmits radiated sound break out from another f exible duct after the terminal unit, and it is similar to path 2. The duct-borne noise is transmited along sound path 6 at another slot diffuser and is similar to sound path 4. The airf ow noise of another slot diffuser transmits along sound path 7 and is similar to sound path 5. The radiated sound power level that breaks out at the inlet of the terminal unit has the same kind of sound path as paths 2 and 3. The sound power level of duct-borne noise at Do1 equals the sound power level at D1 or D2 minus the sound attenuation of the f exible duct and the slot diffuser.

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FIGURE 19.7 Sound path of a typical terminal unit installed in the ceiling plenum.

Environmental Adjustment Factor According to Industry Standard (IS) 880, sound power levels of the terminal unit as listed in the manufacturers’ catalogs are based on free f eld (outdoor) calibration of the reference sound source. At low frequencies, actual rooms are highly re verberant. The adjustment f actor that tak es into account the difference in sound power levels of a commonly used reference sound source measured in a free f eld and a re verberant f eld is called the environmental adjustment factor E f, in dB. IS 885 recommends the following Ef values: Band frequency Ef, dB, re 10

12

W

63

125

250

500

1000

2000

4000

8000

7

3

2

1

1

1

1

1

Plenum Ceiling Effect Sound attenuation of the plenum ceiling combination not only includes the transmission loss of the ceiling, but also considers the ef fect of the plenum. F or a 58 -in.- (16-mm-) thick mineral f ber acoustic tile ceiling with a density of 35 lb / ft3 (560 kg / m3), IS 885 recommends the following values, which are the same as those in Table 19.14: SH__ ST__ LG__ DF

Octave band center frequency, Hz

125

250

500

1000

2000

4000

Plenum ceiling effect, dB

5

9

 10

 12

 14

 15

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Example 19.1. A v ariable-air-volume box (terminal unit) recei ves conditioned air from an air handling unit through main and branch ducts. The VAV box is installed inside a ceiling plenum and is connected to two slot diffusers by 10-in.- (250-mm-) diameter f exible ducts, each of which is 5 ft (1.5 m) long. Information is a vailable from manuf acturers’ catalogs on the duct-borne f an noise at the inlet of the VAV box, the radiated sound po wer le vel Lw,rad, the dischar ge sound po wer le vel Lw,dis of the VAV box, and the sound po wer level of the slot dif fuser Lw,s. From measurements, the transmission loss of the breakout noise through f exible duct wall and the insertion loss of the f exible duct are known. These data are summarized as follows:

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Octave band center frequency, Hz Fan noise at inlet, dB Lw,rad, dB Lw,dis, dB Lw,s, dB Flexible, TLout, dB Flexible, IL, dB / ft

125

250

500

1000

2000

4000

65 60.5 63 40 8 0.50

63 55 60.5 39  11 0.85

61 50.5 47 35  14 1.48

52 46.5 47.5 33  17 2.2

46 41.5 54.5 31  20 2.04

45 37.5 52 24  25 1.64

If the volume of the room is 5040 ft 3 (143 m 3), estimate the total sound pressure le vel in the occupied zone with an a verage distance from the recei ver of 8 ft (2.4 m). Ignore the f an noise transmitted through the return duct system. Solution 1. From Eq. (19.11), the room effect at the octave band center frequency of 125 Hz of this occupied zone is L pr  L wr   5 log V  3 log f  10 log r  25   5 log 5040  3 log 125  10 log 8  25   8.8 dB The room effect of other octave bands can be estimated similarly as follows: Octave band center frequency, Hz

125

250

500

1000

2000

4000

Room effect, dB

9

 9.5

 11

 11.5

 12.5

 13.5

2. For path 1, after considering the in f uence of the en vironmental adjustment f actor Ef , one f nds that the radiated sound power levels of the VAV box Lrad,r  Lw,rad-Ef, in dB, are as follows: Octave band center frequency, Hz

125

250

500

1000

2000

4000

Reverberant f eld, dB

57.5

53

49.5

45.5

40.5

36.5

Because the transmission loss of the casing of the VAV box is rather high, the break out of ductborne fan noise through the casing is far smaller than Lrad,r, and is ignored. The sound pressure le vel in the occupied zone due to the radiated sound po wer level from the VAV box can be calculated by subtracting the plenum ceiling ef fect from Lrad,r and then converting to the sound pressure level through the room effect.

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Octave band center frequency, Hz Lrad,r, dB Ceiling plenum effect, dB Room effect, dB

125

250

500

1000

2000

4000

57.5 5 9

53 9  9.5

49.5  10  11

45.5  12  11.5

40.5  14  12.5

36.5  15  13.5

43.5

34.5

28.5

22

14

Total sound pressure level, dB

8

3. For path 2, the discharge sound power level of the VAV box changes from the free f eld Lw,dis to the reverberant f eld Ldis,r and the duct-borne fan noise at the terminal unit Lw,f. Their combination is as follows: Octave band center frequency, Hz Ldis,r, dB Fan noise Lw,f, dB Ldis,r  Lw,f, dB

125

250

500

1000

2000

4000

60 65 66.5

58.5 63 64.5

46 61 61.5

46.5 52 53.5

53.5 46 54.5

51 45 52

The sound pressure level in the occupied zone due to the combination Ldis,r  Lw,f transmitted along sound path 2 can be estimated by subtracting the transmission loss of the f exible duct w all, the plenum ceiling effect, and the room effect from Ldis,r  Lw, f . Octave band center frequency, Hz Ldis,r  Lw,f , dB Flexible, TLout, dB Ceiling-plenum, dB Room effect, dB Sound pressure level, dB

125

250

500

1000

2000

4000

66.5 8 5 9

64.5  11 9  9.5

61.5  14  10  11

53.5  17  12  11.5

54.5  20  14  12.5

52  25  15  13.5

44.5

35

26.5

13

8

0

4. For path 4, there are two f exible ducts and slot diffusers connected to the terminal unit; therefore, the branch power division is 3 dB. The sound pressure level in the occupied zone due to ductborne noise transmitted along sound path 4 can be e valuated by subtracting the branch po wer division Lw,b, the insertion loss of the f exible duct IL f ex, and the room ef fect from the combination Ldis,r  Lw,f . Octave band center frequency, Hz Ldis,r  Lw,f, dB Lw,b, dB ILf ex, dB Room effect, dB Sound pressure level, dB

SH__ ST__ LG__ DF

125

250

500

1000

2000

4000

66.5 3  2.5 9

64.5 3 4  9.5

61.5 3 7  11

53.5 3  11  11.5

54.5 3  10  12.5

52 3 8  13.5

52

48

28

29

40.5

27.5

5. For path 5, change the sound power level of the slot diffuser from free f eld Lw,s to reverberant f eld Ls,r.

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Then the sound pressure level in the occupied zone due to the slot diffuser can be estimated by subtracting the room effect from Ls,r.

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Octave band center frequency, Hz 125 Ls,r, dB Room effect, dB

1000

2000

4000

37 9

37  9.5

34  11

32  11.5

30  12.5

23  13.5

28

27.5

23

20.5

17.5

9.5

Sound pressure level, dB

250

500

6. The resultant estimated sound pressure level in the occupied space Lp,t, in dB re 20 Pa, is the sum of the contributions of all sound paths and can be calculated as follows: Octave band center frequency, Hz 125

250

500

1000

2000

4000

Path 1 Path 2 Path 3 Path 4 Path 5 Path 6 Path 7

43.5 44.5 44.5 52 52 28 28

34.5 35 35 48 48 27.5 27.5

28.5 26.5 26.5 40.5 40.5 23 23

22 13 13 28 28 20.5 20.5

14 8 8 29 29 17.5 17.5

8 0 0 27.5 27.5 9.5 9.5

Lp,t, dB

56

51.5

44

32.5

32.5

31

These resultant sound pressure levels at the center frequencies of v arious octave bands are all equal to or below the NC 40 curve.

19.9 ROOFTOP PACKAGED UNITS Basics Rooftop packaged units are widely used in lo w-rise commercial b uildings. Because of the use of large rooftop packaged units, many rooftop packaged units are curb-mounted on lightweight roof deck construction. The distances between the fan and ceiling plenum, and between the fan and conditioned space, are sometimes shorter than the distance between the f an in the AHU and space. They may cause unique noise control problems. Objectionable noise from a rooftop packaged unit usually includes direct transmission of fan and compressor noise through the building roof and ceiling, structure-borne vibration, duct-borne fan noise, duct breakout noise, and duct rumble. Locating the rooftop packaged unit at least 25 ft (7.6 m) horizontally away from areas with critical sound control requirements, such as conference rooms and e xecutive off ces whose NC curv e should be 35 dB or less, is the most ef fective and economical method of noise control. A rooftop packaged unit can be located o ver buffer spaces, low-use storage areas, high-noise loading areas, and acoustically nonsensitive spaces such as rest rooms and stairwell buffer spaces. Using a number of small packaged units instead of one or tw o lar ge units di vides the impact over a greater roof area and is sometimes bene f cial. A rooftop packaged unit with a lo wer sound power level is often one size lar ger than the smallest unit for that capacity . The fan in such a unit is properly sized, with less restricted inlet and outlet conditions inside the unit.

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In a typical rooftop packaged unit, the paths by which fan and compressor noise is transmitted from the rooftop packaged units to the receiver in the occupied space are shown in Fig. 19.8a: 1. Radiated f an noise breaks out from the supply duct in the ceiling plenum : plenum ceiling effect : room effect : receiver. 2. Duct-borne f an noise on the dischar ge side : attenuation of ducts, elbows, and silencer : end ref ection : room effect : receiver. 3. Radiated fan noise breaks out from the return air duct in the ceiling plenum : plenum ceiling effect : room effect : receiver. 4. Duct-borne f an noise on the return side : attenuation of ducts, elbows, and silencer : end ref ection : room effect : receiver.

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FIGURE 19.8 Sound path and sound control for a typical rooftop packaged unit: (a) installation of rooftop unit with rectangular supply duct; ( b) rectangular to multiple drop: round mitered elbo ws with turning v anes; (c) rectangular to multiple drop: round mitered elbows with turning vanes, inner-lined round duct after mitered elbow.

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5. Structure-borne f an or compressor noises : roof construction : plenum ceiling ef fect : receiver.

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Discharge Side Duct Breakout Experience has shown that the low-frequency noise breakout from the supply duct inside the ceiling plenum (path 1) and subsequent transmissions to the recei ver through the ceiling are the critical noise problem in most rooftop packaged units. A supply f an that has a do wn shot outlet and that penetrates the roof directly produces many noise problems, such as very loud breakout sound inside the ceiling plenum. A supply f an with horizontal outlet dischar ge to an inner -lined plenum inside the rooftop packaged unit is better . Use 2-in. (50-mm) acoustic linnings for f an, compressor, and condenser sections to reduce sound transmission directly through the roof. An even better sound attenuation arrangement is to run the supply duct horizontally from the rooftop unit 10 to 20 ft (3 to 6 m), combined with silencers along the roof, and then penetrate into the b uilding, directly above acoustically sensitive spaces. Duct e xposed to outside atmosphere must be inner -lined, externally insulated, weatherproof enclosures that are covered with noise barriers. Round ducts ha ve a f ar greater TLout than rectangular ducts. Based on laboratory tests, Beatty (1987) recommended using rectangular to multiple round duct drops with round mitered elbo ws and turning vanes, as shown in Fig. 19.8 b and c to contain low-frequency rumble noise. Compared with rectangular ducts without turning v anes, these conf gurations have an additional noise reduction of 17.5 dB in octa ve band 1, 11.5 dB in octa ve band 2, and 13 dB in octa ve band 3 when the conf guration is the same as in Fig. 19.8b. If the multiple drops of round duct have the conf guration shown in Fig. 19.8c, there will be an additional noise reduction of 18 dB in octa ve band 1, 13 dB in octave band 2, and 16 dB in octave band 3 compared to rectangular ducts without turning vanes. Use heavier-gauge sheet metal to f abricate the rectangular plenum before the round duct drops to reduce breakout noise. If the f rst run of duct from the f an outlet must be rectangular and passes a noise sensiti ve area, Harold (1993) recommended the following: ●

●

●

●

Use 16-gauge sheet metal for duct wall for the f rst 25-ft (7.5-m) run of duct. Use 1500 fpm (7.5 m / s) as the maximum duct velocity for the rectangular duct. Enclose the supply duct in a dryw all enclosure of tw o layers of 58-in. (16-mm) gypsum board with freestanding w alls that ha ve a 4-in. (100-mm) airspace around the duct and contain a 3-in.(75-mm-) thick f berglass blanket that touches only the duct or drywall. Employ an active silencer.

Careful design of the fan outlet conf gurations will prevent sharp turning of airstreams. Inner lining of the ductw ork and sometimes an additional silencer are necessary at the do wnstream side for more effective noise control. The duct system should not be designed for an e xcessive velocity, i.e., normally not greater than 3000 fpm (15 m / s). Design the duct f ttings at minimum dynamic losses. VAV boxes and reheating VAV boxes in a VAV reheat system should not be operated at a completely shut-off condition because a high f an total pressure at a lo w volume f ow rate is generally a noisy operating point for centrifugal fans.

Sound Path on Return Side On the return side, there are three types of return air systems: ●

A rooftop unit whose return air system connects directly the return ceiling plenum to the return intake of the rooftop unit through a short return duct system. This is the type of return air system

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●

●

widely used for rooftop units. A rectangular return duct may be used if its duct w all is made of 16-gauge (1.6-mm thickness) galv anized sheet metal and break out noise is not critical. Adopt a duct velocity of 1500 fpm (7.5 m / s) or less. Use 2-in. (50-mm) thick inner lining in the return side (including ductwork) of the rooftop unit. Use tees with a minimum leg length of 3 to 5 equivalent diameters to increase end ref ection loss. For ducted return air systems, the sound power level at the supply or return f an intake should be attenuated by the return duct liners, return side silencer, and room effect. For a rooftop unit with a collective return grille, a sound-attenuating plenum, as shown in Fig. 19.3, is often cost-effective if space is available. For an open fan room (Fig. 16.7), return air is often transferred to the open f an room from the return ceiling plenum through a transfer duct or a transfer wall passage. Both require an inner lining 2 in. (50 mm) thick with a perforated metal sheet facing. Air velocity in the return transfer duct or passage should be 1000 fpm (5 m / s) or less.

Structure-Borne Noise Structure-borne noise can be controlled by using suitable vibration isolators. Special curb-mounting bases should be used to provide better sound and vibration control. Sound structural support should be pro vided for rooftop packaged units. Borzym (1991) says that the best place for the rooftop packaged unit is directly o ver a column or straddling a major beam close to a column. Good locations are those near the intersection of beam with a heavy end close to a column, straddling a beam, or close to a beam. The rooftop packaged unit should be mounted on a spring-isolated roof curb and should f oat on the springs. The roof below the rooftop packaged unit should not vibrate. The roof directly under the rooftop packaged unit within the roof curb should contain a good acoustic barrier to pre vent sound transmission through the bottom panel of the packaged unit, the roof, and the ceiling plenum. A typical sound barrier is made of semirigid f berglass insulation, at least 4 in. (100 mm) thick, covered by tw o layers of 58 -in. (16-mm) gypsum board. The staggered gypsum boards are used to seal the joints and seams airtight.

REFERENCES

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ASHRAE, ASHRAE Handbook 1991, HVAC Applications, ASHRAE Inc., Atlanta, GA, 1991. ASHRAE, ASHRAE Handbook 1995, HVAC Applications, Atlanta, 1995. ASHRAE, ASHRAE Handbook 1999, HVAC Applications, Atlanta, 1999. Beatty, J., Discharge Duct Conf gurations to Control Rooftop Sound, Heating / Piping / Air Conditioning, July 1987, pp. 53 – 58. Blazier, W. E., Noise Rating of Variable-Air-Volume Terminal Devices, ASHRAE Transactions, 1981, Part I, pp. 140 – 152. Bodley, J. D., An Analysis of Acoustically Lined Duct and Fittings and the State of the Art of Their Use, ASHRAE Transactions, 1981, Part I, pp. 658 – 671. Borzym, J. X., Acoustical Design Guidelines for Location of Packaged Rooftop Air Conditioners, ASHRAE Transactions, 1991, Part I, pp. 437 – 441. Cummings, A., Acoustic Noise Transmission through Duct Walls, ASHRAE Transactions, 1985, Part II A, pp. 48 – 61. Ebbing, C., and Waeldner, W. J., Industry Standard 885: An Overview, Estimating Space Sound Levels for Air Terminal Devices, ASHRAE Transactions, 1989, Part I, pp. 529 – 533. Ebbing, C. E., Fragnito, D., and Inglis, S., Control of Low Frequency Duct-Generated Noise in Building Air Distribution Systems, ASHRAE Transactions, 1978, Part II, pp. 191 – 203. Gelin, L. J., Active Noise Control: A Tutorial for HVAC Designers, ASHRAE Journal, no. 8, 1997, pp. 43 – 49.

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Grimm, N. R., and Rosaler, R. C., HVAC Systems and Components Handbook, McGraw-Hill, New York, 1997. Harold, R. G., Achieving a Quiet Rooftop Installation, ASHRAE Journal, no. 12, 1993, pp. 32 – 38. Hirschorn, M., Acoustic and Aerodynamic Characteristics of Duct Silencers for Air Handling Systems, ASHRAE Transactions, 1981, Part I, pp. 625 – 646. Phelps, G. R., Letters to Editor — Fibrous Facts, Engineered Systems, no. 10, 1997, pp. 13 – 16. Reese, J., The Case for Certif ed Ratings, Heating / Piping / Air Conditioning, August 1986, pp. 85 – 87. Reynolds, D. D., and Bledsoe, J. M., Sound Transmission through Mechanical Equipment Room Walls, Floor, or Ceiling, ASHRAE Transactions, 1989, Part I, pp. 83 – 89. Schultz, T. J., Relationship between Sound Power Level and Sound Pressure Level in Dwellings and Off ces, ASHRAE Transactions, 1985, Part I A, pp. 124 – 153. Sessler, S. M., and Angevine, E. N., HVAC System Noise Calculation Procedure — An Update, ASHRAE Transactions, 1983, Part II B, pp. 697 – 709. Smith, M. C., Industry Standard 885, Acoustical Level Estimation Procedure Compared to Actual Acoustic Levels in an Air Distribution Mock-Up, ASHRAE Transactions, 1989, Part I, pp. 543 – 548. Smith, M. C., and Int-Hout, D., Using Manufacturers’ Catalog Data to Predict Ambient Sound Levels, ASHRAE Transactions, 1984, Part II B, pp. 85 – 96. Tinsley, R. W., Duct Liner: Problem or Solution? HPAC, no. 5, 1998, pp. 65 – 70. The Trane Company, Acoustics in Air Conditioning, American Standard Inc., La Crosse, WI, 1986. Ver, I. L., Noise Generation and Noise Attenuation of Duct Fittings — A Review: Part II, ASHRAE Transactions, 1984, Part II A, pp. 383 – 390.

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__SH __ST __LG DF

CHAPTER 20

AIR SYSTEMS: BASICS AND CONSTANT-VOLUME SYSTEMS 20.1 AIR SYSTEM BASICS 20.2 Air Systems 20.2 Air Distribution Systems 20.3 Ventilation Systems 20.3 Mechanical Ventilation Systems 20.3 Regenerative Systems 20.3 Smoke Control Systems 20.4 Terminals 20.4 Primary, Secondary, and Transfer Air 20.4 20.2 BUILDING LEAKAGE AREA AND BUILDING TIGHTNESS 20.5 Effective Leakage Area 20.5 Building Air Leakage 20.5 Air Change Per Hour at 50 Pa (ACH50) 20.6 20.3 SPACE PRESSURIZATION 20.7 Space Pressure Characteristics 20.7 Stack Effect 20.7 Stack Effect for High-Rise Buildings 20.9 Wind Effect 20.10 Air Systems and Mechanical Ventilation Systems 20.11 Airflow Balance and Space Pressurization by Differential Flow 20.11 20.4 INFILTRATION AND EXFILTRATION 20.14 Volume Flow Rate of Infiltration 20.14 20.5 FAN-DUCT SYSTEMS 20.14 System Operating Point 20.15 Buckingham  Method and Fan Laws 20.15 20.6 SYSTEM EFFECT 20.17 Mechanism of System Effect 20.17 Inlet System Effect 20.18 Inlet System Effect Loss 20.19 Inlet System Effect Loss Coefficient 20.19 Outlet System Effect 20.20 Outlet System Effect Loss Coefficient 20.22 Selecting Fan Considering System Effect Losses 20.23 20.7 COMBINATION OF FAN-DUCT SYSTEMS 20.24 Two Fan-Duct Systems Connected in Series 20.25 Fans Combined in Parallel and Connected in Series with a Duct System 20.26 Two Parallel Fan-Duct Systems Connected with Another Duct System 20.28

20.8 MODULATION OF THE FAN-DUCT SYSTEM 20.31 Modulation Curve 20.31 Modulation of Fan-Duct Systems 20.33 20.9 CLASSIFICATION OF AIR SYSTEMS 20.39 20.10 CONSTANT-VOLUME SYSTEMS 20.40 System Characteristics 20.40 Energy per Unit Volume Flow 20.41 20.11 AIR CONDITIONING PROCESSES 20.41 Sensible Heat Ratio 20.41 20.12 SPACE CONDITIONING AND SENSIBLE COOLING AND HEATING PROCESSES 20.43 Space Conditioning Process 20.43 Sensible Heating and Cooling Processes 20.44 20.13 HUMIDIFYING PROCESSES 20.45 Steam Injection and Heating Element Humidifier 20.45 Air Washer 20.46 Oversaturation 20.46 20.14 COOLING AND DEHUMIDIFYING PROCESS 20.46 20.15 ADIABATIC MIXING AND BYPASS MIXING PROCESSES 20.50 Two-Stream Adiabatic Mixing Process 20.50 Bypass Mixing Process 20.52 20.16 SINGLE-ZONE, CONSTANT-VOLUME SYSTEMS — COOLING MODE OPERATION 20.53 Air Conditioning Cycle 20.53 Cooling Mode Operation in Summer 20.53 Cooling Mode Operation in Winter without Space Humidity Control 20.56 Cooling Mode Operation in Winter with Space Humidity Control 20.57 Part-Load Operation and Controls 20.58 Outdoor Ventilation Air and Exhaust Fans 20.58 20.17 SUPPLY VOLUME FLOW RATE 20.59 Based on Space Cooling and Heating Load 20.59 Based on Requirements Other Than Cooling Load 20.60 Rated Volume Flow of Supply and Return Fans 20.61

20.1

20.2

CHAPTER TWENTY

20.18 DETERMINATION OF THE SUPPLY AIR CONDITION 20.62 Air Conditioning Rules 20.63 Graphical Method 20.63 Influence of Sensible Heat Ratio 20.64 20.19 CONSTANT-VOLUME SINGLE-ZONE SYSTEMS — HEATING MODE OPERATION 20.69 Heating Mode without Space Humidity Control 20.69 Heating Mode with Space Humidity Control 20.71 Part-Load Operation 20.73

Dual-Thermostat Year-Round Zone Temperature Control 20.73 20.20 CONSTANT-VOLUME MULTIZONE SYSTEM WITH REHEAT 20.74 Reheating, Recooling, and Mixing 20.74 Constant-Volume Multizone System with Reheat 20.75 Control Systems 20.75 Operating Parameters and Calculations 20.77 System Characteristics 20.78 REFERENCES 20.79

20.1 AIR SYSTEM BASICS Air Systems Air systems and their controls directly af fect the indoor en vironment and indoor air quality (IA Q). In a broad sense, air systems are a group of subsystems in an air conditioning system. Air systems inclued supply and return air systems (space recirculating systems), mechanical ventilation systems, air distrib ution systems, regenerative systems, smoke control systems, and terminals. The main functions of air systems are as follows: ●

●

●

●

●

Conditioning the supply air including heating or cooling, humidification or dehumidificatio cleaning and purifying, and attenuation of objectional noise produced by f ans, compressors, and pumps Distributing the conditioned supply air with adequate outdoor air to the conditioned space, extracting space air for recirculating, and exhausting or relieving unwanted space air to the outdoors Providing space pressurization, toxic gas e xhaust, and smok e control for occupants’ safety and fire protectio Controlling and maintaining required space temperature, humidity, cleanliness, air mo vement, sound level, and pressure dif ferential within predetermined limits at optimum ener gy consumption and cost Using high-temperature airstream to reactivate the desiccant, if any

In actual practice or in a narro wer sense, an air system or an air -handling system often denotes a more specific system that conditions transports, distributes the supply air , recirculates the space air, and controls the indoor en vironment according to requirements. Such an air system often consists of fans, heat exchangers (including coils, evaporative coolers, and dehumidifiers) a direct-fire gas heater, f lters, a mixing box, dampers, ductwork, terminals, space diffusion devices, and controls. In a central hydronic air conditioning system, an air system consists of an air -handling unit (air is cooled and dehumidified by a ater cooling coil), ducts, terminals, space diffusion devices, and controls. In a unitary packaged air conditioning system, the air system consists of an air handler (in which air is cooled and dehumidified by a DX coil or rotary dehumidifier ducts, terminals, space diffusion devices, and controls. In a space air conditioning system, its air system is a combination of an outdoor air v entilation system and a space air recirculating system, which consists of a f an, a coil, a filte , ductwork, and controls. A separate outdoor ventilation air system provides conditioned

AIR SYSTEMS: BASICS AND CONSTANT-VOLUME SYSTEMS

20.3

outdoor air for occupants. A space recirculating system conditions the space recirculating air to of fset the space cooling or heating load. In an indi vidual air conditioning system, the air system, heating system, and refrigeration systems are inte grated in a single packaged air conditioner . It is not necessary to distinguish them separately . The air system in a thermal storage, desiccant cooling, or an evaporative-cooling air conditioning system is similar to that of a central hydronic system. The air system is the foundation in an air conditioning system. It directly contacts the occupants, manufacturing products, and conditioned space. Air systems play a k ey role in maintaining a healthy, safe, and comfortable indoor environment with an acceptable indoor air quality at optimum energy use. It is more specific convenient, and clear that the systems which condition, transport, distribute, and control the supply air and recirculate the space air are called air systems; and all the other air systems in a broad sense that do not cool and dehumidify the air are called mechanical v entilation systems, exhaust systems, and smoke control systems.

Air Distribution Systems An air distrib ution system is a subsystem of an air system which consists of ductw ork, terminals, space diffusion devices, and controls to transport and distrib ute the conditioned air to the occupied zone as well as to control the indoor environment within required limits.

Ventilation Systems As discussed in Sec. 4.10, ventilation air comprises outdoor air plus treated recirculated air with acceptable indoor air quality. Ventilation air is also the air used for the dilution of the contaminant concentrations in the conditioned space. Ventilation air may or may not be conditioned. Ventilation indicates the process to supply or to remo ve ventilation air. Because of the importance of the v entilation air supply , a ventilation system is usually considered as part of air systems that use f ans to supply and to remo ve ventilation air to and from an y space. Natural v entilation (including infiltra tion) is a process that provides outdoor ventilation air by stack or wind effect, or both.

Mechanical Ventilation Systems A mechanical ventilation system uses mechanical means (fans) to transport the air without cooling and dehumidifying it. The follo wing are the types of mechanical v entilation systems currently used: ●

●

●

Mechanical ventilation systems for cooling and contaminant dilution, which supply outdoor air to a room to cool equipment or to dilute air contamination to acceptable le vel, as in a transformer room or a garage Exhaust systems, which discharge indoor air, or exhaust toxic gas and contaminated air from localized sources such as e xhaust systems from bathrooms or from chemical hoods in laboratories Makeup air systems as a part of mechanical v entilation systems in which outdoor air is often heated and filtered to compensate for the xhausted air

Regenerative Systems In a re generative or drying system, outdoor air is heated to a higher temperature such as 150 to 200°F (65 to 93°C) and is then used to remo ve the absorbed moisture from the desiccant by drying. After drying, the desiccant is reactivated and is ready to dehumidify the process air again.

20.4

CHAPTER TWENTY

Smoke Control Systems Smoke control systems include stairwell pressurizing systems and zone smok e control systems. Stairwell pressurizing systems pressurize stairwells as an escape route in high-rise b uildings during building fire. Zone smo e control systems e xhaust the smok e in the fired floor and pressurize t floors immediately ab ve and below the fired floor to p vent smoke contamination. Smoke control systems are discussed in Chap. 22.

Terminals A terminal is a factory-made device installed inside a control zone either in a ceiling plenum abo ve the conditioned space or sometimes directly mounted adjacent to the e xternal w all of the conditioned space or in a floor plenum. A terminal usually has an individual control system to maintain a selected indoor enviromental parameter within predetermined limits for a control zone. Conditioned air including outdoor air is supplied to a terminal from the supply main duct through branch takeoffs. It may be mix ed with recirculating air and is supplied to the control zone served by the terminal in a multizone air system. A terminal may perform any of the following functions or combinations thereof: ●

●

●

Heating / cooling or filtratio Mixing of outdoor v entilation air or conditioned air with recirculating air , or mixing of cold supply air with warm supply air Controlling and maintaining a predetermined zone temperature by modulating the v olume fl w rate of supply air, the temperature of supply air, or the ratio of the v olume fl w of cold and w arm air supplies. VAV boxes, reheating VAV boxes, mixing VAV boxes, fan-powered VAV boxes, and fan-coil units are all terminals, and will be discussed in Chap. 21.

Primary, Secondary, and Transfer Air Primary air is conditioned mak eup outdoor air or a conditioned mixture of outdoor and recirculating air. Primary air is either conditioned and supplied from a separate mak eup AHU (PU) or from an AHU (PU) using a mixture of outdoor and recirculating air to the terminals. Sometimes the primary air is supplied directly to the space. Primary air is usually mix ed with the recirculating air , which is often called secondary air . The mixing most often tak es place in the terminals, or another AHU or PU, or sometimes directly in the conditioned space. The mixture of primary air and recirculating air (secondary air) may be conditioned again in the terminals. It is then supplied to the conditioned space to of fset the space load. The ratio of the v olume fl w rate of primary air to the mixture (zone supply air) v aries from 0.15 to 0.5. Recirculating air is the air recirculated from the conditioned space, through an AHU, a PU, or a terminal, and supplied again to the conditioned space. Recirculating air is dif ferent from the return air which returns from the conditioned space to an AHU, or PU, and exhausts outdoors. The purpose of mixing primary air with recirculating air in the terminal is to sa ve energy including fan energy, as well as to sa ve investment by using smaller supply and return ducts, which occup y less space. Transfer air is indoor air that mo ves or supplies to a conditioned space from an adjacent area. Transfer air is entirely composed of recirculating air (secondary air). Transfer air often has less unused outdoor air than a mixture of outdoor and recirculating air . Unused outdoor air is outdoor air whose capability of dilution of air contaminants has not been e xpended. Recirculating air still may contain a certain amount of unused outdoor air . Restrooms are normally supplied by transfer air.

AIR SYSTEMS: BASICS AND CONSTANT-VOLUME SYSTEMS

20.5

20.2 BUILDING LEAKAGE AREA AND BUILDING TIGHTNESS Currently, the b uilding tightness or the b uilding air leakage can be indicated by the follo wing indices: (1) an ef fective leakage area, (2) air leakage per unit area of b uilding shell, and (3) air change per hour (ACH) at a specific indoo -outdoor pressure differential.

Effective Leakage Area The effective leakage area of a b uilding Ae,l, in ft 2 (m2), is the amount of open w all area that allows the same volume fl w rate of air fl wing through it as fl wing through the actual building when the pressure difference across the open w all area and the shell of the actual b uilding is the same. The Ae,l value is usually determined by a f an pressurization method, in which a blo wer installed at the door of a b uilding is used to maintain a pressure dif ference across the b uilding shell. The airfl w rate V˙ , in cfm (m3), can then be measured , and the Ae,l value is calculated as

√

V˙  Cd Ae,1

2 p  Cd Ae,l 4005√p 

(20.1)

where Cd  discharge coefficient; its alue is 1 for effective leakage area   air density, lb / ft3 (kg / m3) p  pressure differential between outdoor and indoor air at same level, in. WC (Pa) According to ASHRAE Handbook 1997 , Fundamentals, effective air leakage areas for representative residential b uilding components at an outdoor -indoor pressure dif ference of 0.016 in. WC (4 Pa) are as follows:

Best estimate, in.2 /ft2 External walls Cast-in-place concrete Clay brick cavity wall, f nished Lightweight concrete block, painted Window Wood frame Uncaulked Caulked Double-hung, not weatherstripped Double-hung, weatherstripped Door, average Ceiling

0.007 0.0098 0.016 0.025 0.004 0.12 0.031 0.015 in.2 / lftc (linear ft. crack) 0.026

Building Air Leakage When air fl ws through a crack or gap, the fl w never becomes fully de veloped. The relationship between overall air leakage rate V˙ leak, in cfm (L / s), and the indoor-outdoor pressure difference p, in in. WC (Pa), can often be represented by a widely used air leakage equation as V˙leak  Cflow Aw (p)n

(20.2)

20.6

CHAPTER TWENTY

where Cf ow  f ow coeff cient, cfm / ft2 (m3 / h  m2) at p  in. WC (Pa) Aw  area of exterior wall or roof, ft2 (m2) p  pressure difference across building envelope, in. WC (Pa) Because the cracks ha ve lar ge f ow resistance, in Eq. (20.2), the e xponent n may ha ve a v alue closer to 1 than openings with less f ow resistance. Sha w et al. (1993) recommended n  0.65 based on f eld-measured results. Persily (1999) found signif cant levels of air leakage in commercial and institutional b uildings. Among the data measured from 139 b uildings around the w orld, f ow coeff cients Cf ow, in cfm / ft2 (m3 / h  m2), varies from 0.15 to 6.85 cfm / ft2 (2.7 to 124.5 m3 / h  m2) at an indoor-outdoor pressure dif ference p of 0.3 in. WC (75 P a) using the f an pressurization test method with a mean Cf ow  1.5 cfm / ft2 (27.1 m 3 / h  m2) of exterior wall area at p  0.3 in. WC (75 Pa). There is no correlation between air leakage and the age of the b uilding tested. Building height does sho w an impact on airtightness with high-rise b uildings being tighter , especially onestory buildings are leakier. Based on Persily (1999) who analyzed b uilding envelope airtightness data measured from 139 commercial and institutional b uildings, f ow coef f cients Cf ow, in cfm / ft2 (m3 / h  m2), per unit exterior wall area at 0.3 in. WC (75 Pa) for tight, average, and leaky buildings can be estimated as follows: Cf ow, cfm / ft2 (m3 / h  m2) of exterior wall area at 0.3 in. WC (75 Pa) Tight building Average Leaky building

0.30 1.00 2.00

Air Change per Hour at 50 Pa (ACH50) The air leakage rate of a b uilding can also be e xpressed by the air changes per hour at a speci f c oudoor-indoor pressure difference of which 0.2 in. (50 P a) is most widely used. As in the ef fective leakage area, a fan is installed at the door of a b uilding and is used to maintain a speci f c pressure difference across the b uilding shell. The air leakage V˙ leak, in cfm (L / min), supplied by the f an is squeezed outdoors through the cracks and gaps on the b uilding shell and is then measured. The air change per hour of air leakage, denoted by ACH, can be calculated as ACH 

60 V˙leak Vr

(20.3)

where Vr  the space volume of room or enclosure, ft3 (m3). For residential buildings, in ASHRAE Handbook 1997, Fundamentals, ACH50  3 is considered e xceptionally tight, and for residential buildings 5  ACH50  20 is considered con ventional. There are 10 percent of the 204 lo wincome houses whose ACH50  47.5. Withers et al. (1996) measured the ACH50 values for 18 small commercial buildings after repair of duct leaks, the sealing of the b uilding shell, and other repairs, four of them with ACH50  5.2, seven of them with 10.7  ACH50  19.7, f ve of them with 21.2  ACH50  28.8, one with ACH50  46, and the highest one with ACH50  52.8, a court off ce. Because of the desire for ener gy conservation, the airtightness of b uildings has been impro ved during the past decades. Since the ratio of the e xterior wall area to the f oor area is quite dif ferent for different types of b uildings, it is diff cult to compare the airtighness of dif ferent types of b uildings, especially using different air leakage indices.

AIR SYSTEMS: BASICS AND CONSTANT-VOLUME SYSTEMS

20.7

20.3 SPACE PRESSURIZATION Space Pressure Characteristics Space pressurization or b uilding pressurization is a process that maintains a static pressure dif ference between space air and the air in the surrounding area, either outdoor air at atmospheric pressure or air in adjacent rooms, for the purpose of providing a required indoor environmental control. As mentioned in Chap. 4, normally, slightly positive space pressure should be maintained if the required le vel of indoor air quality and en vironmental control of the conditioned space is higher than that of the surrounding area. In laboratories with toxic gas e xhaust systems, negative space pressure should be maintained. As co vered in Chap. 4, for comfort air conditioning systems for low-rise b uildings, the pressure dif ferential between space air pr and outdoor air po is usually between 0.005 and 0.05 in. WC ( 1.25 and 12.5 P a), and 0.01 to 0.03 in. WC (5 to 7.5 Pa) is often adopted. F or high-rise buildings, the pressure differences between pr and po are far higher because of the stack ef fect on various f oors. For clean rooms, the minimum pressure differential between a clean room and an y adjacent area of less strict cleanliness requirements should be 0.05 in. WC (12.5 Pa) with all entryways closed. A space positve pressure of 0.05 in. WC (12.5 Pa) exerts a force of 4 lb f (18 N) on a door with an area of 16 ft 2 (1.5 m2). An excessive pressure differential may make opening and closing doors diff cult. A positive pressure means that pr  po or pr  psur. Here psur denotes static pressure in the surrounding area, in in. WG (Pa). When a positive pressure is maintained in a conditioned space, in addition to the air returned and e xhausted from the space, a certain amount of air is squeezed out or exf ltrated from the space through openings and cracks in the building shell. Unconditioned outdoor air or contaminated air from the surrounding area is pre vented from in f ltrating the conditioned space because of this positi ve space pressure. A negative pressure means that pr  po or pr  psur, which induces inf ltration. Except in air jets, the velocity pressure in the conditioned space is small and is usually ignored in space pressurization analysis. The space pressure dif ference between the conditioned space and the surrounding area is affected by the stack effect, wind effect, and operation of the air systems and e xhaust systems as well as the airtightness (amount of leakage area) of the building shell.

Stack Effect The temperature difference between the cold outdoor and warm indoor air columns causes a density difference between these air columns that, in turn, creates a pressure dif ference between the cold and the warm air columns. This phenomenon is called the stack effect. In an y b uilding under the stack ef fect alone, there is a le vel at which the indoor and outdoor pressures are equal. This level is called the neutral pressure le vel (NPL), as shown in Fig. 20.1 b. During the heating season, outdoor cold air pressure is greater than indoor w arm air pressure below the NPL. This pressure dif ference causes the outdoor cold air to enter the b uilding through the lower inlets, and the indoor w arm air to f ow upward and discharge from the top openings. During the cooling season, the reverse effect occurs, but the stack ef fect is smaller because of the smaller temperature difference between outdoor and indoor air. The pressure dif ference due to the stack ef fect between outdoor and indoor air pst, in in. WC (Pa, 1 in. WC  248.6 Pa) at height H (in ft or m), from a reference level, is given as pst  pso psi  0.19256(o i)(H HNPL)  0.19256 i

(H HNPL )(TRi TRo) TRo

(20.4)

20.8

CHAPTER TWENTY

Building

Height ft

Exhaust duct for restrooms Elevator shaft

p between space air and air in exhaust duct (exhaust fan is shut off )

(c)

Building Outside Outside p in WC

(a)

Outside

Elevator shaft

Neutral pressure level (NPL)

Elevator shaft

pel po

Elevator Outside door door (b)

Building

(d) FIGURE 20.1 Stack effect in a high-rise b uilding (a) schematic diagram of b uilding airf ow; (b) pressure pro f les; (c) pressure pro f les for tight building shell; (d ) pressure prof les for leaky building shell.

where pso , psi  outdoor and indoor absolute static pressure, in. WG (Pa) o, i  outdoor and indoor air density, lb / ft3 (kg / m3) HNPL  vertical distance between NPL and reference level, ft (m) TRo,TRi  outdoor and indoor absolute temperature, °R For a high-rise b uilding with tw o openings on the e xternal w all, NNPL, in ft (m), measured from lower opening can be calculated as HNPL 

Ho 1 (A1/A2)2(TRi / TRo)

(20.5)

where Ho  vertical distance between two openings, ft (m) A1, A2  area of lower and higher openings, ft2 (m2) TRi, TRo  absolute temperature of indoor and outdoor air, °R (K) In Eq. (20.5), TRi  TRo. If the indoor air is cooler , that is, TRo  TRi, then TRi / TRo should be changed to TRo / TRi. For conditioned spaces where TRi  TRo, the chimney and exhaust systems raise the NPL level and the cold outdoor air supply system lowers the NPL level. In Fig. 20.1 b, it can be seen that the absolute static pressure of outdoor air pso is greater at the low level than at the high level. This is because the air column is taller at the lower level. In an ideal single-cell building with no inside partitions, with a vertical distance of 100 ft (30 m) between the lo wer air inlet and upper air outlets and an indoor and outdoor air temperature dif ference of 50°F (10°C), the static pressure difference pst is about 0.15 in. WC (38 Pa). Most buildings have inside partitions; i.e., they are multicell b uildings. They have doors at the lower entering inlet, windows at the upper dischar ge outlets, and doors to the stairwells. Usually

AIR SYSTEMS: BASICS AND CONSTANT-VOLUME SYSTEMS

20.9

doors and windo ws are closed. When outdoor cold air f ows through the cracks and gaps around these doors, windows, and partitions, its velocity and f ow rate are signi f cantly reduced because of these dynamic losses. Stack effect depends on the magnitude of outdoor -indoor temperature dif ference and therefore varies from maximum to zero as the outdoor climate changes.

Stack Effect for High-Rise Buildings Lovatt and Wilson (1994) measured the pressure differences of the various elements of the building shell and internal partitions of a 44-story high-rise of f ce building of 577 ft (176 m) abo ve grade located in Montreal, Quebec, Canada. The air leakage rate for this b uilding was found by Tamura and Wilson (1967) to be 0.55 cfm / ft2 (2.8 L / s  m2). The stairwells in this b uilding are not directly connected to a stairwell pressurization system. The results were as follows: Condition Outdoor temperature Lobby temperature Pressure difference p, in. WC (Pa) Outside – lobby vestibule Lobby vestibule – lobby Low-rise elevators – lobby Mid-rise elevators – lobby High-rise elevators – lobby M2 stairwell – lobby Roof elevator – 41st f oor Roof elevator – roof vestibule Observation deck – roof vestibule Roof – stairwell 43d Floor – stairwell 42d Floor vestibule – stairwell Exhaust fan room – 42d f oor vestibule 42d Floor – 42d f oor vestibule High-rise elevator mech. room – 43d f oor Mid-rise elevator mech. room – 32d f oor Low-rise elevator mech. room – 16th f oor

Normal occupied 20.3°F ( 6.5°C) 73.0°F (22.8°C) 0.29 (72) 0.02 (5)

0.10 ( 24)

0.12 ( 30)

0.16 ( 40)

0.008 ( 2)

0.23 ( 56)

0.32 (80)

0.07 ( 18)

0.78 ( 195)

0.19 ( 47)

0.07 ( 17) 0.06 (16)

0.17( 43) 0.21 (53) 0.02 (5) 0 (0)

The stack effect for high-rise buildings in cold climates may cause the following problems: ●

●

●

●

●

Pressure differences across elevator doors pelev, d generally do not exceed the 0.1-in. WG (25-Pa) limit. This is the limit recommended by the manuf acturer for proper door operation. Sticky elevator doors may occur when pelev, d  0.1 in. WC (25 Pa). Pressure difference across the stairwell doors psta,w is generally less than 0.04 in. WC (10 P a). These stairwell doors to occupied f oors are easy to open and pro vide an escape route to outdoors during a building f re. During cold weather , the pressure dif ference across the roof door proof may well be abo ve the NFPA limit (1988) between 0.21 and 0.45 in. WC (52.5 and 112.5 Pa). The lobby entrance doors show greater pressure difference across them. The stack effect may affect the airf ow distribution of a vertical exhaust duct from restrooms.

For a tight high-rise b uilding with less air leakage area on the b uilding shell, the pressure prof le of the high-rise building is shown in Fig. 20.1c, where the pressure difference across the outside (entrance) door is greater than the pressure dif ference across the ele vator doors. F or a leak y high-rise

20.10

CHAPTER TWENTY

building, the pressure pro f le is shown in Fig. 20.1 d; the pressure dif ference across the outside door is smaller than the pressure difference across the elevator doors. Lovatt and Wilson (1994) recommended that if the pressure dif ference of the elevator doors of a high-rise building during cold climate pelev,d  0.1 in. WC (25 P a), then repair of the air leakage area on the building shell is often the f rst remedy. Reduction of air leakage will reduce the in f ltration and the heat energy use as well. Add airtight v estibules around all entrance doors, and around doors serving ele vators and stairwells. These setups di vide the pressure dif ference of the entrance door , elevator doors, and stairwell doors into two parts, and again they reduce inf ltration. Add airtight separations in vertical shafts whenever possible, to reduce the pressure difference across the roof door in the stairwell. To pressurize or depressurize the entrance lobby or an y other f oor by a mechanical v entilation system is often not an ef fective and economical means to reduce the pressure dif ference across doors or other b uilding elements, because one of the pressure dif ferences is reduced at the e xpense of increasing the other one.

Wind Effect As wind f ows over a building, it creates positive pressure on the windward side of the building and negative pressure on the leeward side, as shown in Fig. 20.2. Wind pressure on the building surface relative to the approaching wind speed pw, in in. WC (Pa), can be calculated as pw  CpPvw  Cp

vwa

2

 4005 

(20.6)

where Cp  surface pressure coeff cient pvw  velocity pressure at approaching wind speed, in. WC (Pa) vwa  approaching wind speed at wall height on windward side, fpm (m / s) If the wind direction is normal to the windw ard surface of a high-rise b uilding with a width-depth ratio of 4, the average Cp value is about 0.60 on the windward side, about 0.5 on the leeward side and on the f at roof, and about 0.25 on the other two sides.

FIGURE 20.2 Airf ow around buildings.

AIR SYSTEMS: BASICS AND CONSTANT-VOLUME SYSTEMS

20.11

Wind speed from a meteorological station vmet, in fpm (m / s), is usually measured at a height of 33 ft (10 m). The approaching wind speed at w all height vwa should be corrected for height and terrain roughness as follows: Hw

n

H 

vwa  Aovmet

(20.7)

met

where Hw  height of wall where wind pressure is exerted, ft (m) Hmet  height at which wind speed is measured at meteorological station, ft (m) In Eq. (20.7), Ao is a correction f actor for terrain roughness, and n is the velocity prof le exponent. For suburban areas, Ao  0.6 and n  0.28; for urban areas, Ao  0.35 and n  0.4; for airports, Ao  1.0 and n  0.15. The mean wind speeds measured at meteorological station vmet in most U.S. cities in winter are below 1200 fpm (6 m / s). Wind pressure on a wall at a height of 33 ft (10 m) in suburban areas with vmet  1200 fpm (6 m / s) is about 0.054 in. WC (13 Pa). The inf uence of the wind effect on space and building pressure characteristics is as follows: 1. Rooms on the windw ard side of the b uilding are usually at a positi ve pressure and on the leeward side at a ne gative pressure relati ve to the corridor pressure. It is best to b uild clean spaces, such as conference rooms, on the windward side of the prevailing wind, and laboratories with toxic gas exhaust systems on the leeward side of the building. 2. Outdoor air intak e should be located on the side with a positi ve surface pressure coef f cient Cp in the prevailing wind. Exhaust outlets should be located where Cp is negative, preferably on the rooftop. 3. Suff cient total pressure must be provided by the supply fan to overcome the negative pressure at the outdoor intake and by the exhaust fan to overcome the positive pressure at the exhaust outlet. Alternatives should be provided to allow outdoor air intake and outlet when the wind direction is different from that of the prevailing wind. Wind speed may vary from zero to its maximum speed. Wind velocity is a vector quantity.

Air Systems and Mechanical Ventilation Systems If a space is served by air systems including supply, return, and exhaust systems, the space pressure characteristics are closely related to the characteristics of air systems during their operating time. When a hurricane comes, the windward pressure may be as high as several inches WC (several hundred Pa), and the leeward negative pressure has a similar ne gative pressure. Hurricanes and storms last a comparati vely short time. Hurricanes af fect entrance doors, windows, as well as the performance of outdoor intak e and e xhaust outlets in the air systems and HVAC&R equipment outdoors that are not well shielded and protected. For high-rise buildings in cold winter , the stack effect has a signi f cant inf uence on the indoor outdoor pressure dif ference on v arious f oors. F or lo w-rise b uildings, air systems (including mechanical ventilation systems) have a dominant inf uence on space pressure characteristics e xcept on stormy days.

Airflow Balance and Space Pressurization by Differential Flow Consider a single-cell room model of an enclosed space surrounded by an area of uniform pressure. Based on the principle of continuity of mass, for an incompressible steady f ow process, the total mass f ow rate of air entering an enclosed space must equal the total mass f ow rate of air lea ving

20.12

CHAPTER TWENTY

the space (see Fig. 20.3), or m˙ s m˙ inf  m˙ rt m˙ ex where m˙ s  mass f ow rate of supply air, lb / min (kg / s) m˙ rt  mass f ow rate of return air, lb / min (kg / s) m˙ ex  mass f ow rate of exhaust air, lb / min (kg / s)

FIGURE 20.3 Mass f ow balance of an enclosed space: (a) mass f ow balance; (b) f ow circuit.

(20.8)

AIR SYSTEMS: BASICS AND CONSTANT-VOLUME SYSTEMS

20.13

In Eq. (20.8), m˙ inf indicates the mass f ow rate of in f ltration air, lb / min (kg / s). The term m˙ inf means the mass f ow rate of e xf ltration. Because m˙  V˙ , for simplicity, if the dif ference in air density is ignored, then V˙s V˙inf  V˙rt V˙ex

(20.9)

where V˙ s, V˙ inf, V˙ rt, V˙ ex  volume f ow rate of supply, inf ltration or exf ltration, return, and exhaust air, cfm (L / s). For convenience, if the stack effect and wind effect are zero, and the pressure of the surrounded area or outside pressure po is equal to the atmospheric pressure and is assumed to be zero, that is, po  pat  0, then the pressure of the enclosed space is represented by pi, in in. WG (Pa). Let po,st  po,st pat be the increase of the outside pressure due to the stack ef fect, and let pow  pow pat be the increase of the outside pressure due to the wind effect; both are in in. WC (Pa). Space pressurization by using dif ferential f ow control is the most widely used method today . When po,st and pow both are zero and the difference in air density is ignored, if an enclosed space is pressurized or depressurized because of dif ferential f ow of air (supply and return/e xhaust airf ow) systems, the following relationships must hold:

V˙inf  V˙s (V˙rt V˙ex)

(20.10 )

From Eq. (20.1)

V˙inf  4005Cd Ae,l √( pim po) and ( pim po) 

V˙inf

 4005C A  d

(20.11)

e,l

From Eq. (20.10), if V˙ s  (V˙ rt V˙ ex), then the enclosed space pressure is pressurized from atmospheric pressure po up to pim, in in. WG (Pa), and V˙ inf is negative, that is, V˙ inf, an exf ltration. If V˙ s  V˙ rt V˙ ex, pim  po, then V˙ inf is positive, an inf ltration. From Eq. (20.11), space pressurization (pim po) or space depressurization (pim po) depends on the amount of differential f ow V˙ s (V˙ rt V˙ ex)  V˙ inf that has been inf ltrated or exf ltrated through the b uilding shell and the air leakage area Ae,l (airtightness of the b uilding). The greater the V˙ inf and the smaller the Ae,l , the higher will be the pressurization. Similarly, the greater the V˙ inf and the smaller the Ae,l , the higher the depressurization. Cummings et al. (1996) made f eld measurements of uncontrolled air f ow and depressurization of eight restaurants. From the measured results, the average building pressures ( pim po) of f ve restaurants was greater than or equal to 0.032 in. WC ( 0.006, 0.003, 0.024, 0.032, and

0.018 in. WC). These are the results of the following two restaurants:

Chicken 1 Chicken 2

Floor area, ft

ACH50

Ae,l, ft2

OA & MA, cfm

Exhaust, cfm

Building pressure, in. WC (Pa)

3,161 3,321

14.81 7.70

3.90 2.16

8,410 6,360

10,606 9,272

0.032 (7.6)

0.172 (43)

Here OA represents outdoor air and MA the makeup air. The building pressure of the restaurant had a smaller Ae,l  2.16 ft 2 (0.242 m 2), and a greater dif ferential of air f ow between e xhaust f ow and OA and MA f ow was depressurized further ( 0.172 in. WC or 43 Pa). Coogan (1996) recommended a space pressurization by dif ferential airf ow control using a more complicated room model surrounded by tw o zones, which is useful when the enclosed space is adjacent to two spaces of different pressures.

20.14

CHAPTER TWENTY

20.4 INFILTRATION AND EXFILTRATION Inf ltration is the uncontrolled inw ard f ow of outdoor air through cracks and openings in the b uilding shell when the pressure of the outdoor air is higher than the pressure of the indoor air at the same level. Exf ltration is the uncontrolled outw ard f ow of air through cracks and openings in the building shell when the pressure of indoor air is higher than that of outdoor air . Exf ltration is negative inf ltration. Inf ltration and e xf ltration can be induced by the stack ef fect, wind effect, air systems or mechanical ventilating systems, or a combination thereof. If atmospheric pressure is pat zero, the outdoor pressure po is the sum of the stack ef fect po,st and wind effect pow, both in in. WC (Pa), and for the building shell at a normal direction to the approaching wind, outdoor pressure po, in in. WC (Pa), can be calculated as po  po,st pow pat

(20.12 )

If both po,st and pow are zero and the space pressure is not pressurized or depressurized, then the indoor pressure p  pat  0. If the space is pressurized by the supply or e xhaust system, pi  pim, both in in. WC (Pa).

Volume Flow Rate of Infiltration The volume f ow rate of inf ltration V˙ inf, in cfm (m3 /s), for a residential building can often be calculated by Eq. (20.1) using the ef fective leakage area Ae,l. For low-rise commercial buildings in summer, as discussed in Chap. 6, the stack effect is small; if windo ws are well sealed and the space is pressurized with a positive pressure, the inf ltration can be ignored. For low-rise buildings in winter, according to f eld measurements, the follo wing in f ltration rates, in A CH, may be considered in space heating load calculations for rooms with windo ws on one side of a commercial b uilding that are not well sealed: Winter mean wind speed vmet, fpm (m / s)

Inf ltration rate, ACH

1000 (5) 1200 (6)

0.2 0.3

For high-rise b uildings in both summer and winter design conditions, the in f ltration induced by stack and wind effect should be considered. First, the outdoor pressure due to stack effect po,st and wind ef fect pow for each f oor can be calculated by Eqs. (20.4) and (20.6), and the outdoor air pressure po can be calculated by Eq. (20.12). As the pressurized indoor pressure is pim, the pressure differece across the building shell p  po pim. Second, from the building construction data, f nd the area of the exterior wall, in ft2 (m2). Third, calculate the volume f ow rate of the inf ltration V˙ inf, in cfm (m3 /s), for each f oor by the air leakage equation, Eq. (20.2). According to the airtightness of the building, select the f ow coeff cient Cf ow as listed previously.

20.5 FAN-DUCT SYSTEMS A fan-duct system is an idealized air system in which a f an is connected to ductw ork or equipment (a f ow resistance), or se veral f ans are connected to man y sections of ductw ork and equipment. More precisely, a fan-duct system is a combination of fan and f ow resistance. Fan-duct systems are used to study and analyze the operating characteristics of air systems.

AIR SYSTEMS: BASICS AND CONSTANT-VOLUME SYSTEMS

20.15

FIGURE 20.4 System operating point.

System Operating Point A system operating point indicates the operating condition and characteristics of a f an-duct system. Because the operating point of the f an must lie on the f an performance curv e and the operating point of the duct system must lie on the system curv e, the system operating point of a f an-duct system must be the intersection point P of the fan performance curve and the system curv e, as shown in Fig. 20.4. At the operating point of the f an-duct system, its volume f ow rate is V˙ P, and its total pressure loss is pP. If the f ow resistance of the duct system remains the same while f an performance changes, and its characteristics are represented by a ne w fan performance curv e, the operating point of the f anduct system moves to point Q in Fig. 20.4. On the other hand, if the fan performance curve remains the same while the f ow resistance of the duct system is v aried and represented by a ne w system curve, the operating point of the fan-duct system moves from P to S. The fan in a constant-v olume air system should be selected with a V˙ and pt so that the system operating point P is near the point of maximum fan total eff ciency t. The actual operating point of a constant-volume system is usually within a range of 10 percent of the wide-open v olume from the point of maximum t. Buckingham  Method and Fan Laws The relationship between v arious interdependent variables that represent f an characteristics for dynamically similar systems can be determined by using the Buckingham method. For two systems

20.16

CHAPTER TWENTY

to be dynamically similar , they must be similar in shape and construction, i.e., geometrically similar. Their velocity distribution or pro f le of f uid f ow should also be similar . The Buckingham method gives the relationship as

 FD an bc

(20.13)

where F  dependent variable such as V˙ , pt , or power D  diameter of impeller, ft (m) n  fan speed, rpm   air density, lb / ft3 (kg / m3) In Eq. (20.13), a, b, c are indices to be determined so that the group as a whole is dimensionless. Variables corresponding to fan-duct systems and their dimensions in terms of mass M, length L, and time t are as follows: Indices Variable

Unit

M

L

t

V˙ Volume f ow cfm Total pressure pt Power Pf Impeller diameter D Speed n Air density 

(m3 /s) lbf /ft2 (N / m2) ft lbf /min (kJ / s) ft (m) rpm lbm /ft3 (kg / m3)

0 1 1 0 0 1

3

1 2 1 0

3

1

2

3 0 1 0

For the dimensionless group 1 including dependent variable volume f ow V˙ and independent variables D, n, and , their relationship can be found by determining the indices as follows:

1  V˙D an bc [0] [L3t 1 ][L]a[t 1 ]b[ML 3]c Setting up equations in M, L, and t, we have M

0c

L

0  3 a 3c

a  3

t

0 1 b

b  1

That is,

1  V˙D 3n 1 

V˙ D 3n

Or, V˙ / (D3n)  constant, then V˙  K v D 3n

(20.14)

where Kv  volume f ow constant. Similarly, pt  K pD 2n 2 Pf  K PD 5n 3 where Kp  pressure constant and KP  power constant. Equations (20.14) and (20.15) can be used to determine the performance of a f under the following two conditions:

(20.15)

an-duct system

AIR SYSTEMS: BASICS AND CONSTANT-VOLUME SYSTEMS

●

Condition 1. For the same f an-duct system operating at dif ferent speeds D1  D2, if the difference in air density is ne gligible, the v olume f ow rate ratio is proportional to the speed ratio, or K v D 32n 2 V˙2 n2   3 ˙ n1 K v D 1n 1 V1 In Eq. (20.16), subscripts 1 and 2 indicate the original and changed operating conditions, shown in Fig. 20.4. The total pressure rise ratio is equal to the square of the speed ratio K p D 22n 22 2 pt 2 n 22   2 pt1 n1 K p D 21 n 21 1 The fan power ratio is equal to the cube of the speed ratio

●

20.17

(20.16) as

(20.17)

K PD 52n 322 P2 n 32 (20.18)   3 5 3 P1 n 1 K PD 1n 11 Condition 2. For fan-duct systems with geometrically and dynamically similar f ans and the same duct system, i.e., the same type of f ans ha ving the same Kv, Kp, and KP, operating at high Reynolds numbers and installed in the same duct system, the volume f ow ratio, total pressure rise ratio, and power ratio can be expressed as K v D 32 n 2 V˙2 D 32 n 2   K v D 31n 1 D 3 V˙1 1 n1 K p D 22n 222 pt2 D 22n 22   2 2 pt1 D 1n 1 K p D 21n 211

(20.19)

K p D 52n 322 p2 D 52n 32   5 3 p1 D 1n 1 K p D 51n 311 Equations (20.16) through (20.19) relate the parameters of the f an-duct system at v arious operating conditions. They are often called the fan laws.

20.6 SYSTEM EFFECT Mechanism of System Effect The system ef fect of a f an-duct system describes the loss of f an performance caused by une ven and nonuniform velocity prof les at the inlet entering and after the f an outlet using actual operating inlet and outlet connections, compared to the performance of that f an test unit during laboratory ratings. Individual f ans are tested based on ANSI/ASHRAE Standard 51-1985 and ANSI/AMCA Standard 210-85. The con f guration of the test unit is sho wn in Fig. 20.5 b. There are no inlet connections. There is also a minimum length of 10 duct diameters of straight duct connected to the fan outlet. The system effect of a fan-duct system consists of inlet system effect and outlet system effect.

20.18

CHAPTER TWENTY

Inlet System Effect Consider a 90 ° elbow that is connected to the f an inlet with a v ery short connecting duct between the elbow and the f an inlet. Eddies, large-scale turbulence, and a nonuniform v elocity prof le form after the elbo w and are not smoothed to a uniform v elocity distribution because of the short connecting duct. Air enters the f an with a nonuniform v elocity prof le that is quite dif ferent from the

FIGURE 20.5 Inlet system effect: (a) inlet system effect on the pt V˙ diagram; (b) fan test unit at laboratory; (c) round elbow before fan inlet.

AIR SYSTEMS: BASICS AND CONSTANT-VOLUME SYSTEMS

20.19

fan inlet conditions in the standard laboratory test unit. This results in a fan performance curve with lower f an total pressure and v olume f ow than the catalog f an performance curv e, as sho wn in Fig. 20.5a. The difference between the fan total pressure in the f an-duct system with a nonuniform v elocity prof le at fan inlet and the fan total pressure during standard rating at the laboratory is called the inlet system effect. Fans must be selected at a f an total pressure that compensates for the inlet system effect, so that the actual f an performance curve can meet the required f an total pressure at designed volume f ow at operating point Pd, as shown in Fig. 20.5a. If a fan-duct system does not consider the pressure loss caused by the inlet system ef fect, it will actually operate at point Pa with a lower volume f ow and fan total pressure than the calculated v alues, even though the calculated design operating point is at Pd.

Inlet System Effect Loss Inlet system ef fect loss ps,i, in in. WC (Pa), is the dif ference in pressure loss between actual and catalog performance curv es caused by the inlet system ef fect, and it can be calculated as dynamic loss, as def ned in Eq. (17.52) ps,i  Cs,i

vfi

2

 4005 

(20.20)

where Cs,i  inlet system effect loss coeff cient. In Eq. (20.20), vf is the air velocity at the fan inlet based on the area of the inlet collar, fpm (m / s). Inlet system ef fect loss should be added to the system total pressure loss calculated at the designed volume f ow rate, as described in Eq. (17.23). The selected catalog f an performance curv e should meet the system total pressure loss, including the inlet system effect loss.

Inlet System Effect Loss Coefficient According to the Air Movement and Control Association (AMCA) Fan and Systems, Publication 201 (1990), if an elbow is connected by a round duct to the inlet of a centrifugal f an and the length of the connecting duct is equal to 2 duct diameters, the inlet system ef fect loss coeff cients are as follo ws: R / D of elbow Cs,i

0.75 0.8

1.0 0.66

2.0 0.53

Here R is the radius of the elbo w, and D is the duct diameter , both in ft (m). If the length of the connecting duct is 5 D, then Cs,i is approximately one-half the v alues listed abo ve. If there is no connecting duct, then Cs,i is about twice the values listed above. According to AMCA (1973), for a square elbo w (with or without turning v anes) with an inlet transition and a connecting duct of length 2 D, as shown in Fig. 20.6, the inlet system effect loss coeff cients are as follows: R / H of square elbow Cs,i without turning vanes Cs,i with turning vanes

0.5 1.6 0.47

0.75 1.2 —

1.0 0.66 0.33

Here H indicates the dimension of either side of the square duct, in ft (m).

2.0 0.47 0.22

20.20

CHAPTER TWENTY

FIGURE 20.6 Square elbow with turning vanes before fan inlet.

For a square elbo w before a f an inlet with a connecting duct of 5 D and with turning v anes and an inlet transition, Cs,i is approximately 0.6 of the value listed above. If a square elbow before a f an inlet has turning v anes and an inlet transition with no connecting duct, for R / H  1, Cs,i  0.53; and for R / H  0.5, Cs,i  0.8. For double-width double-inlet fan installed in a plenum or in a cabinet, the distance between the fan inlet and the w all of the plenum or cabinet should not be less than 0.4 diameter of the f an inlet. If it is only 0.2 diameter of the f an inlet, an inlet system ef fect loss of Cs,i  0.08 should be added to the system total pressure loss at the design volume f ow rate. For details, refer to AMCA Fan and Systems, Publication 201 (1973). The fans in most air-handling units and packaged units installed according to ASHRAE and ARI standards are rated and tested as complete units. The condition of f an inlet of the actual operating fan-duct system is usually the same as the rated condition. There is no inlet system ef fect loss in such a unit.

Outlet System Effect If a duct f tting is connected to a f an outlet with a short connecting duct, its dynamic loss pdy and local loss coef f cient Cdy increase because of the nonuniform v elocity prof le of the airstream discharged from the f an outlet. The difference between the dynamic losses of a duct f tting connected to a f an outlet with a nonuniform and uniform v elocity prof le at the duct f tting inlet is kno wn as the outlet system effect loss ps,o, in in. WC (Pa). The outlet system effect results in an increase of f ow resistance of the duct system, as shown in Fig. 20.7a. As in Eq. (20.20), outlet system effect loss at design volume f ow rate can be calculated as vfo

2

 4005 

ps,o  Cs,o

where Cs,o  outlet system effect loss coeff cient vfo  fan outlet velocity based on fan outlet area, fpm (m / s)

(20.21)

FIGURE 20.7 Outlet system effect: (a) loss on pt V˙ diagram; (b) effective duct length; ( c) orientation of elbow connected to duct.

20.21

20.22

CHAPTER TWENTY

FIGURE 20.7 (Continued)

Outlet system effect loss at the design v olume f ow should be added to the calculated total system pressure loss.

Outlet System Effect Loss Coefficient Outlet system ef fect loss ps,o and outlet system ef fect loss coef f cient Cs,o both depend on the effective duct length ratio L / Leff, the air velocity at the fan outlet vfo, and the conf guration of the duct f tting with respect to the f an outlet. Here L indicates the length of duct section between f an outlet and the duct f tting in duct diameters. The effective duct length Leff is the length of duct between the f an outlet and a cross-sectional plane at which the uniform v elocity pro f le has been completely reco vered, in duct diameters or equivalent diameters D and L and Leff must be expressed in the same units. Effective duct length can be evaluated as L eff  2.5

vfo 2500 1000

(20.22)

For a single-width single-inlet (SWSI) centrifugal f an, the conf gurations of a 90 ° elbow with respect to the f an outlet at the end of the connecting duct can be classi f ed as 90 ° turn right or 90 ° turn do wn, 90° turn left or 90 ° turn up, parallel with inlet, and opposite to inlet, as sho wn in Fig. 20.7c. According to AMCA Fan and Systems, Publication 201 (1973), the magnitudes of Cs,0.8 at various L / Leff values and elbow orientations with a blast area to fan outlet area ratio Ablast /Aout  0.8 are as follows:

AIR SYSTEMS: BASICS AND CONSTANT-VOLUME SYSTEMS

20.23

L / Leff 90° turn right or down 90° turn left or up, parallel with inlet 90° turn left or up, opposite to inlet

0.12

0.25

0.5

1.0

0.64 1.16 0.80

0.44 0.76 0.52

0.20 0.36 0.24

0 0 0

For a SWSI centrifugal f an with Ablast /Aout  0.7, a multiplier Kblast  1.4 should be applied to the above values. If Ablast /Aout  0.6, then Kblast  2; and if Ablast /Aout  0.9, then Kblast  0.75. For double-width double-inlet (D WDI) centrifugal f ans, an additional multiplier K DI should be applied to the Cs,0.8 values. For a 90° elbow connected to the duct at horizontal turning orientation, KDI  1.25. The outlet system effect coeff cient Cs,o can then be determined as Cs,o  K DIK blastCs,0.8

(20.23)

For details, refer to AMCA, Fan and Systems (1973). For fans installed in air -handling units and packaged units and rated as complete units, the ratings are based on a straight duct length of 2 duct diameters from the unit outlet. The outlet system effect loss is not included in the laboratory-rated fan performance curve. Selecting Fans Considering System Effect Losses If the calculated v olume f ow rate of the design operating point Pd (as shown in Fig. 20.8) is V˙ d, in cfm (L / s), and the calculated system total pressure loss is psy, in in. WC (Pa), the selected fan that

FIGURE 20.8 Selecting fan with performance including system effect loss.

20.24

CHAPTER TWENTY

compensates for inlet and outlet system ef fect losses must operate on a performance curv e that intersects with point Ps, which has a volume f ow rate of V˙ d and a total pressure rise of pss including system effect losses, in in. WC (Pa). That is, pss  psy p s,i p s,o

(20.24)

System effect directly af fects the performance of the air system as well as the function of the air conditioning system. Because the air velocity at the fan inlet and outlet is usually between 1500 and 3000 fpm (7.5 and 15 m / s), system effect losses may amount to 5 to 20 percent of the system pressure loss of the air system. Inlet and outlet system effect losses must be calculated correctly. Moreover, air systems must be designed to reduce system effect losses to save energy.

20.7 COMBINATION OF FAN-DUCT SYSTEMS The purposes of combining fan-duct systems are as follows: ●

●

●

●

To increase the volume f ow rate of the combined fan-duct system To provide the required pressure characteristics for the combined fan-duct system To balance the volume f ow of the combined fan-duct system To maintain a required space pressure characteristic

FIGURE 20.9 Two-fan fan-duct system connected in series.

AIR SYSTEMS: BASICS AND CONSTANT-VOLUME SYSTEMS

20.25

Two Fan-Duct Systems Connected in Series Figure 20.9 shows two fan-duct systems connected in series equipped with f ans F1 and F2 and the corresponding two duct systems represented by f ow resistances R1 and R2. Flow resistance R1 may be the combination of tw o f ow resistances R1A and R1B connected in series, and f ow resistance R2 may be the combination of tw o f ow resistances R2A and R2B. The volume f ow rate of the combined f an-duct system V˙ p, in cfm (L / s), must be the same as the volume f ow rate V˙ 1 or V˙ 2 of fan-duct system 1 or 2 and their components fan F1, fan F2, or V˙p  V˙1  V˙2

(20.25)

The total pressure loss across f ow resistance R1 has a magnitude p1, and the relationship p1  R1V˙ 12 holds. This relationship also determines the system curv e S1. Here the total pressure loss in the duct system includes system ef fect loss. F an performance curv e Ft1 can be obtained from the data given in the manufacturer’s catalog. Total pressure loss p1 and v olume f ow V˙ 1 also determine the operating point of f an-duct system 1 at point Q, which is graphically the intersection of f an performance curve Ft1 and system curve S1. F or f an-duct system 2, p2, V˙ 2, R2, fan performance curv e Ft2, system curv e S2, and system operating point T can be similarly determined.

FIGURE 20.10 fan-duct system.

Comparison of pressure characteristics: (a) combined f an-duct system; ( b) single

20.26

CHAPTER TWENTY

The total pressure loss of the combined fan-duct system pt, in in. WC (Pa), is pt  p1 p2

(20.26)

Graphically, the combination of the curv es of two fans connected in series on a pt-V˙ diagram can be performed by dra wing several constant-volume f ow rate lines. F or each constant-v olume line, the fan total pressure of the combined f an curve Ft1 Ft2 always equals the sum of the sections of fan total pressure of f an curve 1, Ft1 and fan curve 2, Ft2, represented by p1 and p2, as shown in Fig. 20.9. The purpose of connecting tw o f an-duct systems in series is to increase the f an total pressure that can be provided by the combined system. Figure 20.10 sho ws the system pressure characteristics of a combined f an-duct system with fans F1 and F2 and a single f an-duct system that uses only one f an F. It can be seen that the static pressure across the duct w all is considerably higher in a f an-duct system with a single f an than in the combined fan-duct system with two fans. Of course, a single-fan system has a lower initial cost. For two fan-duct systems connected in series, the volume f ow rates of the tw o fans should be similar. If a large fan is connected in series with a small fan, at large volume f ow rates, the fan total pressure of the combined system may be less than if the lar ge fan were operated alone. This loss of eff ciency results from the inf uence of negative total pressure of the small fan when it is operated at a volume f ow rate greater than its wide-open volume f ow.

Fans Combined in Parallel and Connected in Series with a Duct System When two fans F1 and F2 combined in parallel are connected to a duct system represented by f ow resistance R, as shown in Fig. 20.11, the volume f ow rate of the parallel combined f ans V˙ P is the sum of the volume f ow rates of individual fans V˙ 1 and V˙ 2. The purpose of such a combination is to increase the v olume f ow rate of the system. Usually , the fans in parallel combination are identical f ans, i.e., fans of the same model with the same rated

FIGURE 20.11

Two fans combined in parallel and connected in series with a duct system.

AIR SYSTEMS: BASICS AND CONSTANT-VOLUME SYSTEMS

20.27

volume f ow and f an total pressure. As in f an-duct systems connected in series, a large fan should not be combined in parallel with a small fan because of the possible negative effect of the small fan. The procedure for determining the combined f an performance curve Ft1 Ft2 of the two identical fans connected in parallel is as follows: 1. Plot the fan performance curves Ft1 and Ft2 according to the manufacturer’s data. 2. Draw the constant total pressure loss pt lines (horizontal lines). Points on the combined curv e along the constant pt lines have a v olume f ow of 2 V˙ 1. Here V˙ 1 indicates the v olume f ow o f each identical fan along the constant pt lines. 3. Plot the combined curv e Ft1 Ft2 by connecting points ha ving V˙  2V˙ 1 on the constant pt lines. The system curve can be plotted based on the relationship pt  RV˙ 2. The intersection of the combined curve Ft1 Ft2 and the system curve is the system operating point P. The volume f ow at the system operating point V˙ P, in cfm (L / s), is the volume f ow rate f owing through the duct system at a f ow resistance R. The combined system total pressure at the system operating point pP, in in. WC (Pa), is the f an total pressure of the identical f ans or the system total pressure loss of the duct system. Example 20.1. An air system uses two identical fans combined in parallel and connected in series with a duct system, as shown in Fig. 20.11. The pressure-volume and power-volume characteristics of each of the identical fans are listed below: V˙ , cfm pt, in. WC Power Pf, hp

2500 5.00 3.93

5000 4.88 5.90

7500 4.54 7.14

10,000 4.00 7.87

12,500 3.13 8.20

15,000 1.75 7.60

If the total pressure loss of this air system is kno wn to be 1 in. WC (250 Pa) at an airf ow of 10,000 cfm (4719 L / s), determine the following: 1. The system operating point, the corresponding air v olume f ow rate, and the system total pressure when both fans are operating 2. The operating point, the corresponding airf ow rate, and the fan total pressure when only one fan is running, assuming that the air damper after the other fan is closed 3. The fan total eff ciency when two fans or only one fan is running Solution 1. From the given data, the f ow resistance of this air system is R

pt 1   1  10 8 in. WC / (cfm)2 (10,000)2 V˙ 2

The system pressure loss at various airf ow rates is as follows: V˙ , cfm pt, in. WC

5000 0.25

10,000 1.0

15,000 2.25

20,000 4.0

25,000 6.25

After the single-fan performance curve is plotted, the combined curve of the two fans in parallel can be dra wn. Plot the system curv e from the data gi ven above. The volume f ow rate and the

20.28

CHAPTER TWENTY

system pressure loss at operating point P, which is the intersection of the combined curv e and the system curve when two fans are running (as shown in Fig. 20.11), are found to be V˙p  20,000 cfm (9440 L / s)

pP  4.0 in. WC (995 Pa)

2. When only one f an is operating and the damper after the other f an is closed, the intersection of the single-fan curve and the system curve is at point Q (see Fig. 20.11). At point Q V˙ Q  14,500 cfm (6840 L / s),

pQ  2.06 in. WC (510 Pa)

3. When two fans are running, each fan has a volume f ow of 20,000 / 2  10,000 cfm and a fan total pressure of 4 in. WC. From the po wer curve, the corresponding f an power input is 7.75 hp. From Eq. (15.6), the fan total eff ciency is then

two 

ptV˙ 4  10,000   0.81, or 81 percent 6356Pf 6356  7.75

When only one f an is running, the corresponding fan power input is also 7.75 hp. The fan total eff ciency is

one 

2.06  14,500  0.61, or 61 percent 6356  7.75

Two Parallel Fan-Duct Systems Connected with Another Duct System When two parallel f an-duct systems are connected in series to a third duct system represented by f ow resistance R3, as shown in Fig. 20.12, the volume f ow rate V˙ 3 f owing through R3 is the sum of volume f ow rates f owing through the parallel combined fan-duct systems V˙ 1 and V˙ 2 V˙BC  V˙3  V˙1 V˙2 Although the total pressure loss pAB between points A and B of parallel combined f an-duct systems 1 and 2 is the same, f ow resistances R1 and R2 may be different from each other; therefore, at their own operating points, the fan total pressure of f an F1, denoted by pt1, may also be dif ferent from the fan total pressure of fan F2, pt2, even though fans F1 and F2 are identical. Through the concept of residual pressure of a f an-duct system, the performance of tw o parallel combined fan-duct systems can be determined. When fan F1 is connected in series with f ow resistance R1, the total pressure loss across R1 is pR1  R1V˙ 2. If the v olume f ow rate of air f owing through R1 is of a given value V˙ 1, the residual pressure after f owing through the f ow resistance R1, pres1, in in. WC (Pa), is pres 1  pF1 pR1  pF1 R 1V˙ 12  Ft1 S1

(20.27)

where pF1, Ft1  fan total pressure of fan F1, in. WC (Pa) pR1, S1  total pressure loss across f ow resistance R1 or total pressure loss of duct system 1, in. WC (Pa) For fan-duct system 2, residual pressure can be similarly calculated. For two fan-duct systems connected in parallel, the residual pressure of f an-duct system 1 after R1, pres1, must be equal to the residual pressure of f an-duct system 2 after R2, pres2, because there is only one unique total pressure ptB at point B. If two parallel combined fan-duct systems are connected in series with another duct system ha ving f ow resistance R3, the residual pressure provided after fan-duct systems 1 and 2 can be used to overcome the total pressure loss across f ow resistance R3 is as follows: pres1  pres2  R 3V˙ 23

(20.28)

AIR SYSTEMS: BASICS AND CONSTANT-VOLUME SYSTEMS

FIGURE 20.12

20.29

Two parallel fan-duct systems connected with another duct system.

Because the pressure at points A and C is equal to atmospheric pressure (zero), the performance of two parallel combined f an-duct systems connected in series with another duct system can be determined as follows: 1. Plot two fan curves Ft1 and Ft2, the residual pressure curv es Ft1 S1 and Ft2 S2, and the combined residual pressure curv e Ft1 S1 Ft2 S2. The residual pressure Ft S curve can be plotted as follows: At a v olume f ow of V˙A, f nd a corresponding f an total pressure ptfA  FtA from the f an curve given by the manufacturer. Calculate the pressure drop across the duct system at a v olume f ow V˙A : SA  ptRA  RV˙A2. Calculate residual pressure presA  FtA SA at a volume f ow of V˙A. Plot point A on the residual pressure Ft S curve at a volume f ow of V˙A and a residual pressure of presA. Similarly, plot points B, C, . . . on the residual pressure curve. Residual pressure curve Ft S can be drawn by joining points A, B, C, . . . etc. ●

● ● ●

● ●

At each point on a residual curv e, there is a corresponding residual pressure pres  (Ft RV 2) that can be pro vided at a corresponding v olume f ow of V˙ to overcome the f ow resistance R3 of a

20.30

CHAPTER TWENTY

connected series. Combining of residual pressure curv es is performed in the same w ay as combining of fan curves. 2. Calculate R3 from the required v olume f ow rate and total pressure loss in that duct system represented by R3. From the relationship pt  R 3V˙ 23 , draw system curve S3. 3. Determine the intersection point P of the combined residual pressure curv e Ft1 S1 Ft2

S2 and the system curv e S3, which denotes the operating condition at R3. The residual pressure at point P is the pressure that can be pro vided by fan-duct systems 1 and 2 to o vercome the total pressure loss across R3 at that volume f ow rate V˙3  V˙ V˙2. In fan-duct system design, if the volume f ow rate through R3 thus found is greater or smaller than the required v olume f ow rate, select a larger or smaller fan in duct systems 1 or 2 to make them approximately the same. 4. Draw a horizontal line PQ from point P, which intersects with residual pressure curv es at points T and Q, and draw vertical lines QQ and TT from points Q and T. Points Q and T on fan curve F1 or F2 indicate the volume f ow, fan total pressure of the f an-duct systems 1 and 2, and fan power input to fan F1 and fan F2. Example 20.2. An air system is equipped with two parallel combined fan-duct systems connected in series with another duct system, as sho wn in Fig. 20.12. The f ans in the f an-duct systems are identical and ha ve the same pressure-v olume and power-volume characteristics as listed in Example 20.1. Flow resistances R1, R2, and R3 have the following pressure-volume characteristics: ●

●

●

R1 has a pt loss of 2.29 in. WC at a V˙ of 10,500 cfm. R2 has a pt loss of 2.82 in. WC at a V˙ of 9500 cfm. R3 has a pt loss of 1.50 in. WC at a V˙ of 20,000 cfm.

Determine 1. Air volume f ow rate f owing through the duct system having f ow resistance R3 2. Operating point of fan-duct systems 1 and 2 3. Fan total eff ciency of fans 1 and 2 Solution 1. From the given data, the f ow resistances can be calculated as R1 

pt 2.29   2.08  10 8 in. WC / (cfm)2 (10,500)2 V˙ 2

R2 

2.82  3.12  10 8 in. WC/(cfm)2 (9500)2

R3 

1.5  3.75  10 9 in. WC/(cfm)2 (20,000)2

2. From Eq. (20.27), the residual pressure of f an-duct systems 1 and 2 after R1 and R2 at a volume f ow rate of 2500 cfm can be calculated as pres1  Ft1 S1  pF1 R 1V˙ 21  5.0 2.08  10 8  (2500)2  4.87 in. WC pres2  Ft2 S2  5.0 3.12  10 8  (2500)2  4.80 in. WC

AIR SYSTEMS: BASICS AND CONSTANT-VOLUME SYSTEMS

20.31

Similarly, the residual pressure of other volume f ow rates can be calculated as listed below: V˙ , cfm 2500

5000

7500

10,000

12,500

5.0 0.13 0.20 4.87 4.80

4.88 0.52 0.78 4.36 4.10

4.54 1.17 1.76 3.37 2.78

4.0 2.08 3.12 1.92 0.88

3.13 3.25 4.88

Ft, in. WC R1 V˙ 21 , in.WC R2 V˙ 22 , in.WC Ft1 S1, in. WC Ft2 S2, in. WC

3. At various volume f ow rates f owing through the duct system with f ow resistance R3, the total pressure loss pt  R 3V˙ 23 can be calculated as follows: V˙ , cfm

5000

10,000

15,000

20,000

25,000

S3  R 3V˙ 23 , in. WC

0.09

0.38

0.84

1.50

2.34

4. Plot the single-f an curve Ft, the residual pressure curv es Ft1 S1 and Ft2 S2, the combined residual pressure curv e Ft1 S1 Ft2 S2, and the system curv e S3 on pt V˙ diagram, as shown in Fig. 20.12. The intersection point P of the combined residual pressure curv e Ft1

S1 Ft2 S2, and the system curve S3, gives the volume f ow rate in the duct system with f ow resistance R3. From the diagram, V˙p  20,000 cfm (9438 L / s). 5. Draw horizontal line PQ from point P to intersect the Ft2 S2 curve at Q and Ft1 S1 curve at point T. Again, draw vertical lines QQ and TT from points Qand T. The lines intersect fan curve Ft at points Q and T. For fan-duct system 1, operating point T gives V˙1  10,500 cfm(4955 L / s)

and

pR1  3.8 in. WC (945 Pa)

For fan-duct system 2, the operating point Q gives V˙2  9500 cfm(4480 L / s)

and

pR2  4.15 in. WC (1032 Pa)

6. From the plotted po wer-volume curve in Fig. 20.12, for operating point T, the fan power input Pf 1  7.8 hp, and for operating point Q, the fan power input Pf 2  7.6 hp. F an total ef f ciency of fan 1 is therefore

1 

V˙1 pR1 10,500  3.8   0.80, or 80 percent 6356Pf1 6356  7.8

Fan total eff ciency for fan 2 is

2 

9500  4.15  0.82, or 82 percent 6356  7.6

20.8 MODULATION OF THE FAN-DUCT SYSTEM Modulation Curve The modulation curve of a fan-duct system or, practically, an air system, is its operating curve when its volume f ow rate is modulated at part-load operation. A modulation curve is also the locus of the system operating points at reduced system loads and volume f ow rates.

20.32

CHAPTER TWENTY

FIGURE 20.13

Modulation curve of a VAV system installed with duct static pressure control.

For a v ariable-air-volume system with duct static pressure control, the pressure loss of the air system can be divided into f xed part pf x and variable part pvar when the volume f ow rate varies. This is described in detail in Chap. 21. Figure 20.13 sho ws a typical modulation curv e of a VAV system with duct static pressure control using inlet vanes to modulate f an capacity. This VAV system has man y fan curves Ft at various volume f ow rates and a system curve for the variable pressure loss Svar. During reduction of the v olume f ow rate of a VAV system, the dampers in the VAV boxes close partially. The variable part of the pressure loss of the air system pvar from the AHU or PU up to the pressure sensor of duct static pressure control varies as the volume f ow is reduced. Its pressure-volume characteristics are indicated by the system curv e Svar. However, the pressure loss of the duct system, branch takeoffs, and VAV boxes beyond the duct static pressure sensor remains constant, and is equal to the set point of duct static pressure control pf x. The modulation curv e is actually the system curve of the part of VAV air system with variable pressure loss Svar. At the same time, the inlet v anes at the f an inlet also reduce their opening in response to the sensed higher duct static pressure. Therefore, a new fan curve Ft2 is formed. The intersection of Ft2 and system curve Svar at the reduced v olume f ow rate is the system operating point B. Similarly, at a further reduction of volume f ow rate, the system operating points are C, D, E . . . . As the v olume f ow rate of the air system approaches zero, theoretically, the system total pressure will equal the f xed part of pressure loss pf x, which is equal to the set point of duct static pressure control, as shown in Fig. 20.13. The modulation curve of a fan-duct system should not enter the f an surge area or stall region, or unstable operation will result. Ho wever, a forw ard-curved centrifugal f an may be operated in the surge area if the system total pressure is belo w 1.5 in. WC. Refer to the manufacturer’s data for detailed analysis.

AIR SYSTEMS: BASICS AND CONSTANT-VOLUME SYSTEMS

20.33

Modulation of Fan-Duct Systems The performance of a f an-duct system can be modulated either by changing the f an characteristics or by varying the f ow resistance of the duct system. This can be achie ved by any of the follo wing methods. Modulation Using Dampers. Use of a damper is the simplest method of modulation. A multiblade damper is usually installed inside the main duct after the centrifugal supply f an. When the damper closes, the system f ow resistance increases and the system operating point mo ves from point A to B on a different system curve. This method is kno wn as “riding on the curv e,” as shown in Fig. 20.14a. If the f an Ft curve is f atter, then the increase in pt during reduction of the v olume f ow rate of the fan-duct system is small. Because of the considerable total pressure loss across the damper , power input to the fan is only slightly decreased by this method. Modulation dampers ha ve the lo west installation cost and only small energy savings.

FIGURE 20.14 Modulation of fan-duct systems: (a) using dampers; (b) using inlet vanes; (c) using ac inverter to vary fan speed.

20.34

CHAPTER TWENTY

FIGURE 20.14

(Continued)

Modulation Using Inlet Vanes. Modulation by varying the opening of inlet v anes at the centrifugal fan inlet gives different fan performance curves and therefore different system operating points. It is widely used in many VAV systems. The surge area of a fan with inlet vanes is smaller than that of a fan without inlet vanes. For a VAV system with duct static pressure control, a modulation curve and the corresponding f an power inputs are sho wn by points A, B, and C on the po wer-volume curves, as shown in Fig. 20.14b. Compared to modulation by damper , modulation using inlet v anes has a moderately higher installation cost and considerable energy savings. Modulation Using Inlet Cone. Moving the inlet cone of a backw ard-curved or airfoil centrifugal fan also gi ves different f an performance curv es, so that its modulation curv e is similar to that of

AIR SYSTEMS: BASICS AND CONSTANT-VOLUME SYSTEMS

FIGURE 20.14

20.35

(Continued)

inlet vanes. Compared to modulation by inlet v anes, modulation using inlet cone blocks far less fan inlet area. Modulation by inlet cone for backw ard-curved and airfoil centrifugal f ans has a comparati vely lower installation cost and also considerable energy savings. Modulation by Blade Pitch Variation of Axial Fan. Modulation of the volume f ow of a fan-duct system by blade pitch (blade angle) v ariation of an axial f an changes the f an characteristics and, therefore, the system operating points. Modulation by automatic blade pitch v ariation signif cantly

20.36

CHAPTER TWENTY

lowers the f an ener gy consumption at reduced v olume f ow rate. Although automatic blade pitch variation is rather expensive, it has become more popular then before. Modulation by Varying Fan Speed. Modulation of the f an-duct system by v arying the f an speed renders new fan performance curves n2 and n3 and, therefore, new system operating points B and C at reduced volume f ow rates, as shown in Fig. 20.14c. The modulation curve of a VAV system with duct static pressure control can be obtained in the same way as for inlet vanes. Fan power input for f an speed modulation is lo wer than for damper or inlet v ane modulations. Because of the fan speed reduction at reduced V˙ , the fan sound power level is reduced accordingly. Fan speed v ariation using adjustable-frequenc y variable-speed drives in a VAV system has become more popular. An adjustable-frequency ac inverter consumes only 7 to 8 percent more ener gy input to the f an motor. The high initial cost of adjustable-frequenc y variable-speed drives is often cost-effective for large centrifugal fans. The initial cost may drop further in the future as more and more variable-speed drives are installed. Because of the differences in fan characteristics, inlet vane, inlet cone, and variable-speed drive modulation is widely used in centrifugal fan-duct systems, and blade-pitch and variable-speed drive modulation is used for axial fan-duct systems. Example 20.3. A v ariable-air-volume system equipped with a backw ard-curved centrifugal f an with airfoil blades operated at 1700 rpm has the follo wing pressure-v olume and po wer-volume characteristics: V˙ , cfm pt, in. WC Power, hp

5000 6.0 10.0

10,000 6.0 15.2

15,000 6.0 18.7

20,000 4.65 21.0

25,000 2.70 20.0

At design condition, this system has a volume f ow rate of 20, 000 cfm (9438 L / s) and a system total pressure loss of 4.65 in. WC (1156 Pa). When it is operated at 50 percent of its design f ow rate and its f ow rate is modulated by inlet v anes, its performance curv es have the follo wing characteristics: V˙ , cfm pt, in. WC Pf, hp

5000 4.88 9.6

7500 4.55 11.2

10,000 4.0 12.0

12,500 3.12 12.3

15,000 1.87 12.0

Determine the f an power input if the v olume f ow rate is reduced to 50 percent of its design rate by the following methods: 1. Volume control damper 2. Inlet vanes, where system total pressure loss at 50 percent design f ow is 4 in. WC 3. Fan speed variation by means of an adjustable-frequency variable-speed drive Solution 1. From the given data, the f ow resistance of this VAV system at design condition is Rd 

pt 4.65   1.16  10 8 in. WC / (cfm)2 2 2 ˙ (20,000) V

The system total pressure loss at various volume f ow rates can then be calculated as S  pt  R dV˙ 2  1.16  10 8 V˙ 2

f ow

AIR SYSTEMS: BASICS AND CONSTANT-VOLUME SYSTEMS

20.37

Plot the fan performance curve Ft and the system curve S. The intersection of these curves is the system operating point P at design condition, as shown in Fig. 20.15a. 2. When the volume f ow rate is reduced by a damper to 50 percent of its design v olume f ow, that is, 0.5  20,000  10,000 cfm (4719 L / s), its operating point must lie on the f an performance curve with a V˙  10,000 cfm (4719 L / s), point Q. The fan power input at point Q is 15.2 hp (11.3 kW), as shown in Fig. 20.15a. 3. When the v olume f ow is reduced by modulation of inlet v anes to 10,000 cfm (4719 L / s), the system total pressure loss is 4 in. WC (994 Pa), so the f ow resistance of this VAV system at reduced f ow using inlet vanes is

R iv 

4  4  10 8 in. WC / (cfm)2 (10,000)2

From the given data, plot the f an performance curv e using inlet v anes Fiv and the system curv e Siv  Riv V˙ 2 , as shown in Fig. 20.15b. Also plot the power-volume curve. The intersection point Piv of the Fiv and Siv curves is the system operating point at 10,000 cfm (4719 L / s) volume f ow, and its fan power input is 12 hp (8.95 kW), as shown in Fig. 20.15b.

FIGURE 20.15 Modulation of a VAV system using dampers, inlet v anes, and f an-speed v ariation: (a) using dampers and fan-speed variation; (b) using inlet vanes.

20.38

CHAPTER TWENTY

FIGURE 20.15

(Continued)

4. For the same f an-duct system, from the f an law, the fan speed required at reduced v olume f ow of 10,000 cfm (4719 L / s) is V˙2

 850 rpm  V˙   1700  10,000 20,000 

n2  n1

1

Fan power input at a fan speed n2  850 rpm is n2

 2.63 hp (1.96 kW)  n   21  10,000 20,000 

P2  P1

3

3

1

The dif ference in f an po wer inputs between using damper and v ariable-speed dri ve f an modulation is the po wer consumption at the damper: 15.2 2.63  12.5 hp (9.32 kW). At reduced f an speeds, fan total ef f ciency may be reduced. Because of the de gradation of motor eff ciency and ac in verter loss at part-load operations actual po wer input at 50 percent of part-load operation for v ariable-speed dri ve modulation is f ar higher than 2.63 hp (1.96 kW).

AIR SYSTEMS: BASICS AND CONSTANT-VOLUME SYSTEMS

20.39

20.9 CLASSIFICATION OF AIR SYSTEMS Air systems, in a narro wer sense, can be classi f ed into the follo wing categories according to their system characteristics: ●

●

●

●

Air-handling systems, air handlers, and air distribution systems. Supply air systems, return air systems, recirulating air systems, and exhaust air systems. A supply air system may supply conditioned air , outdoor ventilation air, makeup air (either conditioned air or nonconditioned air), or recirculating air. A return air system often returns the conditioned space air to the fanroom or machinery room. A recirculating air system transports the conditioned space air to the AHU, PU, fan-coil unit, or w ater-source heat pump for conditioning or mixing with outdoor air again. An e xhaust system e xhausts the contaminated air or space air to the outside atmosphere. Single-zone or multizone systems. In a single-zone system, there is no terminal. Space temperature, relative humidity, and volume f ow rate are controlled directly by the coils, humidif ers, and inlet vanes or ac in verter in the air -handling unit or packaged unit. In a multizone system, zone temperature or zone supply volume f ow rate is controlled by terminals. Fan combination systems. Three fan combination systems are often used: supply and exhaust fan combination, supply and relief f an combination, and supply and return f an combination systems.

Relief fan

Condensing unit Ceiling plenum VAV box

Vent Supply air

Optional exhaust fan

Exfiltration

Return fan Supply air

DXcoil

Conditioned space

Primary and secondary heat exchanger

(a)

PU Slot diffuser

Filter

Return slots

Return air

Fan room

Dedicated outdoor ventilation air system

m s

cf

Conditioned space

T2

(c) FIGURE 20.16 Various types of air systems: space recirculating system.

I

(b)

Space recirculating air system Fan coil

fc

O

Outdoor air

Outdoor air

Recirculating air

Return duct AHU

O

o

Outdoor air AHU

(a) constant-v olume system; ( b) v ariable-volume system; ( c) dedicated v entilation and

20.40

CHAPTER TWENTY ●

●

●

Single-duct or dual-duct systems. In a single-duct system, conditioned air is supplied to the conditioned space by a single supply duct. In a dual-duct system, conditioned air is supplied to the conditioned space in the perimeter zone through tw o supply ducts: a warm air duct and a cold air duct. In the interior zone, only a cold air duct is needed. Constant-volume (CV) or v ariable-air-volume systems. In a constant-v olume system, the temperature of supply air is modulated to match the v ariation of space load during part-load operation, as shown in Fig. 20.16a. In a VAV system, the supply volume f ow rate is modulated to maintain a predetermined space temperature as space load varies, as shown in Fig. 20.16b. Dedicated ventilation and space recirculating system. In a dedicated ventilation and space recirculating system, a required amount of outdoor ventilation air is supplied by a separate ventilation air system; at the same time, there is a parallel space recirculating system to condition the space recirculating air to offset the space heating or cooling load, as shown in Fig. 20.16c. Outdoor ventilation air is either mix ed with the space recirculating air f rst or directly supplied to the conditioned space. When the outdoor and recirculating air are mixed in the mixing box (mixed plenum) f rst; the mixture is then conditioned and supplied to the conditioned space.

Because a constant-v olume system is often a single-duct and outdoor recirculating air mixing system, a variable-air volume system is usually also an outdoor recirculating air mixing system, and a space conditioning system al ways employs a separate v entilation air and a parallel space recirculating air system. Therefore, air systems can be mainly classi f ed into following three primary categories: ●

●

●

Constant-volume, single-duct outdoor recirculating air mixing systems, or simply constantvolume systems Variable-air-volume outdoor recirculating air mixing systems, or simply v ariable-air-volume systems Dedicated ventilation and space recirculating systems

Each cate gory of air system has supply , return, recirculating, or e xhaust systems and other specif c functions. Dedicated ventilation and space recirculating systems are discussed in Chap. 28.

20.10 CONSTANT-VOLUME SYSTEMS System Characteristics A single-zone, single-duct, constant-volume, and outdoor air and recirculating air mixing system is shown in Fig. 20.16 a. A constant-volume system means that there is a constant v olume of supply air throughout the operating period. Supply air temperature is varied when the space load reduces in part-load operation. Single-zone states indicates that the system serv es a conditioned space which is controlled to maintain a unique indoor temperature, relative humidity, cleanliness, and pressure differential. Single-zone constant-volume systems ha ve been widely used in residential b uildings and small retail stores in commercial b uildings. They often emplo y a packaged unit with a DX coil instead of water cooling coil in an air-handling unit. In the PU, there is often a single supply fan, and the conditioned space is generally positi vely pressurized at 0.005 to 0.03 in. WC (1.25 to 7.5 P a), so that the amount of outdoor v entilation air intak e at the packaged unit can be e xf ltrated through the cracks and gaps on the b uilding shell; or sometimes a small e xhaust fan is installed in the restroom. F or a shopping mall or an indoor stadium, an AHU with a return f an is often used. Man y air systems that serve clean rooms are single-zone constant-v olume systems. These air systems are demanding, complicated, and expensive, and are discussed in Chap. 30. In health care facilities and industrial applications, a single-zone constant-volume system or sometimes a multizone system with reheat is used.

AIR SYSTEMS: BASICS AND CONSTANT-VOLUME SYSTEMS

20.41

To sa ve ener gy, most AHU and PU manuf acturers also of fer air economizers for single-zone constant-volume systems when the supply v olume f ow rate of the air system is 2000 cfm (1000 L / s) and greater . Most single-zone constant-v olume systems emplo y zone temperature control, night setback, diagnostics, emergency stop, and safety controls. Air economizer and air system controls are discussed in the chapters that follow.

Energy per Unit Volume Flow As speci f ed in ASHRAE / IESNA Standard 90.1-1999 and discussed in Sec. 17.2, the f an power input to the f an motors per unit v olume f ow Psy / V˙sd in hp/cfm (W ·s / L), for constant-volume air systems as calculated by Eq. (17.24a): Psy  0.0012 hp / cfm (0.0019 W s / L) V˙sd

when V˙sd  20,000 cfm (9440 L / s) (20.29)

Psy  0.0011 hp / cfm (0.00174 W s / L) V˙sd

when V˙sd  20,000 cfm

For variable-air-volume (VAV) systems, the fan power input per unit volume f ow can be calculated by Eq. (17.24b) as: Psy  0.0017 hp / cfm (0.0027 W s / L) V˙sd

when V˙sd  20,000 cfm

Psy  0.0015 hp / cfm (0.0024 W s / L) V˙sd

when V˙sd  20,000 cfm

20.11 AIR CONDITIONING PROCESSES An air conditioning process determines the change in thermodynamic properties of moist air between the initial and f nal states of conditioning and also the corresponding ener gy and mass transfer between the moist air and a medium, such as w ater, refrigerant, or moist air itself during this change. The energy balance and the conservation of mass in nonnuclear processes are the tw o principles most often used in the analysis and calculation of the change of thermodynamic properties in air conditioning processes. In general, for a single air conditioning process, heat transfer or mass transfer is positi ve. However, for calculations that in volve several air conditioning processes, the heat supplied to the moist air is taken to be positive, and the heat rejected from the moist air is taken to be negative.

Sensible Heat Ratio The sensible heat ratio (SHR) of an air conditioning process is de f ned as the ratio of the absolute value of sensible heat to the absolute value of total heat, or SHR 

 qsen   qtotal 



 qsen   qsen   ql 

In Eq. (20.30), total heat, in Btu / h (W), is given as qtotal   qsen   ql 

(20.30)

20.42

CHAPTER TWENTY

FIGURE 20.17

Space conditioning lines.

Figure 20.17 shows an air conditioning process from an initial state s to f nal state r. Refer to Secs. 2.9, 2.10, and Appendix Table B.4 for the SI unit conversion of the parameters on the psychrometric chart in Fig. 20.17 as well as other f gures. The sensible heat change qsen in this process, in Btu / h (W), can be calculated as qsen  60V˙s scpa(Tr Ts)  60m˙acpa(Tr Ts) where V˙s  volume f ow rate of supply air, cfm [m /(60 s)] m˙a  mass f ow rate of supply air, lb / min [kg / (60 s)] s  density of supply air, lb / ft3(kg / m3) cpa  specif c heat of moist air, Btu / lb  °F (J / kg · °C) Ts,Tr  moist air temperature at initial and f nal states, °F (°C) 3

Latent heat change ql in this process (in Btu / h or W) is given by ql  60V˙ss(wr ws)h fg

(20.31)

AIR SYSTEMS: BASICS AND CONSTANT-VOLUME SYSTEMS

20.43

where ws,wr  humidity ratio at the initial and f nal states, lb / lb (kg / kg). Theoretically, hfg, in Btu / lb (J / kg), is the latent heat of v aporization of w ater at the temperature where v aporization or condensation occurs. But because the sensible heat change of m˙acpa(wr ws)(Tr Ts) is already included in Eq. (20.31), it is convenient to use hfg,32 cps(T 32) to replace hfg. The error is usually less than 0.03 percent. Then the latent heat change ql can be calculated as ql  60V˙ss(wr ws)h fg,32  60  1075V˙ss(wr ws)  60m˙a(wr ws) (1075)  64,500V˙ss(wr ws)

(20.32)

In Fig. 20.17, the slope of the air conditioning process sr is given by tan  

Lw Ct(wr ws)  Lt Cw(Tr Ts)

(20.33)

where   angle between air conditioning process and horizontal line on psychrometric chart, deg Lw, Lt  vertical and horizontal distance between state points r and s, ft (m) Cw,Ct  scale factor for humidity ratio lines and temperature lines, lb / lb  ft (kg / kg  m) and °F / ft (°C / m) Substituting Eqs. (20.31) and (20.32) into Eq. (20.30) gives SHR 

m˙acpa(Tr Ts) V˙aacpa(Tr Ts ) m˙a(wr ws )h fg,32 

1

(20.34)

1 (h fg,32Cw tan ) / (cpaCt)

In ASHRAE’s psychrometric chart, as shown in App. Fig. B.1, additional SHR lines are gi ven by joining the intersection of T  78°F,   50 percent, and the outer sensible heat ratio scale. An air conditioning process that is parallel to any one of these SHR lines has the same SHR.

20.12 SPACE CONDITIONING AND SENSIBLE COOLING AND HEATING PROCESSES Space Conditioning Process A space conditioning process is an air conditioning process in which either: 1. Heat and moisture are absorbed by the supply air and removed from the space, or 2. Heat, or sometimes heat and moisture, is supplied to the space to compensate for the transmission and in f ltration losses through the b uilding shell, with moisture being gi ven up by the supply air to the space. The purpose of these processes is to maintain a desirable space temperature and relative humidity. Figure 20.17 sho ws two lines, sr and sr, to indicate these tw o space conditioning processes. These lines are called space conditioning lines. The upper line, sr denotes the absorbtion of space heat and moisture during summer , and the bottom line indicates the supply of heat and absorption of moisture during winter. In space conditioning processes, assuming that the kinetic energy difference between the supply inlet s and return e xit r is negligible, and that there is no w ork being done during these processes,

20.44

CHAPTER TWENTY

the steady f ow energy equation can be simplif ed to 60 m˙ah s qrc  60 m˙ah r

(20.35)

where hr, hs  enthalpy of space air and supply air, Btu / lb (J / kg) Qrc  heat to be removed from conditioned space, or space cooling load, Btu / h (W) Rearranging the terms, then, we can calculate the space cooling load as Q rc  60 m˙a(h r h s)  60 V˙ss(h r h s)

(20.36)

where V˙s  volume f ow rate of supply air, cfm [m /(60 s)] s  density of supply air, lb / ft3 (kg / m3) 3

Considering the mass balance during the space conditioning process m˙aws m˙g  m˙awr where wr, ws  humidity ratio at exit r and inlet s, lb / lb (kg / kg) m˙g  rate of moisture gain in conditioned space, lb / min [kg / (60 s)] Again, by rearranging the terms, the rate of space moisture gain can be evaluated as m˙g  m˙a(wr ws )  V˙ss(wr ws )

(20.37)

Sensible Heating and Cooling Processes A sensible heating process is a process in which heat is added to the moist air , resulting in an increase in its temperature, while its humidity ratio remains constant, as discussed in Sec. 15.9. This process is represented by line el in Fig. 15.29. The sensible heating process occurs when moist air f ows through a heating coil in which heat is transferred from the hot w ater inside the coil tubes to the moist air , or through a heat e xchanger where heat transfer tak es place between tw o f uid streams. The rate of heat transfer from the hot f uid to the cold f uid Qch, in Btu / h (W), is called the heating coil load or the heating capacity of the heat exchanger. From Eq. (15.24) the heating coil load can be evaluated as Q ch  60m˙a(h l h e )  60V˙s s(h l h e )

(20.38)

where he,hl  enthalpy of moist air entering and lea ving coil (or heat e xchanger), Btu / lb (J / kg). From Eq. (2.29), h  cpdT w (h g0 cpsT ) and from Eq. (2.36) cpa  cpd wcps Substituting Eq. (2.29) into Eq. (20.38), assuming that the dif ferences between cpd1 and cpd2, and cps1 and cps2, are negligible, and that for a sensible heating process the humidity ratio of moist air entering is equal to leaving air we  wl, Q ch  60m˙acpa(Tl Te)  60V˙sscpa(T1 T2)

(20.39)

A sensible cooling process removes heat from the moist air, resulting in a drop in its temperature while maintaining a constant humidity ratio of the moist air , as discussed in Sec. 15.9. This process is represented by line e l in Fig. 15.29. The sensible cooling process occurs when moist air f ows through a cooling coil and there is an indirect heat transfer from the moist air to the chilled w ater inside the cooling coil; or in an air-to-air heat exchanger where two airstreams do not make contact. In a sensible cooling process, the rate of heat transfer Qcs, in Btu / h (W), from the moist air to the chilled w ater inside the cooling coil or from one airstream to another airstream in the heat

AIR SYSTEMS: BASICS AND CONSTANT-VOLUME SYSTEMS

20.45

exchanger is called the sensible cooling coil load or sensible cooling capacity of the heat exchanger. It can be calculated as Q cs  60m˙acpa(Te Tl)  60V˙sscpa(Te Tl)

(20.40)

20.13 HUMIDIFYING PROCESSES In a humidifying process, water vapor is added to the moist air , which increases the humidity ratio if the initial moist air is unsaturated. As discussed in Chap. 15, air is humidi f ed by the follo wing methods: 1. Steam injection or submerged heating element 2. Atomizing devices 3. Wetted media evaporation or spraying air washers Steam Injection and Heating Element Humidifier In a humidifying process using steam injection, steam is often supplied from the main line to a grid-type humidif er and then injected directly into the air through small holes, as discussed in Sec. 15.19 and sho wn in Fig. 20.18. The steam injection humidifying process is indicated by line 12,

FIGURE 20.18

Humidifying processes.

20.46

CHAPTER TWENTY

which is approximately parallel to the constant temperature lines on the psychrometric chart. The slight inclination to ward the right-hand side at the top of line 12 is due to the high temperature of the injected steam. The increase in the moist air temperature because of the steam injection can be calculated as follows: When a mass f ow rate of m˙s lb / h [kg / (60 s)] of dry saturated steam at lo w pressure is injected into a moist airstream of a mass f ow rate of m˙a, lb / h [kg / (60 s)], according to the principle of heat balance m˙ah 1 m˙sh s  m˙ah 2

(20.41)

where h1, h2  enthalpy of moist air entering and leaving steam injection humidifier, Btu / lb (J / kg) hs  enthalpy of injecting steam, Btu / lb (J / kg) The mass f ow rate of injecting steam is def ned as m˙s  m˙a(w2 w1)

(20.42)

where w1, w2  humidity ratio of moist air entering and lea ving steam injection humidi f er, lb / lb (kg / kg). Let wsm  m˙s / m˙a  w2 w1 ; then h 1 wsmh s  h 2

(20.43)

Assuming that cpd and cps are constants, as wcps  cpd, let w12  (w1 w2) / 2 and replace w1 and w2 by w12. Then T2 T1 

wsmcpsTs cpd w12cps

(20.44)

where Ts  steam temperature, °F (°C) T1, T2  temperature of moist air entering and leaving steam injection humidif er, °F (°C) A heating element humidif er employs a steam coil or electric heating element to pro vide the latent heat of vaporized water. The saturated water vapor added to the moist air is generally at a temperature higher than that of the airstream.

Air Washer An air w asher is a de vice that sprays w ater into air to humidify , to cool and dehumidify , and to clean the air, as discussed in Sec. 15.21. When moist air f ows through an air w asher sprayed with recirculating w ater, as sho wn in Fig. 20.18 b, or w ater is distrib uted o ver a wetted medium, the moist air is humidi f ed, tending to approach saturation. This actual adiabatic saturation process follows the thermodynamic wet-b ulb temperature line on the psychrometric chart represented by line 12 in Fig. 20.18. This process increases the humidity ratio of the airstream while resulting in a reduced air temperature. The cooling ef fect of the adiabatic saturation process is often called direct evaporative cooling and is covered in greater detail in Chap. 27. A saturated end state is not usually achie ved by such a process. The saturation eff ciency sat of the moist air leaving the air washer is def ned by Eq. (15.54) and discussed in Sec. 15.21.

Oversaturation In water atomization for air humidif cation, an atomizing device breaks water into f ne particles that are injected into the airstream. Whether the atomization is accomplished by the centrifugal force,

AIR SYSTEMS: BASICS AND CONSTANT-VOLUME SYSTEMS

20.47

compressed air , or ultrasonic force, the humidi f cation process is an actual adiabatic saturation process. With some atomizing devices, moist air leaving the air washer or pulverizing fan (as shown in Fig. 20.18c) can contain unevaporated water particles ww great enough to exceed the humidity ratio of saturated air at the thermodynamic wet-bulb temperature w *s . The excess amount of water particles present in the moist air is called oversaturation. Oversaturation is def ned as wo w* s  w 2 ww w* s

(20.45)

The quantity of unevaporated water particles at state point 2, in lb / lb (kg / kg) is ww  wo w2

(20.46)

where wo  sum of humidity ratio w2 and minute water particles ww at state point 2, lb / lb (kg / kg) w2  humidity ratio at state point 2, lb / lb (kg / kg) When adiabatic heat transfer occurs between the airstream and the minute w ater particles, some evaporation takes place, so the humidity ratio of the moist air will increase. Such a transformation still follows the thermodynamic wet-bulb temperature line 10, as shown in Fig. 20.18. As the moist air f ows through an air w asher or an atomizing de vice, there can be minute w ater particles present in the moist air , even if the w ater eliminators are emplo yed. The magnitude of ww depends mainly on the construction of the humidifying de vice and the water eliminators, as well as the air v elocity f owing through them. When moist air f ows through an air w asher, ww may vary from 0.0002 to 0.001 lb / lb (kg / kg). If a pulv erizing f an is used, ww may be as high as 0.00135 lb / lb (kg / kg). Oversaturation is bene f cial to a space or to a process where humidi f cation is required. Ho wever, oversaturation causes wetted surf aces and dampness and often f acilitates the gro wth of mold and fungus, resulting in a serious IA Q problem. Necessary remedies must be tak en to eliminate hazardous oversaturation and dampness along the airf ow passage.

20.14 COOLING AND DEHUMIDIFYING PROCESS In a cooling and dehumidifying process, both the temperature and the humidity ratio of the moist air will drop. This process is represented by curv e m-cc on the psychrometric chart in Fig. 20.19, where m is the entering mixture temperature of the outside and recirculating air. There are three types of heat e xchangers commonly used in the cooling and dehumidifying process: 1. Direct-expansion DX coil in which refrigerant e vaporates directly inside the coil ’s tubes, as discussed in Sec. 10.2 2. Water cooling coils with chilled w ater f owing inside the coil’s tubes, as discussed in Sec. 15.10 3. Air washer where chilled w ater rather than recirculated w ater is used for spraying, as discussed in Sec. 15.21 The evaporating temperature of the refrigerant Tev in a DX coil and the outer surf ace temperature of the coil ’s tubes Ts,t are usually lo wer than the de w point of entering air Tae. If the temperature of the entering chilled water in a water cooling coil Twe and the outer surface temperature of the coil’s tubes Ts,t are lower than the de w point of entering air T ae , condensation occurs on the outer surface of the coil tubes on both DX and w ater cooling coils. Temperature, humidity ratio, and enthalpy of moist air are changed during such a cooling and dehumidifying process. Based on the principle of heat balance, Total enthalpy total enthalpy Cooling coil load (or coolHeat energy of condensate of entering air  of leaving air ing capacity of washer)

20.48

CHAPTER TWENTY

FIGURE 20.19

Cooling and dehumidifying process.

That is, 60m˙ah ae  60m˙ah cc Q cc 60m˙wh w where hae, hcc  enthalpy of moist air entering and leaving the cooling coil or air washer, Btu / lb (J / kg) m˙w  mass f ow rate of condensate, lb / min [kg / (60 s)] hw  enthalpy of condensate, Btu / lb (J / kg ) The cooling coil load or the cooling capacity of the air w asher Qcc, in Btu / h (W), can be calculated as Q cc  60m˙a(h ae h cc) 60m˙wh w  60V˙ss(h ae h cc) 60m˙wh w

(20.47)

The mass f ow rate of the condensate can be evaluated as m˙w  m˙a(wae wcc) where wae, wcc  humidity ratio of moist air entering and lea ving coil or air w asher, lb / lb (kg / kg). Assuming that the temperature of the condensate Tw  Tcc, then we see that Eq. (20.47) becomes Q cc  60m˙a[h ae h cc (wae wcc)cpwTcc]  60 V˙s s[h ae h cc (wae wcc)cpwTcc] where cpw  specif c heat of w ater, Btu / lb  °F (J / kg  °C). In most cases,

(20.48a) m˙wh w is less than

AIR SYSTEMS: BASICS AND CONSTANT-VOLUME SYSTEMS

20.49

0.02 Qcc. Because m˙wh w is small compared to m˙a(h ae h cc), for practical calculations, m˙wh w is often neglected, so Q cc  60m˙a(h ae h cc)  60V˙ss(h ae h cc)

(20.48b)

A straight line has been used to represent a cooling and dehumidifying process, and the intersection of this straight line and the saturation curv e is considered the ef fective surface temperature of the coil and is called the apparatus de w point. This is a misconception. First, the cooling and dehumidifying process is actually a curve instead of a straight line. Second, the effective surface temperature as a f ctitious reference point is af fected by many factors and is diff cult to determine in heatand mass-transfer calculations. It is more accurate to def ne the apparatus dew point as the dew point of the moist air leaving the conditioning apparatus, the coil, the air w asher, or other heat e xchangers. Apparatus dew point is represented by point ad on the saturation curve of the psychrometric chart in Fig. 20.19. However, the sensible heat ratio of the cooling and dehumidifying process SHR c can be indicated by the slope of the line joining points m and cc, as shown in Fig. 20.19. SHR c can be calculated as the ratio of sensible heat remo ved during cooling and dehumidifying process Qcs to the

FIGURE 20.20

Cooling and dehumidifying process based on manufacturers’ data.

20.50

CHAPTER TWENTY

total heat removed Qcc, both in Btu / h (W) SHR c 

Q cs Q cc

(20.49)

According to the data published by U.S. manuf acturers, a cooling and dehumidifying curv e can be plotted for the moist air f owing through a w ater cooling coil. Fig. 20.20 sho ws such a curv e. The operating conditions are as follows: Entering air

80°F (26.7°C) dry-bulb and 67°F (19.4°C) wet-bulb temperatures 500 fpm (2.5 m / s) 45°F (7.2°C) 10°F (5.6°C)

Face air velocity Entering water temperature Water temperature rise

The number of ro ws of depth, i.e., the number of coil tubes along the air f ow, varies from 2 to 10 rows, and the f n spacing v aries from 8 to 18 f ns per inch (1.4- to 3.2-mm f n spacing). Both of these factors affect the outer surface area of the coil. Note that the relati ve humidity of the air lea ving the coil approaches saturation as the outer surface area of the coil increases. The cooling and dehumidifying process for a DX coil is similar to that of the w ater cooling coil. Therefore, for coils with f n spacing of less than 10 f ns per inch and for entering air at 80 °F (26.7 °C) dry-b ulb and 68 °F (20 °C) wet-b ulb temperatures, the relative humidity of the conditioned air lea ving the cooling coil may be estimated approximately as follows: 4-row coils 6-row and 8-row coils

90 – 95 percent 96 – 98 percent

Many factors affect the f nal state of the conditioned air leaving the coil. Use manufacturer’s data or refer to Secs. 10.2 and 15.10 for a more detailed analysis.

20.15 ADIABATIC MIXING AND BYPASS MIXING PROCESSES Two-Stream Adiabatic Mixing Process In a tw o-stream adiabatic mixing process, two moist airstreams (say , 1 and 2) are mix ed together adiabatically, forming a uniform mixture m in a mixing chamber . Such a mixing process is illustrated by line 1- m1-2 on the psychrometric chart in Fig. 20.21. In actual practice, for the mixing of either recirculating and outdoor airstreams or the conditioned airstream and bypass airstream, the rate of heat transfer between the mixing chamber and the ambient air is small. It is usually ignored. Based on the principle of heat balance and the continuation of mass m˙1h 1 m˙2h 2  m˙mh m m˙1w1 m˙2w2  m˙mwm

(20.50)

m˙1 m˙2  m˙m In Eqs. (20.50), subscripts 1, 2, and m represent airstream 1, airstream 2, and the mixture, respectively. Because m˙1  m˙m m˙2 and m˙2  m˙m m˙1 , substituting into Eq. (20.50) gives m˙1/m˙m 

h2 hm w2 wm   K1 h2 h1 w2 w1

AIR SYSTEMS: BASICS AND CONSTANT-VOLUME SYSTEMS

FIGURE 20.21

20.51

Adiabatic mixing and bypass mixing processes.

and

m˙2 / m˙m 

hm h1 wm w1   K2 h2 h1 w2 w1

(20.51)

where K1, K2  fraction of airstreams 1 and 2 in the mixture. From the above-mentioned relationships 1. The mixing point m must lie on the line that connects state points 1 and 2. 2. The ratio of the mass f ow rate of airstream 1 to the mass f ow rate of the mixture is e xactly equal to the ratio of segment m1-2 to segment 1-2, as shown in Fig. 20.21, that is, m˙1 m1-2  m˙m 1-2 Similarly, the ratio of the mass f ow rate of airstream 2 to the mass exactly equal to the ratio of segment m1-1 to segment 1-2, that is, m˙2 m1-1  m˙m 1-2

(20.52) f ow rate of the mixture is

(20.53)

As the mass f ow rate m˙  V˙ , substituting into Eqs. (20.50) gives V˙11h 1 V˙22h 2  V˙mmh m V˙11w1 V˙22w2  V˙mmwm V˙11 V˙22  V˙mm

(20.54)

20.52

CHAPTER TWENTY

Because the magnitude of m always lies between 1 and 2, if the differences between m and 1, m and 2, and the differences in specif c heat cpa are negligible, then V˙1h 1 V˙2h 2  V˙mh m V˙1w1 V˙2w2  V˙mwm V˙1T1 V˙2T2  V˙mTm V˙1 V˙2  V˙m

(20.55)

Bypass Mixing Process In man y air -handling units, one airstream is di vided into an upper hot deck air -stream and a lower cold deck airstream, as shown in Fig. 20.21. During the cooling season, the lower airstream is cooled and dehumidi f ed by the cooling coil, and the hot w ater supply to the heating coil in upper hot deck is shut of f. Therefore, the upper airstream becomes a bypass airstream, i.e., it bypasses the cooling coil. After the coils, the conditioned airstream is mixed with the bypass airstream and forms a mixture at state point m2. This bypass mixing process is sho wn by line cc-m2-sf in Fig. 20.21. During the heating season, the cooling coil is not ener gized, but the heating coil is. Under such circumstances, the upper conditioned w arm airstream and the lo wer bypass airstream combine to form a mixture m2, as indicated by line ch-m2-sf in Fig. 20.21. Similar to the two-stream adiabatic mixing process, the ratio of the mass f ow rate of the conditioned stream m˙cc to the mixture m˙m is sf-m2 m˙cc   K cc m˙m sf-cc and

m˙ch sf-m2   K ch m˙m sf-ch

(20.56)

where Kcc, Kch = fraction of conditioning stream f owing through the cooling coil or heating coil. Also m˙m  m˙cc m˙by  m˙ch m˙by m˙mwm  m˙ccwcc m˙by  wsf

(20.57)

Subscript cc represents the condition of f the cooling coil, ch the heating coil, by the bypass airstream, which is the same as the state point of air at the supply f an discharge sf. According to the principle of energy balance hm   wm 

m˙cch cc m˙byh sf m˙m m˙chh ch m˙byh sf m˙m

(20.58)

m˙ccwcc m˙bywsf m˙m

Based on the mass f ow rate of the mixture, the cooling coil load can be calculated as Q cc  K ccVs s(h sf h cc)

(20.59)

AIR SYSTEMS: BASICS AND CONSTANT-VOLUME SYSTEMS

20.53

and the heating coil load is Q ch  K chV˙sscpa(Tch Tsf)

(20.60)

20.16 SINGLE-ZONE, CONSTANT-VOLUME SYSTEMS — COOLING MODE OPERATION Air Conditioning Cycle An air conditioning c ycle consists of se veral air conditioning processes connected in a serial order . Strictly speaking, such processes determine the operating performance of v arious components of the air system in an air conditioning system including the heat, mass, and other energy transfer and the changes in property of the moist air. Different air systems are characterized by dif ferent types of air conditioning c ycles. The psychrometric analysis of an air conditioning c ycle is an important tool to determine the properties of moist air at various state points, the volume f ow rates, and the capacities of the major components of the air system. Air conditioning c ycles can be di vided into tw o categories: open and closed c ycles. In an open air conditioning cycle, the moist air at the end state will not resume its original state. An air conditioning cycle of a 100 percent outdoor air mak eup system is an open c ycle. In a closed air conditioning cycle, moist air at the end state will resume its original state. The cycle of conditioning the mixture of recirculating and outdoor air , supplying it to the space, maintaining it at a required temperature and relative humidity, and recirculating it again is a closed cycle. Based on space load characteristics, the air conditioning cycles for constant-volume single-zone systems operate mainly in two modes: cooling mode or heating mode. 1. Cooling mode . When a single-zone, constant-volume system is turned on and its space temperature Tr is raised above a predetermined limit, such as 75°F (24.4°C), the control system calls for cooling (cold air supply), and the air economizer and refrigeration will be turned on in sequence to maintain Tr within required limits. The constant-volume system is said to be operated at cooling mode. Its equipment is selected to satisfy the summer outdoor and indoor design conditions. 2. Heating mode. When a single-zone, constant-volume system is turned on and its Tr drops below a predetermined limit, such as 70°F (21.1°C), the control system calls for heating (warm air supply) to maintain a required Tr within predetermined limits. The constant-volume system is said to be operated at heating mode. The same equipment is also selected to satisfy winter outdoor and indoor design conditions.

Cooling Mode Operation in Summer In Fig. 20.22, the upper air conditioning c ycle is the summer cooling mode c ycle for a constantvolume single-zone system in a f actory. Return air from the conditioned space may pass through a return f an. Recirculating air enters the air -handling unit or packaged unit at state point ru and is mixed with the outdoor air at point o in the mixing box. The mixture m is then cooled and dehumidif ed at the cooling coil and lea ves it at point cc. After that, the conditioned air f ows through the supply fan sf, the supply duct, and then supplies to the conditioned space through the supply outlet at state point s. When supply air absorbs the sensible and latent loads from the space, it becomes space air at point r. Space air then f ows through ceiling plenum rp, absorbs heat released from the light f xtures and the roof, and f ows through the return duct / return fan rf. The recirculated portion enters the AHU or PU again at point ru.

FIGURE 20.22

20.54

Summer cooling mode air conditioning c ycle for a constant-volume single-zone system in a f actory.

AIR SYSTEMS: BASICS AND CONSTANT-VOLUME SYSTEMS

20.55

Cooling mode operation in summer consists of the following processes: 1. Sensible heating process r-ru from the return system heat gain qr,s, in Btu / h (W), when return air f ows through the ceiling plenum, the return duct, and the return fan. It includes ●

●

●

Heat gain from the electric lights recessed on the ceiling plenum qrp Return fan power heat gain qrf, if there is a return fan Return duct heat gain qrd

That is; qr,s  qrp qrf qrd

(20.61)

It is often con venient to use o verall temperature rise to denote the system heat gain, such as, Tr,s, in °F (°C). Then the return system heat gain is qr,s  60V˙retretcpa(Tru Tr)  60V˙retretcpa Tr,s

(20.62)

Subscript ret indicates return air and ru the recirculating air entering the AHU or PU. The volume f ow rate of return air is often higher than that of the recirculated air entering the mixing box if there is leakage or exhaust air through the exhaust / relief damper. 2. Adiabatic mixing process of recirculating air at state point ru and the outdoor air at state point o in the mixing box, which is represented by line ru-m-o on the psychrometric chart in Fig. 20.22.

FIGURE 20.23

Air conditioing c ycle — winter cooling mode operation for a constant-v olume single-zone system.

20.56

CHAPTER TWENTY

After mixing, the condition of the mixture is m, which is also the state point of the moist air entering the cooling coil. 3. Cooling and dehumidifying process, which is indicated by the curv e m-cc. The cooling coil load can be calculated by Eq. (20.48b). 4. Sensible heating process cc-s from the supply system heat gain qs,s when supply air f ows through the supply fan and supply duct. It includes ●

●

Supply fan power heat gain qsf represented by line cc-sf Supply duct heat gain qsd indicated by line sf-s.

Supply system heat gain is then given as qs,s  qsf qsd

(20.63)

The supply fan power heat gain and the supply duct heat gain are f ar greater than the heat gains of the return fan and duct. These differences are caused by the higher f an total pressure in the supply fan and the larger temperature difference between the air inside and outside the supply duct. Like the return system heat gain the supply system heat gain can be represented by an o verall temperature rise Ts,s, in °F (°C). 5. Space conditioning process, which is represented by line sr. From Eqs. (20.31) and (20.36), the sensible heat removed from the conditioned space or the space sensible load Qrs, in Btu / h (W), is Q rs  60m˙acpa(Tr Ts)  60V˙sscpa(Tr Ts)

(20.64)

From Eq. (20.32), the latent heat removed from the conditioned space or the space latent load Qrl, in Btu / h (W), is Q rl  60m˙a(wr ws)h fg,32  60V˙ss(wr ws)h fg,32

(20.65)

Therefore, the space cooling load Qrc, in Btu / h (W), is calculated as Qrc  Qrs Qrl

(20.66)

Cooling Mode Operation in Winter Without Space Humidity Control In winter, many factories that employ constant-volume single-zone systems still ha ve a space cooling load, need a cold air supply , and operate at cooling mode operation. These systems can be divided into two categories: systems without space humidity control, as shown by the lower air conditioning cycle in Fig. 20.23, and systems with space humidity control, as shown by the upper air conditioning cycle in Fig. 20.23. The air conditioning c ycle of a constant-v olume single-zone system without space humidity control operated at cooling mode winter design conditions consists of the following air conditioning processes, and most are similar to those in summer design conditions: ●

●

●

●

Sensible heating process r-ru from the return system heat gain qr,s Adiabatic mixing process of recirculating air ru and the outdoor air o in the mixing box Sensible heating process m-sf-s from the supply system heat gain qs,s Space conditioning process sr

Most of the constant-v olume single-zone systems operated at cooling mode winter design conditions do not need a cooling and dehumidifying process because the mixture of outdoor and

AIR SYSTEMS: BASICS AND CONSTANT-VOLUME SYSTEMS

20.57

recirculating air m is usually lo wer than the supply temperature plus the temperature rise due to supply system heat gain. This winter cooling mode free cooling c ycle can be constructed as follows: 1. From Eq. (20.65), calculate the humidity ratio dif ference wr ws from the kno wn space latent load and supply volume f ow rate. 2. The volume f ow rate of outdoor air V˙o should be greater than the occupant requirements, and suff cient to offset the space cooling load during winter without refrigeration. Based on the same supply volume f ow rate as in summer , use Eq. (20.31) to calculate the supply air temperature difference Tr Ts from the design winter sensible cooling load, and determine Ts. 3. Temperature of the recirculating air entering the AHU or PU Tru can be calculated as Tru  Tr temperature rise due to return system heat gain Mixing temperature of outdoor and recirculating air can be calculated as Tm  Ts temperature rise due to supply system heat gain 4. Because wr  wru and wm  ws, the ratio wr ws wru wm Tru Tm   wru wo wr wo Tru To

(20.67)

From Eq. (20.67), wr wo and wr can be determined.

Cooling Mode Operation In Winter with Space Humidity Control Constant-volume single-zone systems using cold air supply with space humidity control during the winter season ha ve been adopted in man y industrial applications, such as spinning and wea ving rooms in te xtile mills. In these w orkshops, high internal sensible heat loads are released by the spinning machines and wea ving looms, requiring cold air supply and a speci f c controlled relati ve humidity, even in winter. The upper air conditioning c ycle of Fig. 20.23 sho ws such a cooling mode c ycle in winter with space humidity control using a spray-type air w asher for a constant-v olume single-zone system. The adiabatic saturation process of e vaporative cooling humidi f cation in an air w asher saves energy to e vaporate the liquid w ater which is especially suitable for conditioned space that needs a cold air supply in winter . Air washers also cool and dehumidify the air in summer by chilled w ater and clean the air by water spraying. The following are the air conditioning processes in this cycle and their characteristics: 1. Process m-sc is an adiabatic saturation process. It proceeds along the thermodynamic wet-b ulb temperature line on the psychrometric chart. 2. State point sc indicates the conditioned air leaving the air washer. Process lines sc-sf, sf-s, and s-r all undergo heating and humidifying processes. Because the airstream lea ving the air w asher is oversaturated with minute water particles, as the airstream picks heat from the fan, ductwall, and space machine loads, more evaporation occurs. The increase in the humidity ratios in each of these processes varies from 0.0001 to 0.0003 lb / lb (kg / kg). These processes can be determined from both the increase of the enthalpy and the humidity ratios. 3. Because of the high machine load, the sensible heat ratio of the space conditioning line SHR s often exceeds 0.90. 4. Outdoor air is used for free cooling as well as for occupants and the dilution of air contaminants.

20.58

CHAPTER TWENTY

Part-Load Operation and Controls In a constant-volume single-zone system operated at cooling mode, when the space cooling load reduces, two types of control are used to maintain the required space temperature or other controlled variable during part-load operation: Water Flow Rate Modulation Control. For an air system emplo ying a w ater cooling coil in an air-handling unit, the reduction of space sensible cooling load causes a drop in space temperature Tr. When a drop of Tr is sensed, the controller closes the tw o-way valve of the chilled w ater supply to maintain the required space temperature. The reduction of chilled w ater f owing through the coil raises the air temperature and humidity ratio of f the coil. The result is a higher space relati ve humidity rp at part-load operation than at design load when the space temperature is maintained at a nearly constant value. The part-load performance of a w ater cooling coil, as calculated in Example 15.2, is sho wn below:

Load ratio Entering air Leaving air Water velocity Water temperature rise Space temperature Space relative humidity

Design load

Part load

1.0 80°F dry / 67°F wet 57.5°F dry / 56.5°F wet 4 fps 10°F (5.6°C) 75°F (23.9°C) 55 percent

0.8 79°F dry / 68°F wet 61.3°F dry / 59.5°F wet 2.5 fps 16°F (8.9°C) 75°F (23.9°C) 59 percent

If the load ratio drops to 0.5, then space relati ve humidity at part load rp may increase to 65 percent. Two-Position or Cycling Control. For small constant-volume single-zone systems, such as using a packaged unit whose supply volume f ow rate is 1500 cfm (708 L / s) or less, on / off control of the refrigeration compressors or the packaged unit is often used to adjust the cooling capacity of the system within an operating c ycle of a fe w minutes to 20 min. Both the air temperature of f the coil and the space temperature f uctuate. Because of the heat storage capacity of the air system and the building structure in the conditioned space, the space temperature may f uctuate by 1 to 2°F (0.56 to 1.1°C). For large packaged units, the fan is continuously operating. Only the refrigeration compressors are cycling in groups, so that the cooling capacity is reduced at part-load operation. At the same time, the amplitude of the f uctuation of the space temperature becomes smaller accordingly.

Outdoor Ventilation Air and Exhaust Fans Because the supply v olume f ow is constant year -round for a constant-v olume single-zone system only when the f lter is loaded with dirt, the fan total pressure increases and the supply v olume f ow rate decreases accordingly. Manually Adjusted Dampers. Many small constant-v olume single-zone systems for residential and retail stores using a rooftop or split packaged unit with a supply v olume f ow rate of 1500 cfm (708 L / s) or less employ manually adjusted outdoor air dampers and operate as follows: ●

Outdoor damper is f eld-adjustable to allo w preset amount of outdoor air for year -round ventilation purposes.

AIR SYSTEMS: BASICS AND CONSTANT-VOLUME SYSTEMS ●

●

20.59

Outdoor damper acti vates with supply f an to preset position and closes when supply f an is shut off. Outdoor damper can be preset to admit up to 50 percent of outdoor air.

Air Economizer . For constant-v olume systems with a supply v olume f ow rate of 2000 cfm (944 L / s) or greater , many packaged unit manuf acturers offer air economizers to use outdoor air free cooling for ener gy sa vings as well as to impro ve indoor air quality . An air economizer is always used as the f rst stage of cooling before refrigeration is ener gized. Air economizers are discussed in Chap. 21. Exhaust Systems. In most small constant-v olume single-zone systems using a PU, part of space air is exf ltrated through the cracks and gaps on the b uilding shell and the opening of entrance doors because of the positive space pressure to balance the amount of outdoor air intak e at the PU. Sometimes an additional small exhaust fan is installed on the external window or wall of the restroom. For a constant-volume system operated in an air economizer cycle that extracts 100 percent outdoor air for free cooling, a return or relief fan must be used to exhaust the space air outdoors.

20.17 SUPPLY VOLUME FLOW RATE Supply air volume f ow rate and the refrigeration capacity are two primary air system characteristics.

Based on Space Cooling and Heating Load For most air systems including constant-v olume systems that are designed to maintain a healthy and comfortable space temperature Tr, the supply air v olume f ow rate is often calculated from the space cooling load. From Eqs. (20.31) and (20.36), the mass f ow rate of the supply air , in lb / min (g / min), can be calculated as m˙a 

Q rs Q rc  60(h r h s) 60cpa(Tr Ts)

(20.68)

The supply volume f ow rate, in cfm [m3 /(60 s)], can be calculated as V˙s 

Q rc

Q rs

(20.69) 60s(h r h s) 60scpa(Tr Ts) In Eq. (20.69), Tr is the space temperature and Ts the supply temperature at the outlet, both in °F (°C). F or a constant-v olume single-zone system operated at cooling mode summer design conditions, Tr is often the required indoor design temperature in summer , and Ts is the representati ve supply temperature of this single zone, as shown in Fig. 20.22. For a multizone air system for comfort air conditioning, Tr is often the same summer indoor design temperature, and Ts is the supply temperature of the control zone that has the greatest supply system heat gain. In Eq. (20.69), Qrs indicates the design sensible cooling load, in Btu / h (W). F or a constant v olume system Qrs is the peak cooling load of that specif c area which the air system serves because V˙s is not reduced at partload operation. For a VAV system, Qrs is the block cooling load. If Qrs and Tr and Ts are properly selected, Eq. (20.68) can be applied to both single-zone and multizone as well as constant-volume and variable-air-volume air systems. The supply density V˙s,d , in cfm / ft2 (L / s  m2) is def ned as V˙s,d 



V˙s Afl

20.60

CHAPTER TWENTY

where A f  f oor area of the conditioned space, ft2 (m2). Supply air density is often used in air volume f ow estimates. The following items should be considered during the calculation of the supply air f ow rate from Eqs. (20.68) and (20.69): 1. Volumetric versus mass f ow rate. The supply volume f ow rate V˙s is widely used in calculations to determine the size of f ans, grilles, outlets, air-handling units, and packaged units. However, mass f ow rate m˙a is sometimes simpler to apply in cooling coil load calculations, or it may be more appropriate when the variation of air density affects the result. 2. Temperature difference versus enthalpy difference. Theoretically, in Eq. (20.69), it is more accurate to use the enthalp y difference to calculate the supply v olume f ow rate. This is due to the elimination of the term cpa, which is actually a variable depending on the temperature and humidity ratio. In actual practice, it is more convenient to use temperature differences instead of enthalpy difference because usually the space temperature is controlled and maintained at a preset v alue. The variation of cpa is often not signif cant enough to affect the calculated results. 3. Summer cooling load ver sus winter heating load . Summer cooling load is usually used to calculate the supply v olume f ow rate because it is generally greater than using the winter heating load. But if the winter weather is v ery cold, this may require a greater v olume f ow rate. Based on winter heating load, the supply volume f ow rate can be calculated as V˙s 

Q rh

(20.70)

scpa(Ts Tr) where Qrh  space heating load, Btu / h (W). Use the greater v olume f ow rate of the tw o to select the fans, AHU or PU, and other components.

Based on Requirements Other Than Cooling Load The supply volume f ow rate of many air conditioning systems can also be go verned by the following requirements: 1. To dilute the concentration of the air contaminants . Based on the principle of conserv ation of mass, the mass generated rate of the air contaminants in the conditioned space m˙par , is given by m˙par 

V˙s(Ci Cs ) 2118

And the supply volume f ow rate V˙s, in cfm (m3 /s), can be calculated as V˙s 

2118 m˙par Ci Cs

(20.71)

where Ci, Cs  concentrations of air contaminants of space air and supply air, usually in mg / m3; 1 mg  0.01543 gr m˙par  rate of contaminants generated in space, mg / s 2. To maintain a r equired relative humidity r as well as a humidity r atio w r at a speci f c temperature for the conditioned space. Then the supply volume f ow rate V˙s, in cfm [m 3 /(60 s)], can be calculated as V˙s 

Q rl 60s(wr ws)h fg,32

(20.72)

AIR SYSTEMS: BASICS AND CONSTANT-VOLUME SYSTEMS

20.61

3. To provide a desirable air velocity v r , in fpm (m / s), within the working area in a clean room. The supply volume f ow rate V˙s, in cfm (m3 /s), is then V˙s  Arvr

(20.73)

where Ar  cross-sectional area perpendicular to airf ow in working area, ft (m ). 4. To exceed the outdoor air r equirement for acceptable air quality for occupants . The supply volume f ow rate can then be calculated as 2

V˙s  nV˙oc

2

(20.74)

where n  number of occupants V˙oc  outdoor air requirement per person, cfm (m3 /s) 5. To exceed the sum of the volume f ow rates of the exhaust air V˙ex and the space air r elief V˙exf through e xf ltration to maintain a positi ve pressure in the conditioned space, both in cfm (m 3 /s). Then the supply volume f ow rate is V˙s  V˙ex V˙exf

(20.75)

The supply v olume f ow rate should be the one that is the greatest from an y of the abo ve requirements, the space cooling load, or heating load, or outdoor air requirements.

Rated Volume Flow of Supply and Return Fans The supply volume f ow rate V˙s is determined based on the mass f ow rate of supply air m˙a required to offset the space load at the speci f c supply condition, with a speci f c supply air density s. Because the supply volume f ow rate must be provided by the supply fan, or the fan’s rated volumetric capacity ,V˙sf r against a certain f an total pressure pt, it is important to determine the relationship between and V˙s V˙sf,r . If the supply f an is located do wnstream from the cooling coil, the humidity ratios of the moist air between state points cc and s are constant, as shown in Fig. 20.22. According to the principle of conservation of mass, the mass f ow rate of supply air m˙a, in lb / min (kg / min), remains constant, but the volume f ow changes as its state changes. The relationships are m˙a  V˙cccc  V˙sfsf  V˙ss

(20.76)

Subscript cc represents the state point of conditioned air lea ving the cooling coil, sf represents the moist air at the outlet of the supply fan, and s represents the supply air at the outlets. Fans are rated at standard air condition. Standard air means dry air at a temperature of 70 °F (21.1°C) and a barometric pressure of 29.92 in. (760 mm) Hg (14.697 psi) with a density of 0.075 lb / ft3 (1.20 kg / m3). From Eqs. (20.76) and (20.68), the volume f ow rate at the supply f an outlet V˙sf , in cfm (m3 /s), is calculated as V˙sf 

V˙ss

sf



Q rs 60sf cpa (Tr Ts)

(20.77)

For the same fan at a given speed, its volume f ow rate remains unchanged against a f xed pt even if the air density varies during rated conditions. Therefore, V˙sf,r  V˙sf

(20.78)

where V˙sf,r  rated v olume f ow rate, cfm (m /s). From Eq. (2.32), the moist v olume of 1 lb of moist air v is 3

v

R aTR(1 1.6078w) pat

20.62

CHAPTER TWENTY

For normal summer and winter comfort applications, the humidity ratio varies less than 0.006 lb / lb (kg / kg). Therefore, v  1.01RaTR / pat, that is, the inf uence of variation of humidity ratio is smaller than 1 percent. Therefore, it can be ignored. If there is no in f ltration and e xf ltration at the conditioned space, Eq. (20.76) can be rewritten as m˙a  V˙cccc  V˙sfsf  V˙ss  V˙rr  V˙rfrf

(20.79)

Subscript rf represents the outlet of the return fan. Similarly, the rated volume f ow rate of the return fan V˙rf,r is given by V˙rf,r  V˙rf 

Q rs 60 rf cpa(Tr Ts)

(20.80)

For air temperature leaving the cooling coil Tcc at 55°F (12.8°C) with a relative humidity of 92 percent and Tsf at 57 °F (13.9 °C), the psychrometric chart sho ws the moist v olume vsf = 13.20 ft 3 / lb (0.824 m3 / kg). From Eq. (20.77), the rated volume f ow rate of supply fan is V˙sf 

Q rs 13.20 Q rs  60  0.243(Tr Ts) 1.105(Tr Ts)

(20.81)

For cold air distribution, which is covered in detail in Chap. 18, Tcc may equal 40°F (4.4°C) with a relative humidity of 98 percent; and if Tsf  42°F (5.6°C), then, vsf  12.8 ft 3 / lb (0.8 m 3 / kg), and the rated supply volume f ow is V˙sf,r  V˙sf 

Q rs 12.8 Q rs  60  0.243(Tr Ts) 1.14(Tr Ts)

(20.82)

For an air system in which the supply f an is upstream from the cooling coil, and if Tsf  82°F (27.8°C) and sf  43 percent, then, from the psychrometric chart, vsf  13.87 ft 3 / lb (0.866 m3 /kg), and the supply volume f ow rate is V˙sf,r  V˙sf 

13.87 Q rs Q rs  60  0.243(Tr Ts) 1.05(Tr Ts)

(20.83)

Thus, an upstream f an requires 9 percent greater f an capacity than a do wnstream fan with cold air distribution even though Qrs and Tr Ts are the same.

20.18 DETERMINATION OF THE SUPPLY AIR CONDITION The following parametrs are usually kno wn before the constant-v olume system cooling mode air conditioning cycle is developed: 1. 2. 3. 4. 5. 6.

Summer and winter outdoor design temperatures: dry-bulb To and wet-bulb To Summer and winter indoor space air temperature Tr and relative humidity r Space sensible and latent load Qrs and Qrl Outdoor air requirement V˙o Estimated supply system heat gain qs,s Estimated return system heat gain qr,s

If the state point of the supply air can be determined from the psychrometric chart, then the constant-volume system cooling mode air conditioning c ycle can be de veloped. Eventually, the supply volume f ow rate V˙s and the cooling coil load Qcc can also be determined.

AIR SYSTEMS: BASICS AND CONSTANT-VOLUME SYSTEMS

20.63

Air Conditioning Rules To determine the supply air condition, the following air conditioning rules are helpful: 1. For a gi ven summer indoor design temperature Tr, design space sensible cooling load Qrs, and specif c supply system heat gain qs,s, ●

●

A lower air of f-coil temperature Tcc and supply air temperature Ts always result in a greater supply air temperature differential Ts  Tr Ts and a lower space relative humidity s. On the other hand, a higher Tcc means a smaller Ts and a higher r .

2. A greater Ts reduces the volume f ow rate and, therefore, fan size, fan energy use, and terminal and duct sizes. These reduce both f rst cost and operating costs. At the same time, a greater Ts, needs a lo wer Tcc to maintain an indoor design temperature Tr . Excluding the cold air distrib ution in the ice storage system, a Ts between 15 and 20°F (8.3 and 11.1°C) is most widely used. 3. Temperature of air leaving the cooling coil Tcc is closely related to the temperature of the chilled water entering and leaving the coil, Twe and Twl, or to the evaporating temperature inside the DX coil Tev. A lower Tcc or Twe needs a correspondingly lower Tev and, therefore, higher power input to the refrigeration compressors. 4. For a given Tcc, Qrs, and qs,s, a supply volume f ow rate V˙s that is greater than the required v alue results in a lower space temperature Tr, a higher space relative humidity r, and a smaller Ts. On the other hand, a smaller than required V˙s results in a higher Tr, a lower r , and a greater Ts. 5. When an air system is used to serv e a single-zone conditioned space, Ts and Tcc should be properly arranged so that the indoor design temperature and relati ve humidity can be maintained for energy-eff cient operation. 6. For an air system serving multizones, with many space temperatures and relati ve humidity requirements, Twe and Tcc or Tev should satisfy the lo west Tcc requirement. In actual practice, Twe and Tcc, or Tev, are often determined by previous experience with similar applications.

Graphical Method The following graphical method may be used to determine the condition of supply air s for an air system serving a single zone (including constant-v olume single-zone systems) whose Tr and r must be maintained. 1. According to the data gi ven in Sec. 20.14, the relative humidity of the air lea ving a cooling coil cc with 10 or more f ns per inch can be estimated as 4-row coil:

cc  93 percent

6-row coil:

cc  96 percent

8-row coil:

cc  98 percent

2. Draw the space conditioning line sr from the state point r, parallel to the line of kno wn sensible heat ratio SHRs, as shown in Fig. 20.24. 3. State point of supply air s must lie on line sr. 4. The horizontal line cc-s represents the magnitude of the supply system heat gain. As segment cc-s is mo ved up and do wn, there e xists a position where cc lies e xactly on the selected cc curve for the coil, s lies on line sr, and Ts Tcc exactly matches the temperature increase caused by supply system heat gain. 5. In practice, this required Tcc must be reconciled with the selected coil ’s performance for a gi ven coil conf guration, Twe, and air and water velocities.

20.64

CHAPTER TWENTY

FIGURE 20.24

Determination of state point of supply air.

Influence of Sensible Heat Ratio For a specif c design Tr and r, the sensible heat ratio of the space conditioning line SHR s has a signif cant effect on Ts and Tcc. Fig. 20.25 shows such an in f uence. As SHRs becomes smaller, for the sake of maintaining the required Tr and r, the temperature of conditioned air lea ving the cooling coil Tcc must be lowered. This requires a lo wer Twe or a lower evaporating temperature Tev in a DX coil, which eventually increases the ener gy input to the refrigeration compressors, but reduces the air volume f ow that must be handled. For a specif c space cooling load Qrc, a lower SHR s always indicates a smaller Qrs and a greater Ts. A higher Ts means a lower supply volume f ow rate V˙s, smaller equipment size, lower fan energy use, and reduced construction and operating costs. In Fig. 20.25, when SHR s  0.7, the dew point of space air of 75 °F (23.9°C), and 50 percent relative humidity is 55 °F (12.8°C), which is higher than the 54 °F (12.2°C) supply air temperature. Under such circumstances, condensation may occur at the supply outlet when its cold metal frame contacts the space air at dew point 55°F (12.8°C) or higher. Condensation can be avoided by the following measures: ●

●

Add an insulation layer with v apor barrier , on the surf ace of ducts and outlets with suf f cient thickness, typically 1 to 2 in. (25 to 50 mm) that are e xposed to space air or ambient air whose dew point is higher than the external temperature of the exposed surface. Adopt supply outlet that can induce suf f cient space air to mix with lo w-temperature supply air in order to raise Ts.

For an air system to serv e a single-zone conditioned space (including constant-v olume single-zone systems) with specif c indoor Tr and r , say, 75°F (23.9°C) and 50 percent, and a f xed SHRs, there will be only one supply air condition with a corresponding Ts that can provide the required indoor Tr and r (see Fig. 20.26). If Tr remains constant at 75 °F (23.9°C), an increase of Ts due to the

FIGURE 20.25 of supply air.

Effect of sensible heat ratio of space conditioning line on the determination of the condition

FIGURE 20.26

Inf uence of supply temperature difference Tr Ts on Tcc and r .

20.65

20.66

CHAPTER TWENTY

reduction of supply volume f ow rate V˙s at a f xed space sensible load Qrs is possible only when the temperature of air lea ving the cooling coil Tcc and supply temperature Ts are lowered accordingly. The space relative humidity r also will be lo wer. This again needs a lo wer chilled water temperature entering the coil Twe or a lower evaporating temperature inside the DX coil Tev. Example 20.4. One w orkshop in a f actory emplo ys a constant-v olume single-zone air system. The summer space sensible cooling load is 350,000 Btu / h (102,550 W), and the summer space latent load is 62,000 Btu / h (18,166 W). Other design data are as follows: Summer outdoor design temperatures: Dry-bulb Wet-bulb Summer indoor space conditions: Air temperature Relative humidity Temperature rise: In supply fan In supply duct Because of heat released from light f xtures and roof In return duct and return fan Relative humidity of air leaving cooling coil Outdoor air requirement

95°F (35°C) 75°F (23.9°C) 75°F (23.9°C) 50 percent 2°F (1.11°C) 3°F (1.67°C) 3°F (1.67°C) 2°F (1.11°C) 93 percent 3200 cfm (1510 L / s)

Determine 1. The condition of supply air at summer design temperatures 2. The rated volume f ow rate of supply fan 3. The cooling coil load Solution 1. From Eq. (20.30) and the given data, the sensible heat ratio of the space conditioning line can be calculated as SHR s 

| Q rs | 350,000  | qrs | | qrl | 350,000 62,000

 0.85 On the psychrometric chart in the upper summer cool mode air conditioning c ycle in Fig. 20.22, draw a space conditioning line sr from point r with SHR s  0.85. From the gi ven data, the total temperature rise of the supply system heat gain is 2 3  5°F. Point cc will lie on cc 93 percent line, and point s must lie on line sr. Let se gment cc-s  5°F. M ove cc-s until point cc is on the 93 percent curve and point s is on line sr. The state point of supply air s and the condition of air leaving the cooling coil, point cc, are then determined. From the psychrometric chart, Ts  60°F (15.6°C)

s  78 percent

Tcc  55°F (12.8°C)

cc  93 percent

and

ws  0.0086 lb / lb (kg / kg) wcc  0.0086 lb / lb (kg / kg)

hcc  22.6 Btu / lb (52.6 kJ / kg)

AIR SYSTEMS: BASICS AND CONSTANT-VOLUME SYSTEMS

20.67

2. Because the air temperature at the supply f an outlet Tsf  55 2  57°F, and wsf  wcc  0.0086 lb / lb, from the psychrometric chart, moist volume vsf  13.21 ft3 / lb. From Eqs. (20.77) and (20.78) the rated supply volume f ow rate of the supply fan is V˙sf,r  V˙sf 



Q rs 60sf cpa (Tr Ts)

13.21  350,000  21,140 cfm (9976 L /s) 60  0.243(75-60)

3. State point m can be graphically determined or calculated. The total temperature rise due to the return system heat gain is 3 2  5°F; therefore, Trf  Tru  75 5  80°F. Because line r-ru is a horizontal line, point ru can be plotted on the psychrometric chart. From the given data, the summer outdoor temperature To  95°F, and the summer design outdoor wet-b ulb temperature To  75°F, so the state point of outdoor air o can also be plotted. Connect line ru-o. Ignore the difference in density between points ru, m and o, Then ru-m 3200   0.15 ru-o 21,140 From the psychrometric chart, the length of line section ru-o is 1.8 in.; then point m can be determined as Tm  82°F (27.8°C)

m  44 percent

and

hm  30.9 Btu / lb (71.9 kJ / kg)

Mixing point m can also be determined by using the analytical method. From the psychrometric chart, the enthalp y of outdoor air ho  38.4 Btu / lb and the enthalp y hru  29.8 Btu / lb. From Eq. (20.55), hm  

V˙oh o (V˙s V˙o)h ru V˙m 3200  38.4 (21,140 3200)29.8 21,140

 31.10 Btu/lb (72.3 kJ / kg) Tm  

V˙oTo (V˙s V˙o)Tru V˙m 3200  95 (21,140 3200)80  82.3F(27.95C) 21,140

The differences between the enthalpies and temperatures determined from the graphical and analytical methods are small, less than 1 percent. But the analytical method is more accurate because the line segment lengths are small. Because V˙ss  V˙sfsf , from Eq. (20.48 b), the cooling coil load is calculated as Q cc  60V˙sf sf (h m h cc)  60 

21,140(31.1 22.6)  816,154 Btu / h (239,133 W or 239 kW) 13.21

Example 20.5. A f actory uses a constant-v olume single-zone system without space humidity control. The operating parameters at winter design conditions are as follows:

20.68

CHAPTER TWENTY

Winter outdoor design conditions Air temperature Relative humidity Winter indoor design conditions Air temperature Space sensible cooling load Space latent load Temperature rise due to supply system heat gain Temperature rise due to return system heat gain Supply volume f ow rate Minimum outdoor air requirement

32°F (0°C) 30 percent 72°F (22.2°C) 800,000 Btu / h (234,400 W) 600,000 Btu / h (175,800 W) 4°F (2.22°C) 2°F (1.11°C) 100,000 cfm (47,190 L / s) 26,250 cfm (12,387 L / s)

At winter design condition, a mixture of outdoor air and recirculating air forms the cold air supply (cooling mode operation) to offset the space cooling load. Determine the space relative humidity at winter design conditions and the actual amount of outdoor air intake. Solution. Assume that the moist v olume of the supply air vs  13.3 ft 3 / lb. From Eq. (20.65) and the given data, the humidity ratio difference between the space and supply air can be calculated as Q rl wr ws  ˙ 60Vssh fg,32 600,000  13.3  0.00125 lb / lb (kg / kg) 60  100,000  1075 From Eq. (20.31) and the gi ven data, the supply air temperature dif ference Tr Ts can be calculated as Q rs Tr Ts  60V˙sscpa 



800,000  13.3  7.30 F (4.1C) 60  100,000  0.243

From the given data Tru  Tr 2  72 2  74°F and

Tm  Tr (Tr Ts) 4  72 7.3 4  60.7°F (15.9°C)

Because wru  wr and wm  ws, from Eq. (20.67) wr ws Tru Tm  wru wo Tru To  Then

wr wo 

74 60.7  0.317 74 32

wr ws  0.00394 lb / lb (kg / kg) 0.317

From the psychrometric chart, at a temperature of 32°F and a relative humidity of 30 percent, wo  0.0012 lb / lb, so wr  wo 0.00394  0.0012 0.00394  0.00514 lb / lb

AIR SYSTEMS: BASICS AND CONSTANT-VOLUME SYSTEMS

20.69

At a temperature of 72 °F (22.2 °C) and wr  0.00514 lb / lb (kg / kg), the space relati ve humidity r  31 percent, as shown in the lower part of the psychrometric chart in Fig. 20.23. The amount of outdoor air intake can be calculated as V˙o 

V˙s(Tru Trn) Tru To

 100,000  0.317  31,700 cfm (14,959 L / s) This amount is greater than the amount of outdoor air required 26,250 cfm (12,387 L / s). At Ts  72 7.3  64.7°F and ws  0.00514 0.00125  0.00389 lb / lb, from the psychrometric chart, vs  13.3 ft3/ lb (0.830 m3/ kg), which is equal to the assumed value.

20.19 CONSTANT-VOLUME SINGLE-ZONE SYSTEMS — HEATING MODE OPERATION Heating Mode without Space Humidity Control In many residential and small commercial b uildings, the sum of the heat loss due to transmission through the windows, walls, and roofs and the heat loss due to in f ltration in winter is greater than the sum of the internal heat gains from occupants, electric lights, equipment, and appliances. Then a warm air supply is required to maintain a desirable indoor temperature. A constant-volume singlezone system operated in the heating mode without space humidity is often used. The lower air conditioning c ycle in Fig. 20.22 sho ws such a heating mode c ycle in winter. Return air at state point r f ows through the return grilles, return duct, and return fan. After that, recirculating air at point ru enters the air-handling unit and mix es with the outdoor air . The mixture, at point m, passes through the cooling coil which is not ener gized, and is then heated at the heating coil. Warm air leaves the heating coil at point ch. It leaves the supply fan at point sf after absorbing the fan power heat gain. The warm supply air loses heat in the supply duct and supplies to the conditioned space at point s. At the conditioned space, the supply air absorbs the latent heat from the occupants; supplies heat to compensate for transmission loss through e xternal windows, walls, and roofs; and f nally changes to the point r. The air conditioning c ycle in heating mode operation without space humidity control during winter design condition (as shown in Fig. 20.22) consists of the following processes: 1. Sensible heating process r-ru due to the return system heat gain qr,s. In winter, the return system heat gain consists of mainly the heat released from the electric lights qrp and the return f an power heat gain qrf, if any. That is, qr,s  qrp qrf

(20.84)

The temperature difference between the air inside and outside the return duct is generally small in winter and can be ignored. 2. Adiabatic mixing process ru-m-o. The ratio of the minimum v olume f ow rate of outdoor air to the supply v olume f ow rate is the same as in the cooling mode in summer design condition in order to provide the required amount of outdoor air for occupants. 3. Sensible heating process m-ch at the heating coil. The heating coil load can be calculated according to Eq. (20.39) as Q ch  60 V˙sscpa(Tch Tm) where Tm, Tch  temperature of mixture entering coil and air off heating coil,°F (°C). 4. Sensible heating process ch-sf due to the supply fan power heat gain qsf.

(20.85)

20.70

CHAPTER TWENTY

5. Sensible cooling process sf-s due to the heat loss from the w arm air inside the supply duct through the duct wall to the ambient air qsd. 6. Space conditioning process sr. During heating mode operation in winter , heat is supplied to the space to compensate for the transmission loss, and moisture is absorbed by the supply air due to the space latent load. Supply air at point s is then changed to r. The heat supplied to the conditioned space to of fset the space heating load Qrh, in Btu / h (W), can be calculated as Q rh  60 V˙sscpa(Ts Tr)

(20.86)

And the latent heat absorbed by the supply air Qrl, in Btu / h (W), is Q rl  60V˙ss(wr ws)h fg,32

(20.87)

Example 20.6. For the same constant-volume single-zone system as that in Example 20.4, shown in Fig. 20.22, the following winter design conditions are given: Winter outdoor design conditions: Dry-bulb temperature Relative humidity Winter indoor design temperature Outdoor air requirement Temperature rise: Due to heat release from light f xtures 2 Return fan power heat gain Supply fan Temperature drop of the supply duct Space heating load Space latent load Supply volume f ow rate

15°F (9.4°C) 30 percent 72°F (22.2°C) 3200 cfm (1510 L / s) °F (1.11°C) 1°F (0.56°C) 2°F (1.11°C) 1°F (0.56°C) 225,000 Btu / h (65,925 W) 60,000 Btu / h (17,580 W) 21,140 cfm (9976 L / s)

Determine 1. The space relative humidity at winter design conditions 2. The temperature of supply air 3. The heating coil load Solution 1. To develop a winter heating mode air conditioning c ycle for a constant-v olume single-zone system, the space humidity ratio wr must be determined. Because the winter indoor temperature is given, the space relative humidity can then be found from the psychrometric chart. Assume that the moist v olume of the w arm supply air vs  13.75 ft 3 /lb (0.86 m 3 /kg). Because s  1 / vs, from Eq. (20.87) Q rl wr ws  60V˙s s h fg,32 60,000  13.75 60  21,140  1075  0.000605 lb / lb 

AIR SYSTEMS: BASICS AND CONSTANT-VOLUME SYSTEMS

20.71

Because wm  ws, wr ws 3200   0.15 wr wo 21,140 Therefore,

wr wo 

0.000605  0.0041 lb / lb 0.15

From the psychrometric chart, at To  15°F and o  30 percent, wo  0.0005 lb / lb. Then wr  wo 0.0041  0.0005 0.0041  0.0046 lb / lb (kg / kg) Point r can thus be plotted on the psychrometric chart. The relative humidity of the space air at winter design condition is 30 percent. 2. From the gi ven data, Tru  Tr 2 1  72 2 1  75°F. Because wru  wr, points ru and o can be plotted from the psychrometric chart, as shown in Fig. 20.22, upper f gure. Connect ru-o. For an adiabatic mixing process ru-m 3200   0.15 ru-o 21,140 Section ru-o measured from the psychrometric chart is 6.15 in. Then mixing point m can be determined from the psychrometric chart: Tm  64.3°F (17.9°C)

and

wm  0.004 lb / lb (kg / kg)

From Eq. (20.86), the supply air temperature Ts  Tr

Q rh 60 V˙sscpa

 70

225,000  13.75  80.0F (26.7C) 60  21,140  0.243

Because wm  wsf  0.004 lb / lb, point s can be plotted. From the psychrometric chart, vs  13.72 ft3/ lb (0.856 m3 /kg) which is approximately equal to the assumed value. If the assumed value is too high or too low, the calculation will be repeated with a new value of vs. 3. Warm air temperature off the heating coil at winter heating mode can be calculated by Tch  Ts 1 2  80.0 1 2  79.0°F And from Eq. (20.85), if s  1 / 13.70, the heating coil load is Q ch  60 V˙sscpa(Tch Tm)  60  21,140 

1 (0.243)(79.0 64.3)  330,719 Btu / h (96,901 W) 13.70

Heating Mode with Space Humidity Control In health care b uildings with special requirements, or in industrial b uildings humidi f cation may have to be incorporated in the w arm air supply in winter . Figure. 20.27 sho ws an air conditioning cycle operated in winter heating mode with steam injection humidi f cation and the corresponding schematic diagram of the constant-v olume single-zone system. This cycle consists of the follo wing processes:

20.72

CHAPTER TWENTY

FIGURE 20.27 system.

Air conditioning cycle — winter heating mode operation with space humidity control for a constant-v olume single-zone

1. 2. 3. 4. 5. 6. 7.

Sensible heating process r-ru from return system heat gain Mixing process ru-m-o of recirculated air and outdoor air Sensible heating process m-ch at the heating coil Steam injection or heated element humidifying process ch-h Sensible heating process h-sf from supply fan power heat gain Sensible cooling process sf-s from duct heat loss Space conditioning process sr

The procedure for constructing this cycle is as follows: 1. Plot the state points of the conditioned space r and outdoor air o. 2. Draw the space conditioning line sr from point r with known sensible heat ratio SHRs. 3. From Eq. (20.87), calculate wr ws. Since wr, Qrl, and the supply v olume f ow rate from summer cooling mode calculations are kno wn values, ws can be calculated. Point s must lie on line sr, so the intersection of sr and ws line determines point s. 4. Draw a horizontal line r-ru from point r, representing the return system sensible heat gain. Plot point ru from its known temperature rise. 5. Connect points ru and o. From the ratio of the v olume f ow rate of the outdoor air to the supply air ,V˙o / V˙s point m can be found and located on line ru-o. 6. Draw horizontal line m-ch from point m. Draw also a horizontal line sf-s-h through point s. Temperature dif ference Tsf Th represents the temperature rise of the supply f an heat gain, and Tsf Ts is the supply duct heat loss. This def nes points sf and h.

AIR SYSTEMS: BASICS AND CONSTANT-VOLUME SYSTEMS

20.73

7. Draw a vertical line through ch to line h-sf extended from point h. The intersection of lines h-ch and m-ch locates point ch. Because wm  wch, the mass f ow rate of the injected steam is the mass f ow rate of the water vapor m˙s, in lb / min (kg / s), added to the moist air, and can be calculated from Eq. (20.42) m˙s  V˙ss(wh wm)

(20.88)

where wh  humidity ratio of air leaving humidif er, lb / lb (kg / kg) wm  humidity ratio of mixture, lb / lb (kg / kg) Part-Load Operation For a constant-volume single-zone system, when the space heating load is reduced during part-load operation, as in the the cooling mode operation, hot water f ow rate modulation is used to reduce the supply air temperature and the heating capacity if a hot water heating coil is employed in an airhandling unit. F or a packaged unit using a gas- f red furnace as a heating source, a DDC unit controller will directly adjust the amount of gas supplied to the b urner when the temperature senser senses the rise of space temperature during part-load operation.

Dual-Thermostat Year-Round Zone Temperature Control For most of the single-zone constant-v olume systems emplo yed in residential b uildings and small retail stores, two separate thermostats with one set point for cooling mode operation and another set point for heating mode operation are often used, as shown in Fig. 20.28. During cooling mode operation, when the zone temperature drops below the cooling set point, the refrigeration system is shut down; or during heating mode operation, when the zone temperature rises abo ve the heating set point, the gas furnace is shut off. On / off control is most often used. Between the cooling and heating mode operation, there is a dead band in which neither cooling nor heating is pro vided. During the dead band mode, the constant-v olume single-zone system is

Heating mode

Dead band

Cooling mode

Dead band

Output 100%

0

70

76 Tr , F

FIGURE 20.28

Dual-thermostat year-round zone temperature control.

20.74

CHAPTER TWENTY

often shut do wn, and sometimes the f an remains continuously operating to pro vide necessary outdoor ventilation air for interior spaces. The on / off control of the cooling or heating mode operation for such a constant-v olume single-zone system is performed manually or automatically . It is possible to ha ve a dead band width of 5 °F (2.8°C) or e ven wider for such a dual dual-thermostat year round zone-temperature control system. If the cooling capacity of a single-zone constant-v olume system is greater than 5 tons (17.6 kW), multiple single-zone constant-volume packaged units may be used.

20.20 CONSTANT-VOLUME MULTIZONE SYSTEM WITH REHEAT Reheating, Recooling, and Mixing In zone controls for multizone systems, reheating is a process in which air is reheated after it has been cooled. Recooling is a process in which air is cooled after it has been pre viously heated by using combustion fuels, and in the mixing process a cooler airstream is often mix ed with w armer recirculating air, or a w armer airstream mix ed with cooler outdoor air . Reheating, recooling, and mixing processes are all simultaneous cooling and heating processes in air conditioning c ycles, and are energy-ineff cient air conditioning processes. ASHRAE / IESNA Standard 90.1-1999 speci f es that zone thermostatic controls (either in constant-volume or VAV systems) shall be capable of operating in sequence the supply of heating and cooling energy to the zone. Zone controls shall pre vent (1) reheating, (2) recooling, (3) mixing or simultaneously supplying air that has been pre viously mechanically heated with air that has been previously cooled by refrigeration or by an economizer , and (4) other simultaneous operation of heating and cooling systems to the same zone e xcept the follo wing (exceptions for VAV systems will be listed and discussed in Chap. 21): a. Zones in which the v olume f ow of reheated, recooled, or mixed air is no greater than the lar ger of the following: ●

●

The v olume f ow of outside air to meet the v entilation requirement of ASHRAE Standard 62-1999. 0.4 cfm / ft2 (2L / s  m2) of the zone conditioned f oor area pro vided that primary air temperature is controlled: 0 to 12°F (0 to 6.7 °C) below the design space heating temperature Tr,h when outdoor air temperature To  60°F (15.6°C) for reheat systems and the cold deck mixing systems. 0 to 12°F (0 to 6.7°C) above design space temperature Tr when To  60°F (15.6°C) for recooling system and the hot deck mixing systems. Each zone need not comply with this e xception provided the a verage of all zones serv ed by this system that have both heating and cooling ability comply.

●

●

300 cfm (142 L / s). This exception is for zones whose peak f ow rate totals no more than 10 percent of the total fan system (air system) volume f ow rate. Any higher rate that is satis f ed by the authority ha ving jurisdication to reduce o verall air system annual ener gy usage by of fsetting reheat / recool energy losses through a reduction in outdoor air intak e in accordance with the multiple space requirements in ASHRAE Standard 62-1999.

b. Zones where special pressurization relationships, cross-contamination requirements, or code-required minimum circulation rates are such that v ariable-air-volume systems are impractical. c. Zones where at least 75 percent of the ener gy for reheating or for w arm air mixing systems is provided from a site recovered (including condenser heat) or site solar energy source.

AIR SYSTEMS: BASICS AND CONSTANT-VOLUME SYSTEMS

20.75

Reheating is a simple and effective means of controlling space temperature and relati ve humidity at part load. Reheating is used with compliance to ASHRAE / IESNA Standard 90.1-1999 for VAV systems at minimum v olume f ow rate settings for comfort air conditioning, and in hospitals, laboratories, computer rooms, museums, and precision manuf acturing environments (such as semiconductor manufacturing plants) in order to maintain a close tolerance in space temperature ( 1°F or 0.56°C) or space relative humidity for processing air conditioning.

Constant-Volume Multizone System with Reheat A constant-v olume, multizone system with reheat is sho wn in Fig. 20.29. During cooling mode operation, as shown in Fig. 20.29, outdoor air at point o is mixed with recirculating air at point ru. The mixture m f ows through the f lter. It is then cooled and dehumidi f ed to point cc at the cooling coil and dischar ged from the AHU or PU at the supply f an outlet sf. After absorbing the duct heat gain, the conditioned air enters the reheating coils 1 and 2 at point s. During summer design load operation, theoretically, reheating coil 1 is not ener gized. Air is supplied to room 1 at condition s1. Conditioned air is reheated at reheating coil 2 to maintain the required room temperature in room 2. After absorbing the space cooling load, the supplied air becomes room air r1 and r2. Returned air from rooms 1 and 2 is mixed together and forms mixture rm in the ceiling plenum. The return air is then returned to the AHU or PU, and the recirculating air enters the AHU or PU at point ru. A portion of return air is e xhausted through the relief f an to the outdoors. During winter operation, outdoor air is mix ed with recirculating air in a ratio such that its amount is always greater than the minimum outdoor air requirement. The mixture m f ows through the f lter, the deenergized cooling coil, and the supply duct and is reheated at the reheating coils 1 and 2. The heated air is supplied to rooms 1 and 2 to of fset the heating load there. Both hot w ater and electric reheating coils can be used in multizone systems. When electric reheating coils are used, the design and installation requirements described in Sec. 8.4 and the requirements of the National Electric Code and local codes must be followed.

Control Systems As described in Sec. 5.14, the controls for a constant-volume multizone system with reheat using an AHU or a PU may include the following: ●

●

●

●

●

Discharge air temperature control Outdoor ventilation air control Space pressurization control Air economizer control Zone smoke control

All these functional controls for an air system are discussed in Chap. 23 except zone smoke control, discussed in Chap. 22. If there are only a fe w reheating coils, both the AHU or PU and the reheating coils can be controlled by a DDC unit controller . If there are man y reheating coils, the AHU or PU and the reheating coils should be controlled by a DDC system controller. When room temperature sensors T2 and T3 sense a drop in either room temperature belo w a limit, a signal is sent to the DDC controller. The reheating coil is then energized. For a hot water reheating coil, the two-way valve is modulated so that room temperature is maintained within predetermined limits. F or an electric reheating coil, heating is added or decreased in steps. To prevent f res, electric heating is energized only when airf ow is sensed by a f ow switch.

FIGURE 20.29

20.76

Constant-volume multizone system with reheat: (a) schematic diagram; (b) air conditioning cycle.

AIR SYSTEMS: BASICS AND CONSTANT-VOLUME SYSTEMS

20.77

Operating Parameters and Calculations 1. The zone supply air condition should be determined so that the space temperature and relative humidity in each zone or room can be maintained at design load operation by using the reheating processes. 2. The zone n supply mass f ow rate m˙an, in lb / min [kg / (60 s)], and volume f ow rate V˙sn, in cfm [kg / (60 s)], at summer design condition can be calculated as m˙an 

Q rsn 60cpa(Trn Ts)

V˙sn  vsm˙an

(20.89) (20.90)

where Qrsn  sensible cooling load for zone n, Btu / h (W) Trn  temperature of zone n,°F (°C) Ts  supply temperature of zone n,°F (°C) vs  1 / s, moist volume of supply air, ft3 / lb (m3 / kg) 3. The zone n supply temperature at minimum part-load operation Tsnp, in°F (°C), can be calculated as Tsnp  Trnp

Q rsn 60m˙ancpa

(20.91)

where Trnp  temperature of zone at minimum part-load, operation °F (°C). 4. The zone n reheating coil load Qhn, in Btu / h (W), at minimum part-load operation can be calculated as Q hnp  60m˙ancpa(Tsnp Ts)

(20.92)

System supply temperature Ts varies along the supply main duct. F or simplicity, Ts can be taken as an average value. 5. The weighted mean v alue of zone temperature Trm, in °F (°C), and mean humidity ratio wrm, in lb / lb (kg / kg), can be calculated as

and

Trm 

m˙a1Tr1 m˙a2Tr2    m˙anTrn m˙a1 m˙a2    m˙an

(20.93)

wrm 

m˙a1wr1 m˙a2wr2    m˙anwrn m˙a1 m˙a2    m˙an

(20.94)

where m˙a1, m˙a2,   , m˙an  mass f ow rate of supply air for zones 1, 2, . . . , n, lb / min (kg / s) Tr1, Tr2, . . ., Trn  air temperatures in zones 1, 2, . . . , n,°F (°C) wr1, wr2, . . ., wrn  air humidity ratio in zones 1, 2, . . . , n, lb / lb (kg / kg) 6. The system supply volume f ow rate V˙s, in cfm (m3 /s), of the AHU or PU is calculated as V˙s  V˙s1 V˙s2    V˙sn

(20.95)

where supply volume f ow rate for zones 1, 2, . . . , n, cfm (m3 /s). V˙s1, V˙s2, . . . , V˙sn  7. The condition of recirculating air entering the AHU or PU, point ru, and the mixture of outdoor and recirculating air m can be determined graphically on the psychrometric chart.The cooling coil load Qcc, in Btu / h (W), can therefore be calculated as Q cc  60(m˙a1 m˙a2    m˙an)(h m h cc)

(20.96)

20.78

CHAPTER TWENTY

where hm, hcc  enthalpy of mixture and conditioned air off coil, Btu / h (W) m˙a  mass f ow rate of mixture, cfm [kg / (60 s)] 8. At part-load operation, space cooling load reduces and more reheating is required. The condition of zone supply air point snp can be determined because snp must lie on horizontal line ccsnp. Zone supply air temperature at part load Tsnp can be calculated from Eq. (20.91). Dra w a line from snp with the known sensible heat ratio for zone n at part load SHR np. This line intersects zone temperature line Trnp at point rnp, which is the state point of zone air at part load. Zone relati ve humidity rnp can then be determined. The operating parameters during winter heating mode operation at design load can be determined as follows: ●

At winter design load, wrm  wru and ws  wm. Here wru indicates the humidity ratio of the recirculating air, ws is the humidity ratio of system supply air before entering the reheating coil, and wm is the humidity ratio of the mixture of the outdoor air and recirculating air , all in lb / lb (kg / kg). Then wru  ws Also,

●

Q rl 60(m˙a1 m˙a2    m˙an)h fg,32 wru wm V˙o  wru wo V˙s

Q rhn

(20.99) 60V˙snscpa where Qrhn  space heating load for zone n, Btu / h (W). The reheating coil load for zone n during winter heating mode operation Qhn, in Btu / h (W), is Q hn  60m˙ancpa(Tsn Ts)

●

(20.98)

where wo  humidity ratio, lb / lb (kg / kg); V˙o  volume f ow rate of outdoor air, cfm [kg / (60 s)]. Zone supply temperature Tsn, in °F (°C), can be calculated as Tsn  Trn

●

(20.97)

(20.100)

where Ts  supply air temperature before reheating coil, °F (°C). Usually, the reheating coil load during winter heating mode operation is greater than in summer cooling mode minimum partload operation, and should be tak en as the design capacity of the reheating coil. Ho wever, it is necessary to select the greater value of these two as the reheating coil design capacity. At part-load operation, space heating load reduces, and the reheating coil load decreases accordingly.

System Characteristics If the humidity ratio of the supply temperature Ts remains the same during summer cooling mode or winter heating mode operation, the space relative humidity of a constant-v olume multizone system with reheat at part load is always lower than that at design load. There is also another constant-volume multizone system that has a hot deck and a cold deck. For each control zone, a separate air duct is connected to the hot deck, and another separate air duct is connected to the cold deck; each has a damper . Constant-volume multizone systems w aste energy and also result in a v ery complicated control system. They are now replaced by other more ener gyeff cient systems.

AIR SYSTEMS: BASICS AND CONSTANT-VOLUME SYSTEMS

20.79

REFERENCES AMCA, Fan and Systems, Application Guide 201 – 90, Air Movement and Control Association, Arlington Heights, IL., 1990. ASHRAE, ASHRAE Handbook 1996, HVAC Systems and Equipment, ASHRAE Inc., Atlanta, GA, 1996. ASHRAE, ASHRAE Handbook 1997, Fundamentals, Atlanta, 1997. ASHRAE / IESNA Standard 90.1-1999, Energy Standard for Buildings Except New Low-Rise Residential Buildings, ASHRAE Inc., Atlanta, 1999. Avery, G., No More Reheat: Banking Center Retrof t, Heating / Piping / Air Conditioning, no. 3, 1984, pp. 98 – 99. Carrier Corporation, Handbook of Air Conditioning System Design, McGraw-Hill, New York, 1965. Coad, W. J., The Air System in Perspective, Heating / Piping / Air Conditioning, no. 10, 1989, pp. 124 – 125. Coogan, J. J., Effects of Surrounding Spaces on Rooms Pressurized by Differential Flow Control, ASHRAE Transactions, 1996, Part I, pp. 18 – 25. Coward, C. W., Jr., A Summary of Pressure Loss Values for Various Fan Inlet and Outlet Duct Fittings, ASHRAE Transactions, 1983, Part I B, pp. 781 – 789. Cummings, J. B., Withers, C. R., Jr., Moyer, N. A., Fairey, P. W., and McKendry, B. B., Field Measurement of Uncontrolled Airf ow and Depressurization in Restaurants, ASHRAE Transactions, 1996, Part I, pp. 859 – 869. Desmone, C. L., and Frank, P. L., Air Conditioning for Precision Manufacturing, Heating / Piping / Air Conditioning, no. 2, 1992, pp. 35 – 44. Driscoll, D. J., System Effect – The Balancer’s Dilemma? ASHRAE Transactions, 1983, Part I B, pp. 795 – 801. Energy Information Administrations, Commercial Buildings Characteristics, EIA, Washington, 1994. Holness, G. R., Pressurization Control: Facts and Fallacies, Heating / Piping / Air Conditioning, no. 2, 1989, pp. 47 – 51. Lovatt, J. E., and Wilson, A. G., Stack Effect in Tall Buildings, ASHRAE Transactions, 1994, Part II, pp. 420 – 431. Manley, D. L., Bowlen, K. L., and Cohen, B. M., Evaluation of Gas-Fired Desiccant-Based Space Conditioning for Supermarkets, ASHRAE Transactions, 1985, Part I B, pp. 447 – 456. Meckler, G., Eff cient Integration of Desiccant Cooling in Commercial HAVC Systems, ASHRAE Transactions, 1988, Part II, pp. 2033 – 2042. O’Connor, J. E., The System Effect and How It Changes Fan Performance, ASHRAE Transactions, 1983, Part I B, pp. 771 – 775. Persily, A. K., Myths about Building Envelopes, ASHRAE Journal, no. 3, 1999, pp. 39 – 47. Persily, A. K., and Linteris, G. T., A Comparison of Measured and Predicted Inf ltration Rates, ASHRAE Transactions, 1983, Part II B, pp. 183 – 200. Shaw, C. Y., Reardon, J. T., and Cheung, M. S., Changes in Air Leakage Levels of Six Canadian Off ce Buildings, ASHRAE Journal, no. 2, 1993, pp. 34 – 36. Tamblyn, R. T., Coping with Air Pressure Problems in Tall Buildings, ASHRAE Transactions, 1991, Part I, pp. 824 – 827. Tamura, G. T., and Wilson, A. G., Pressure Differences Caused by Chimney Effect in Three Tall Buildings, ASHRAE Transactions, vol. 73, no. 2, 1967. Tamura, G. T., and Shaw, C. Y., Studies on Exterior Wall Air Tightness and Air Inf ltration of Tall Buildings, ASHRAE Transactions, 1976, Part I, pp. 122 – 134. The Trane Company, Building Pressurization Control, The Trane Company, La Crosse, WI, 1982. Wang, S. K., Air Conditioning, vol. 4, Hong Kong Polytechnic, Hong Kong, 1987. Withers, C. R., Cummings, J. B., Moyer, N. A., Fairey, P. W., and McKendry, B. B., Energy Savings from Repair of Uncontrolled Airf ow in 18 Small Commercial Buildings, ASHRAE Transactions, 1996, Part II, pp. 549 – 561.

CHAPTER 21

AIR SYSTEMS: VARIABLE-AIR-VOLUME SYSTEMS 21.1 SYSTEM CHARACTERISTICS OF VARIABLE-AIR-VOLUME SYSTEMS 21.2 Variable-Air-Volume Systems 21.2 Types of Vairable-Air-Volume Systems 21.2 21.2 SINGLE-ZONE VAV SYSTEMS 21.3 System Description 21.3 Year-Round Operation of a Single-Zone VAV System 21.5 Economizer Cycle and Economizers 21.8 Differential Enthalpy, Electronic Enthalpy, and Fixed Enthalpy Economizer Control 21.8 Fixed Dry-Bulb and Differential Dry-Bulb Economizer Control 21.9 Comparison of Enthalpy-Based and Temperature-Based Economizer Control 21.10 Water Economizer and Control 21.12 Comparison of Air and Water Economizers 21.14 ASHRAE / IESNA Standard 90.1-1999 Economizer Control Specifications 21.14 Air Conditioning Cycle and System Calculations 21.16 Zone Temperature Control — Sequence of Operations 21.17 21.3 VAV COOLING, VAV REHEAT, AND PERIMETER HEATING VAV SYSTEMS 21.18 VAV Cooling Systems 21.19 VAV Reheat Systems 21.20 System Description 21.20 Perimeter Heating VAV System 21.20 VAV Box 21.21 Reheating VAV Box 21.23 Sound Power Level of a VAV Box 21.23 VAV Reheat Zone Temperature Control and Sequence of Operations 21.25

Stability of Zone Control Using VAV Boxes and Reheating VAV Boxes 21.26 Case Study: A VAV Reheat System 21.27 Conditioned Air Off-Coil and Supply Temperature Differential 21.28 System Volume Flow Rate and Coil Load 21.28 Summer Cooling Mode Part-Load Operation 21.29 Winter Reheating in Perimeter Zone 21.29 Winter Cooling Mode Operation in Interior Zone 21.32 21.4 DUAL-DUCT VAV SYSTEMS 21.33 System Description 21.33 Number of Supply Fans 21.36 Mixing VAV Box 21.36 Mixing Mode Operation 21.38 Zone Controls and Sequence of Operations of Dual-Fan Dual-Duct VAV System 21.38 Discharge Air Temperature Control 21.40 Zone Supply Volume Flow Rate 21.41 Case Study: A Dual-Fan Dual-Duct VAV System 21.42 Winter Heating and Winter Cooling Mode Operation 21.43 Part-Load Operation 21.43 21.5 FAN-POWERED VAV SYSTEMS 21.44 System Description 21.44 Fan-Powered VAV Boxes 21.48 Fan Characteristics in Parallel FanPowered Boxes 21.50 Zone Control and Sequence of Operations of a Fan-Powered VAV System with Parallel Fan-Powered Box 21.52 Supply Volume Flow Rate 21.53 Fan Energy Use 21.54 Design Considerations 21.55 21.6 COMPARISON BETWEEN VARIOUS VAV SYSTEMS 21.56 REFERENCES 21.56

21.1

21.2

CHAPTER TWENTY-ONE

21.1 SYSTEM CHARACTERISTICS OF VARIABLE-AIR-VOLUME SYSTEMS Variable-Air-Volume Systems A variable-air-volume (VAV) system is an air system that v aries its supply air v olume fl w rate to match the reduction of space load during part-load operation to maintain a predetermined space parameter, usually air temperature, and to conserve fan power at reduced v olume fl w. A constantvolume system v aries its supply air temperature to match the reduction of space load during partload operation to maintain a predetermined space air temperature. Compared with a constant-v olume system, a VAV system has mainly the follo wing advantages: ●

●

●

●

●

●

Reduced fan energy use during part-load operation when the supply v olume fl w rate is reduced A slightly lo wer or nearly the same zone relati ve humidity when the supply v olume fl w rate is reduced during summer cooling mode part-load operation More individual control zones Reduction of the construction cost because of taking into consideration of the supply air v olume fl w diversity factor instead of the sum of zone peak loads Capability of self-balancing of zone supply volume fl w rates Convenience during the relocation of the terminals and space dif fusion de vices during future expansion or retrofi

Compared with a constant-volume system, the primary disadvantages of a VAV system are: ●

●

Inadequate outdoor ventilation air when the supply volume fl w rate is reduced A more complicated system structure and controls, which need more demanding design, installation, operation, and maintenance

VAV systems are applicable to air systems whose space load v aries significantly so that there ar fan energy savings. VAV systems became popular after the ener gy crisis in 1973. They are widely used in large commercial buildings in the United States. Types of Variable-Air-Volume Systems Most medium-size and lar ge b uildings need multizone air systems. Ho wever, many indoor stadiums, convention centers, factories, residential buildings, and small retail stores emplo y single-zone air systems. Currently used variable-air-volume systems can be mainly classified into the foll wing types: ●

●

●

●

●

Single-zone VAV systems VAV cooling systems VAV reheat systems Dual-duct VAV systems Fan-powered VAV systems

Recently, another VAV system has been de veloped called the v ariable diffuser VAV system. In variable diffuser systems, the aperture of each dif fuser can be v aried so that the dischar ge velocity is relati vely constant while the supply v olume fl w is reduced, and the thro w from the v ariable diffuser may also be maintained abo ve a certain limit. Variable diffuser VAV systems provide more individual control zones as well as a more complicated dif fuser construction and controls. More field perfomance data and xperience are needed to make an appropriate selection.

AIR SYSTEMS: VARIABLE-AIR-VOLUME SYSTEMS

21.3

Conditioned spaces in b uildings are primarily di vided into tw o categories: (1) perimeter zone, where there are e xternal wall, roofs, and windows and space load v aries depending on solar heat, outdoor-indoor temperature dif ference and internal load; and (2) interior zone where space loads are mainly internal loads. In an interior zone, the air system often operates at cooling mode in both summer and winter e xcept for morning w arm-up. A perimeter zone can often be subdi vided into four subzones according to their orientations. Although all these subzones operate at cooling mode in summer, in f all and winter , zones f acing south in the northern hemisphere may need cooling, whereas zones facing north may need heating. Under such circumstances, a VAV system that serves a perimeter zone is preferable to ha ving a modulating cooling de vice and a modulating heating device to meet the variable cooling and heating requirements for each of its subzones f acing different orientations. The system characteristics of dif ferent types of VAV systems v ary mainly because of the zone control of their different types of VAV boxes to serve perimeter zone, except a single-zone VAV system which directly varies the fan speed or the inlet vanes and a dual-duct VAV system in which two supply ducts (a cold air duct and a w arm air duct) are used. A VAV cooling system with VAV boxes usually serves an interior zone, and reheating VAV boxes or a VAV cooling system plus a perimeter heating system (such as a hot water heating system) will serve a perimeter zone. A single-zone VAV system can be used for either a perimeter zone or an interior zone. Excluding the terminals and zone controls, the follo wing system characteristics of dif ferent VAV systems are similar to each other, and are discussed in later chapters. ●

●

●

●

●

Fan combination System pressure analysis and space pressurization Smoke control Outdoor ventilation air Controls, including discharge temperature, duct static pressure, and safety controls

ASHRAE / IESNA Standard 90.1-1999, Energy Standard for Buildings Except Low-Rise Residential Buildings, specifies that ind vidual variable-air-volume (VAV) fans with 30-hp (22-kW) motors and lar ger shall ha ve other controls and de vices (such as adjustable-frequenc y, variable-speed drives) that will result in fan motor demand of no more than 30 percent of design w attage at 50 percent of design v olume fl w rate when f an static pressure equals one-third of the total design static pressure based on manufacturer’s certified data

21.2 SINGLE-ZONE VAV SYSTEMS A single-zone VAV system is a VAV system that v aries the air v olume fl w rates supplied to and returned from a single-zone conditioned space to maintain a predetermined space parameter at reduced load and to conserv e fan power. Single-zone systems are widely used in indoor stadiums, assembly halls, shopping malls, and factories.

System Description Figure 21.1 sho ws the schematic diagram and the air conditioning c ycles for a typical single-zone VAV system. During the summer cooling mode operation, the mixture of outdoor air and recirculating air at point m is filtered at the filter and cooled and dehumidified at the cooling coi The conditioned air lea ves the coil at point cc. It then fl ws through the supply f an of the air -handling unit (AHU) or packaged unit (PU) and the supply duct before it is dischar ged to the single-zone conditioned space at point s. After absorbing the space cooling load, the supply air becomes the space air r and returns to the AHU or PU through the return air system.

FIGURE 21.1 A single-zone VAV system: (a) schematic diagram; (b) air conditioning cycle.

21.4

AIR SYSTEMS: VARIABLE-AIR-VOLUME SYSTEMS

21.5

When air economizer c ycles are used for free cooling during spring and f all seasons, outdoor air intake may vary from minimal to 100 percent. A relief fan should be installed in combination with the supply fan, if the pressure drop of the return duct system is small. If the pressure drop of the return duct system is greater than 0.6 in. WC (149 Pa), a return fan in series with a supply fan should be considered. During summer cooling mode part-load operation, the supply and return v olume fl w rates are reduced to match the lo wer space cooling load. The temperature rise from the f an power heat gain is reduced accordingly . At the same time, duct heat gain remains approximately the same, which results in a higher temperature rise for supply and return ducts during summer cooling mode partload operation. During winter, if the solar heat and internal load is greater than the space transmission and infil tration loss, then a cooling mode operation of a cold air supply is still needed, such as in fully occupied indoor stadium and arenas. Outdoor air is often mix ed with the recirculating air to form a mixture at a temperature around 55°F (12.8°C) during the air economizer c ycle for free cooling. Warm air may be needed only during the morning w arm-up period to raise the space temperature to an acceptable value before the space is to be occupied. If solar heat and the internal load are less than the transmission and infiltration loss the mixture of outdoor and recirculating air is heated at the heating coil to point hc. It is extracted by the supply fan and forced through the supply duct. Warm air is then supplied to the conditioned space at point s, and the recirculating air is returned to the AHU or PU through the ceiling plenum and connecting return ducts. For comfort air conditioning without humidification in winte , the space relati ve humidity depends mainly on the space latent load and the volume fl w ratio of outdoor air to supply air. For many VAV systems in commercial b uildings, the ceiling plenum is often used as the return plenum. The primary sound transmission path of supply fan noise from the AHU or PU to the space is often duct-borne noise through the return system. A sound trap or silencer should be added in the f an room between the AHU or PU and the connecting duct before penetrating the fan room partition wall. Year-Round Operation of a Single-Zone VAV System In this single-zone VAV system, cooling is provided by a refrigerant plant and e vaporative coolers, and heating is provided by the hot water boiler through a hot water heating coil. For convenience in analysis, consider an interior zone with zone air temperature maintained at 75°F (23.9°C) year round. During cooling mode operation, the required of f-coil temperature is 55°F (12.8°C) with a relative humidity of 95 percent. The recirculating air enters the mixing box at 78°F (25.5°C). It is mixed with an amount of outdoor air approximately equal to 20 percent of the total design supply volume fl w rate and is cooled and dehumidified in a cooling coil. During winter heating mod operation, cold air supply is still required to maintain a year -round space temperature of 75°F (23.9°C); outdoor air and recirculating air are mix ed together to pro vide a mixture of 55°F (12.8°C). It is also assumed that the b uilding is in a location where no humidification is required t maintain a space relative humidity no lower than 30 percent during the winter season. To achieve effective and ener gy-efficient operation when the outside weather changes from ho summer to cold winter , the year-round operation of a single-zone VAV air system described abo ve can be di vided into the follo wing operating modes, corresponding to the v arious outside weather regions on the psychrometric chart, as shown in Fig. 21.2: Region I: Refrigeration and Evaporative Cooling. The lower left boundary line of region I is the (dashed) enthalpy line that passes through the state point ru of recirculating air entering the mixing box. When the outside weather at point o is inside region I, the enthalpy of the outside air is higher than that of the recirculating air , that is, ho  hru. In these instances, it is more ener gy-efficient t condition (cool and dehumidify) the recirculating air than the outside air . As such, the quantity of outdoor air mix ed with the recirculating air will be the minimum needed to meet the required outdoor ventilation rate in the conditioned space. In region I, refrigeration or e vaporative cooling or both are required to cool and dehumidify or to cool and humidify the mixture of outdoor and recirculating air to point cc in order to maintain the predetermined space condition r.

21.6

CHAPTER TWENTY-ONE

FIGURE 21.2 Enthalpy economizer control: (a) outside weather regions; (b) enthalpy economizer.

AIR SYSTEMS: VARIABLE-AIR-VOLUME SYSTEMS

21.7

For region I, the operating characteristics of the VAV system include the following: ●

●

●

●

The outdoor damper is opened to such a de gree that only that minimum required amount of outdoor air is extracted. The recirculating damper is fully opened. Supply air is cooled and dehumidified or cooled and humidified to point cc. A refrigeration system, evaporative coolers, or both are operating.

Region II: Free Cooling and Refrigeration / Evaporative Cooling. Region II is enclosed by four boundary lines: the enthalpy line through point ru at the upper right, part of the saturation curve, the 55°F (12.8°C) temperature line, and the cc-ru line. In region II, the enthalpy of outdoor air is lo wer than that of recirculating air , that is, ho  hru. Less energy will be used in conditioning if 100 percent of outside air is cooled and dehumidifie instead of a mixture outdoor air and recirculating air . Both free cooling by the outside air and refrigeration / indirect evaporative cooling (to be discussed in Chap. 27) are required to bring the 100 percent outside air down to 55°F (12.8°C) with a relative humidity of 95 percent. The operating characteristics of this VAV system when outdoor air is inside region II include the following: ●

●

●

●

The outdoor damper and the exhaust damper or relief damper both are fully opened. The recirculating damper is closed. Outdoor air is cooled and dehumidified to point cc. Point cc is at a temperature of 55°F (12.8°C) and a humidity ratio of 0.0088 lb / lb (kg / kg). Refrigeration system / and indirect evaporative cooler are operating.

Region III: Free Cooling and Evaporative Cooling / Refrigeration. Region III is bounded by the line cc-ru, 55°F (12.8°C) temperature lines, and enthalpy lines passing through points ru and cc for wcc  0.005 lb / lb (kg / kg), as shown in Fig. 21.2. F or outdoor air inside this re gion, ho  hru, the humidity ratio of outdoor air wo is approximately less than 0.009 lb / lb (kg / kg). Indirect and direct evaporative cooling plus necessary refrigeration is required to maintain a space temperature of 75°F (23.8°C) and a space relative humidity greater than or equal to 30 percent. When outdoor temperature To  78°F (25.5°C), it is ener gy-efficient to xtract 100 percent outdoor air (to mak e use of free cooling) and to use direct and indirect e vaporative cooling plus necessary refrigeration to cool and humidify the outdoor air to a temperature of 55°F (12.8°C) with a humidity ratio wcc  0.005 lb / lb (kg / kg). When To  78°F (25.5°C), it is ener gy-efficient t extract minimum outdoor air and mix it with recirculating air . The mixture is then e vaporatively cooled indirectly-directly with neccessary refrigeration to Tcc  55°F (12.8°C) and wcc  0.005 lb / lb (kg / kg). The operating characteristics of the single-zone VAV system when outdoor air is inside re gion III and when 100 percent outdoor air is extracted are as follows: ●

●

●

●

The outdoor air damper and the exhaust or relief air damper both are fully opened. The recirculating damper is closed. Evaporative coolers are used for evaporative cooling and humidifying. Necessary refrigeration is required.

When only a minimum of outdoor air is mixed with recirculating air, then ●

●

●

●

The outdoor air damper is opened to minimum. The recirculating damper is fully opened. Evaporative coolers are used for evaporative cooling and humidifying. Sometimes refrigeration may be required.

21.8

CHAPTER TWENTY-ONE

Region IV: Free Cooling, Humidification and Heating. Region IV is enclosed by part of the saturation curve, the 55°F (12.8°C) temperature line, and the enthalp y line passing through point cc for wcc  0.005 lb / lb (kg / kg), as shown in Fig. 21.2. For outdoor air state points in region IV, either To  55°F (12.8°C) and ho  hcc or To  78°F (25.5°C) and ho  hcc, when wcc  0.005 lb / lb (kg / kg). In modulation of outdoor and recirculating dampers, or in locations with v ery cold winter, heating may be required to pro vide a mixed air temperature of 55°F (12.8°C) when the v olume fl w of outdoor air is equal to the minimum requirement. Free cooling of outdoor air will be utilized. Optimum humidifying methods should be adopted to maintain a space relati ve humidity of preferably not less than 25 percent. In locations where the outdoor temperature may drop belo w 32°F (0°C) in winter , remedies should be provided to protect water from freezing. The operating characteristics of this VAV system include the following: ●

●

●

Outdoor, exhaust, and recirculating air dampers are modulated. Heating, humidification or cooling and humidification may be required The refrigeration system is not operating.

Year-round operation of a multizone VAV system having perimeter and interior zones, is similar to that for an interior zone VAV system with the e xception that heating may be required in re gions II, III, and IV. For a multizone VAV system, the space air temperature can be considered as the average value of multiple zones.

Economizer Cycle and Economizers An economizer c ycle is an air conditioning c ycle that utilizes the free cooling capacity of outdoor air either directly or to cool the condenser w ater in a cooling to wer (or an e vaporative cooler), and then to cool the air indirectly, instead of using refrigeration to pro vide cooling / dehumidification s as to maintain a required space temperature. The component and devices used in the operation of an economizer cycle are collectively called an economizer, and the type of control used to operate the economizer cycle effectively and energy-efficiently is called economizer control There are two types of economizers: 1. Air or air -side economizers. These bring in up to 100 percent outdoor air to utilize its free cooling capacity, rather than use refrigeration or e vaporative cooling, to of fset the space cooling load when the enthalp y of the outdoor air is lo wer than the enthalp y of the recirculating air , or the outdoor air temperature is lo wer than a preset air temperature or a set point which is a function of outdoor climate. An air economizer consists of outdoor , exhaust, relief, and recirculating ducts and dampers in the AHU or PU, as well as a control system to operate them (see Fig. 21.2 b). Air economizer control can be subdivided into enthalpy-based differential enthalpy, fi ed enthalpy, and electronic enthalpy economizer controls and also temperature-based fi ed dry-bulb and differential dry-bulb economizer controls. 2. Water or water -side economizer s. These economizers use the outdoor air to cool the condenser water or cooling w ater in the cooling to wer or evaporative cooler first and then to cool the the mixture of outdoor and recirculating air through a precooling coil. A water economizer consists mainly of a cooling to wer (or an e vaporative cooler), a water precooling coil in the AHU or PU, a circulating pump to circulate cooling water, and the associated control system.

Differential Enthalpy, Electronic Enthalpy, and Fixed Enthalpy Economizer Control The operation of an enthalp y-based economizer control system is based on the comparison of air enthalpies or the comparison of air enthalpies and air temperatures. In a dif ferential enthalp y

AIR SYSTEMS: VARIABLE-AIR-VOLUME SYSTEMS

21.9

economizer, the on-and-off air economizer is based on the comparison of the enthalp y of the outdoor air ho, in Btu / lb (kJ / kg), with the enthalp y of the recirulating air hru, in Btu / lb (kJ / kg). In cooling mode operation, if ho  hru, the differential enthalpy economizer is turned on. If ho  hru, the differential enthalpy economizer is shut of f. Enthalpies of outdoor air and recirculating air can be calculated from their measured dry-b ulb and wet-b ulb temperatures, or from their measured de w points. The sequence of operations of a typical differential enthapy economizer control system based on measured dry- and wet-bulb temperature of the outdoor and recirculating air is as follows: 1. The outdoor air temperature To is sensed by sensor T1, and the relati ve humidity of outdoor air o is sensed by humidity sensor H1. Both signals are sent to a DDC controller , where calculations are performed to combine the signals into an outdoor air enthalp y ho signal. This signal is compared with the recirculating air enthalpy hru, which is the combined signal from temperature T2 and humidity sensor H2, at the DDC controller. 2. If ho  hru, the DDC controller sends signals to actuate the damper actuators so that the outdoor air and exhaust air dampers will be at their minimum opening positions, as determined by the manual position switch. Meanwhile, the recirculating damper is fully opened. 3. If ho  hru, the DDC controller sends signals to actuate the damper actuators to fully open the outdoor air and exhaust air dampers; at the same time, the recirculating damper will be closed. 4. Temperature sensor T3 senses the mix ed air temperature of outdoor and recirculating air Tm for airfl w leaving the mixing box. This signal is compared with the set point Tm,s  55°F (12.8°C) in the DDC controller . When the outdoor air temperature To  55°F (12.8°C), under the condition that the opening of the outdoor damper is equal to or greater than its minimum opening position, the DDC controller then actuates the damper actuator to modulate the opening of the outdoor , exhaust, and recirculating air dampers to maintain a predetermined mix ed air temperature of Tm  55°F (12.8°C). When To  55°F (12.8°C), the refrigeration compressors or chillers that serv e this air system can be shut down. In an electronic enthalp y economizer control, outdoor air is compared with a set point A which can be represented on the psychrometric chart by a curve that goes through a point at approximately 75°F (23.9°C) and 40 percent relative humidity. This curve is nearly parallel to dry-bulb lines at low humidity le vel and nearly parallel to enthalp y lines at high humidity le vels. In a fi ed enthalp y economizer control, outdoor air is often compared with a fi ed enthalpy hoA of 75°F (23.9°C) and 50 percent relative humidity with an enthalpy of 28 Btu / lb (65 kJ / kg) at sea level. In cooling mode operation, if ho  28 Btu / lb (65 kJ / kg) at sea le vel, the fi ed enthalpy economizer shall be shut off. If h  28 Btu / lb (65 kJ/ kg), the fi ed enthalpy economizer shall be turned on.

Fixed Dry-Bulb and Differential Dry-Bulb Economizer Control There are two temperature-based economizer controls: fi ed dry bulb and differential dry bulb. In a fi ed dry-bulb economizer control, outdoor air temperature To is compared with a fi ed dry-bulb temperature Tdb. ASHRAE / IESNA Standard 90.1-1999 recommends the follo wing fi ed dry bulbs: Climate

Fixed dry bulb

High-limit shutoff

Dry Intermediate Humid

75 to 78°F (23.9 to 25.6°C) 70°F (21.1°C) 65°F (18.3°C)

To  75 to 78°F To  70°F To  65°F

From the psychrometric chart, the enthalpies of 75°F , 50 percent relati ve humidity, 70°F, 64 percent relative humidity, and 65°F, 85 percent relative humidity are all near to the v alue 28 Btu / lb (65 kJ / kg). In a dif ferential dry-bulb economizer control, outdoor air temperature To is compared with the fi ed recirculating temperature Tru. In cooling mode operation, if To  Tru, the differential

21.10

CHAPTER TWENTY-ONE

dry-bulb economizer control is turned on. If To  Tru, the differential dry-bulb economizer control is shut of f. For a space temperature Tr  75°F (23.9°C) in summer , Tru is usually equal to 75 3  78°F (25.6°C) after absorbing the lighting sensible heat in the return ceiling plenum. Theoretically speaking, it is more ener gy-efficient to compare the enthal y of outdoor air with that of recirculating air during air economizer control. However, comparing enthalpies necessitates having both temperature and humidity sensors or temperature and de w point sensors. In actual practice, humidity sensors may demonstrate considerable errors (sometimes up to 10 percent) and have extensive maintenance requirements. Dew point sensors are delicate and expensive and cause maintenance dif ficulties. Therefore, it is simpler and more con venient to use only temperature sensors and to compare the outdoor air temperature To with the recirculating temperature Tru (or a predetermined set point) instead of sensing and comparing enthalpies. This method of control is called temperature economizer control. Outside weather re gions using temperature-based economizer control are sho wn in Fig. 21.3 a and are similar to the outside weather re gions for enthalp y economizer control. The sequence of operations of a typical dif ferential dry-b ulb economizer control system is as follows: 1. The outdoor air temperature To is sensed by temperature sensor T1, as shown in Fig. 21.3b. Then To is compared with the recirculating temperature Tru sensed by temperature sensor T2. If To  Tru, the DDC controller sends signals to damper actuators so that the outdoor and e xhaust air dampers will be at their minimum opening, as determined by the manual positioning switch. The recirculating damper should be fully opened. 2. If To  Tru, the outdoor and e xhaust air dampers should be fully opened and the recirculating damper should be closed. 3. Temperature sensor T3 senses the mixed air temperature Tm after the mixing box. When the outdoor air temperature To  55°F (12.8°C), under the condition that the v olume fl w of outdoor air V˙o equal or be greater than the minimum requirement, the DDC controller modulates the outdoor, exhaust, and recirculating air dampers to maintain a nearly constant mix ed air temperature of 55°F (12.8°C). 4. When To  55°F (12.8°C), the refrigeration system used to serv e this single-zone VAV air system is shut off. In a fi ed dry-bulb economizer control, a single dry-bulb temperature set point Tdb, may be used as an indicator of when to change from using the minimum amount of outdoor air to using 100 percent outdoor air. The changeover is determined from a comparison of To with Tdb. The proper magnitude of Tdb depends on the outside climate and the characteristics of the air system as recommended by ASHRAE / IESNA Standard 90.1-1999. According to Wacker (1989), an investigation is required to determine the optimum Tdb.

Comparison of Enthalpy-Based and TemperatureBased Economizer Controls In addition to the instrumentation and maintenance problems of enthalp y economizer control, enthalpy-based and temperature-based economizer controls are dif ferentiated by the location of the status points of outside weather in the dotted area A or shaded area B shown in Fig. 21.3 a. When the outside weather is inside the dotted area A, differential-enthalpy economizer control using minimum outdoor ventilation air recirculation mode often consumes less ener gy to cool and dehumidify the mixture of outdoor and recirculating air than with the 100 percent outdoor air mode. When the outside weather f alls in the shaded area B, a differential dry-bulb economizer control using minimum outdoor v entilation air recirculating mode may sensibly cool the mixture at a smaller enthalpy differential than a 100 percent outdoor air intake in a differential enthalpy economizer control.

FIGURE 21.3 Temperature economizer control: (a) outside weather regions; (b) temperature economizer.

21.11

21.12

CHAPTER TWENTY-ONE

Consider a single-zone VAV system that serv es a b uilding located in Phoenix, Arizona. Assume that the outdoor temperature is 85 °F (29.4 °C), the humidity ratio is 0.006 lb / lb (kg / kg), and the conditioned air leaves the coil at a temperature of 55°F (12.8°C), with a relative humidity of 95 percent. The recirculating air enters the AHU or PU at a temperature of 78 °F (25.5°C) and a humidity ratio of 0.0094 lb / lb (kg / kg). If a dif ferential enthalpy economizer control is used, the state point of outdoor air oh falls in region III, and if a dif ferential dry-bulb economizer control is used, the state point of outdoor air oT falls in re gion I. The properties of moist air are as follows: Differential enthalpy Outdoor Temperature, °F Relative humidity, % Humidity ratio, lb / lb Enthalpy, Btu / lb Enthalpy difference, Btu / lb Amount of outdoor air, %

85 24 0.006 27.0

Off-coil 55 65 0.006 19.8 7.2 100

Differential dry bulb Recirculating 78 45 0.0094

Outdoor

Mixture

Off-coil

85 24 0.006

79.5 40 0.0086 28.6

55 90 0.0086 22.4

6.2 20

A differential dry-bulb economizer control is required to cool the mixture of outdoor and recirculating air do wn to an enthalp y dif ference of 6.2 Btu / lb (14.4 kJ / kg) instead of 7.2 Btu / lb (16.7 kJ / kg) using a differential enthalpy economizer control. Therefore, consider the following: 1. For locations having hot and humid climates where man y status points of outdoor air f all in the dotted area A in Fig. 21.3a, a differential enthalpy economizer control is recommended. 2. For most locations ha ving moderate outdoor humidity ratios, according to Spitler et al. (1987), the energy savings offered by using a dif ferential enthalpy economizer control instead of a temperature-based economizer are small. The two types of economizer control may dif fer by only about 5 to 10 percent. A f xed dry-bulb economizer control is recommended. 3. For locations having dry climates, a f xed dry-bulb of economizer control with a temperature setting of 75 to 78°F (23.9 to 25.1°C) sometimes even saves more cooling energy than a differential enthalpy economizer control. A f xed dry-bulb economizer control is strongly recommended. 4. For locations with humid climates, where the space requires a lo wer relative humidity, a f xed dry-bulb economizer control using a set point Tdb a few degrees Farenheit lower than space temperature sometimes is benef cial. 5. A survey including the simulation and comparison of various air economizer control alternatives is often benef cial. Water Economizer and Control Consider a typical w ater economizer that uses a precooling coil in either an AHU or PU, as shown in Fig. 21.4b, and has the following operating characteristics: ●

●

●

●

The cooling tower approach is assumed to be 5°F (2.8°C). The mixing temperature of outdoor air and recirculating air Tm, is 78°F (25.5°C). The minimum temperature dif ference between Tm and the condenser w ater temperature Tcon,w is 4°F (2.2°C). When the outdoor wet-b ulb temperature To  40°F (4.4 °C), the temperature of the condenser water used to replace the entire refrigeration through a heat e xchanger is less than 45 °F (7.2°C).

FIGURE 21.4 Water economizer control: (a) outside weather regions; (b) water economizer of an indoor PU.

21.13

21.14

CHAPTER TWENTY-ONE

●

The outdoor wet-b ulb temperature To,we below which the w ater economizer can be operated effectively is To,we  78 5 4  69°F (20.6°C)

During the operation of a w ater economizer , the outside weather can be grouped into three regions based on the outdoor wet-bulb temperature, as shown in Fig. 21.4a: Region I: The water economizer cannot be operated effectively. Region II: The water economizer is operating ef fectively and replaces part of the refrigeration. Region III: The water economizer replaces the refrigeration up to 100 percent. The di viding line between re gions I and II is the wet-b ulb To,we line. The di viding line between regions II and III is the wet-b ulb Tr,of line, where Tr,of represents the outdoor wet-b ulb temperature below which the refrigeration plant serving the air system installed with a w ater economizer can be shut down. In locations where the outdoor climate is cold, and when the temperature discharged from the cooling tower drops below 35°F (1.7° C), the water economizer should be shut do wn for freeze protection.

Comparison of Air and Water Economizers The economizer cycle is a popular means of saving energy in the year-round operation of air conditioning systems. The amount of ener gy savings, compared with air systems for which the economizer cycle is not adopted, ranges from 15 to 40 percent of total energy use in various locations. The advantages of air economizers over water economizers include the following: 1. Because of the intake of a large amount of outdoor air, the indoor air quality is improved. 2. The ener gy input to an air conditioning system that uses an air economizer is approximately 20 percent less than that for a water economizer. 3. The cooling to wer requires less mak eup w ater, fewer w ater treatments, and reduced maintenance. However, the disadvantages of air economizers are as follows: 1. For an AHU or PU located in the interior core f an room or machinery room and used to serv e several f oors in a high-rise b uilding, the large vertical shafts used to transport the outdoor air and exhaust air often occupy a lot of valuable rental space. 2. When outdoor weather suddenly changes, the space air may also produce corresponding f uctuations. Building pressurization is not as stable as with a water economizer. 3. During the winter season, the intake of drier outdoor air al ways results in a lo wer space relative humidity than is experienced using recirculating air. Humidif cation may be required. 4. For packaged units, the f rst cost of an air economizer is about 25 to 40 percent of the installation cost of the packaged unit. F or a w ater economizer, the f rst cost is about 10 percent of the installation cost of the PU. 5. Coil freeze protection is more critical in locations where the outdoor temperature drops belo w 32°F (0°C). ASHRAE / IESNA Standard 90.1-1999 Economizer Control Specifications For the use of free cooling of outdoor air to reduce refrigeration, to improve indoor air quality, and to save ener gy use in HV AC&R systems ASHRAE / IESNA Standard 90.1-1999 speci f ed the following:

AIR SYSTEMS: VARIABLE-AIR-VOLUME SYSTEMS

21.15

Installation of Economizers. Each air system that has a fan and cooling capacity equal to or greater than that listed belo w shall include an air or w ater economizer and meet the follo wing requirements: No. of hours between 8 a.m and 4 p.m. with 55°F  Tdb  69°F

Twb  69°F, Btu / h

69°F  Twb  73°F, Btu / h

Twb  73°F, Btu / h

200 – 599 600 – 799 800 – 999 1000 – 1199  1199

135,000 65,000 65,000 65,000 65,000

N.R. 135,000 135,000 65,000 65,000

N.R. N.R. 135,000 135,000 65,000

Here, N. R. indicates “not required,” Tdb the dry-bulb temperature, and Twb the wet-bulb temperature. Exceptions include: ●

●

●

●

●

●

●

Air system used for the purpose of outdoor ventilation air cleaning to meet ASHRAE Standard 62 requirement. Where more than 25 percent of the supply air is used for space humidi f cation at abo ve 35 °F (1.7°C) dew point to satisfy process needs. Air systems that include a condenser heat recovery system. Air systems that serve residential spaces, or serve spaces whose sensible cooling load is rather small. Air systems expected to operate less than 20 hours per week. Where the use of outdoor air will affect supermarket open refrigerated display cases. When the cooling ef f ciency of the packaged unit or heat pump meet the follo wing increasing requirements: Cooling Degree Days (CDD50)

System Size, kBtu / h Qrc

Mandatory Minimum EER

3601– 5400

5401 – 7200

7201 – 9000

9001 – 10,800

65  Qrc  135 135  Qrc  240 240  Qrc  760

10.1 9.3 9.0

12.1 11.3 10.9

11.6 10.8 10.5

11.1 10.4 10.0

10.7 9.9 9.6

For other packaged systems, when mandatory minimum EERs v aries from 10.3 to 9.0, the corresponding minimum increased cooling ef f ciencies (EER) required are from 12.5 to 9.9. F or details, refer to ASHRAE / IESNA Standard 90.1-1999. Design and controls of economizers shall be such that economizer operation does not increase the building heating energy use during normal operation e xcept allowable reheating, recooling, and mixing by Standard 90.1-1999 (refer to Sec. 20.20). Air Economizers. ●

Standard 90.1-1999 specif es the following for air economizers:

Capacity contr ol. Air economizers shall be capable of modulating outdoor air and return air dampers to provide up to 100 percent of outdoor air for cooling. Air economizers shall be capable of automatically reducing outdoor air intake to the required minimum outdoor ventilation air specif ed by ASHRAE Standard 62 when outdoor air intake is no longer reducing cooling energy usage. Operators of air economizers using v arious types of temperature-based or enthalp y-based economizer controls shall be aware that for dry climates (shaded area B in Fig. 21.3a) minimum outdoor ventilation air (required by ASHRAE Standard 62) recirculating mode often needs less cooling energy than 100 percent outdoor air economizer operation. The 100 percent air economizer operation shall be shut of f. Similarly, for humid climates (dotted area A in Fig. 21.3 a) in which minimum

21.16

CHAPTER TWENTY-ONE

●

●

●

outdoor ventilation air recirculating mode often needs less cooling ener gy than 100 percent outdoor air economizer control and the air economizer operation shall also be shut off. Sequence control. Economizer dampers shall be in sequence with the refrigeration and shall not be controlled only by mixed air temperature except for single-zone air systems. Damper leakage. Both recirculating and outdoor air dampers shall ha ve a maximum leakage rate of 20 cfm / ft2 (100 L / sm2) at a pressure difference of 4 in. WC (1000 Pa). Relief of excess outdoor air. Means to relieve excess outdoor air during air economizer operation shall be pro vided to pre vent overpressurization of the b uilding. The relief outlet shall be located to avoid recirculation into the building.

Water Economizers. Water economizers shall be capable of cooling supply air through indirect evaporation and providing up to 100 percent of the system cooling coil load (precooling coil) at an outdoor air temperature of 50 °F (10°C) dry b ulb / 45°F (7.2°C) wet b ulb and belo w [except when the dehumidi f cation requirement cannot be met at these outdoor dry- and wet-b ulb temperatures and the 100 percent system cooling (precooling) coil load must be satis f ed at outdoor air temperatures 45°F (7.2°C) dry bulb / 40°F (4.4°C) wet bulb]. Precooling coils and w ater-to-water heat e xchangers as components of the w ater economizer shall either have a water-side pressure drop of less than 15 ft (44.7 kPa), or using a secondary water loop so that the pressure drop of the coil or heat e xchanger is not tak en care of by the circulating pumps when the system is in normal cooling (noneconomizer) mode. Air systems with w ater cooling and humidi f cation to maintain space humidity at greater than 35°F (1.7°C) dew point temperature shall use a w ater economizer if an economizer is required because the air economizer needs to e xtract outdoor air at lo wer temperature and humidity ratios and thus increases the humidifying load. Integrated Economizers Control. Economizers shall be integrated with refrigeration and can provide partial cooling and e ven additional refrigeration if required to meet the remaining cooling coil load, except the following: ●

●

●

Direct-expansion systems that control and reduce the required quantity of outdoor air to pre vent coil frosting at the lo west step of compressor unloading if this lo west step is no greater than 25 percent of the total system capacity. Individual packaged unit has a capacity less than 65,000 Btu / h (19 kW) and uses noninte grated economizer controls. System in locations with less than 800 average hours annually between 8 a.m. and 4 p.m. when the ambient (outdoor) dry-bulb temperatures are between 55 °F (12.8°C) and 69 °F (20.6°C) inclusive.

Air Conditioning Cycle and System Calculations The air conditioning c ycle of a single-zone VAV system during cooling mode operation at summer design conditions and heating mode operation at winter design conditions without space humidity control can be plotted on the psychrometric chart as sho wn in Fig. 21.1 b During summer cooling mode part-load operation, the state point of outdoor air o varies, and the sensible ratio of the space conditioning line SHR sp usually becomes smaller than at design load. If the zone temperature Tr is maintained at the same predetermined limit as in design load, the zone relati ve humidity will be slightly greater than that in design load if SHR sp becomes smaller. During winter heating mode partload operation, if the sensible heat ratio of space conditioning line SHR sp becomes smaller because of the reduced v olume f ow, the zone relative humidity will also be greater than that in design load. From Eq. (20.69), the design supply v olume f ow rate V˙s, in cfm [m 3 / (60 s)], can be calculated as V˙s 

Q rs 60 s cpa (Tr Ts)

(21.1)

AIR SYSTEMS: VARIABLE-AIR-VOLUME SYSTEMS

21.17

where Qrs  zone sensible cooling load, Btu / h (W), and Tr, Ts  zone and supply air temperature, °F (°C). In Eq. (21.1), [m3 /(60 s)] means that in SI units, the V˙s unit should be 160 m3 /s. From Eq. (20.48b), the design cooling coil load Qcc, in Btu / h (W), can be calculated as Q cc  60V˙s s(h m h cc)

(21.2)

where hm  enthalpy of the mixture of outdoor and recirculating air , Btu / lb (J / kg), and hcc  enthalpy of conditioned air lea ving the cooling coil, Btu / lb (J / kg). From Eq. (20.85), the design heating coil load Qch, in Btu / h (W), which offsets the space heating load or sometimes maintains a setback temperature during unoccupied periods to reduce w arm-up load ne xt morning and pre vent freezing, can be calculated as Q ch  60V˙s scpa(Thc Ten)

(21.3)

where Thc, Ten  air temperature leaving and entering heating coil, °F (°C).

Zone Temperature Control — Sequence of Operations Consider a single-zone VAV system using an AHU with a w ater cooling coil, a water heating coil, and a temperature economizer (air economizer), as shown in Fig. 21.1. The sequence of operations of this VAV system to maintain a preset zone temperature Tr is as follows: 1. When the AHU is in the off position, the outdoor and exhaust dampers are closed, the two-way valves of the cooling and heating coils are also closed, and supply and return fan motors are off. 2. When the AHU is in the on position, zone temperature sensor T1 senses the zone temperature Tr , sends a signal to the DDC controller , and compares it with the set point in summer of 75 °F (23.9°F) in the controller . If Tr > 75 °F (23.9 °C), the DDC controller calls for cooling — cooling mode operation. At the same time, the outdoor temperature To is measured by the outdoor temperature sensor T2. If To  75°F (23.9 °C) (a temperature economizer set point), the DDC controller opens the outdoor and e xhaust dampers to a minimum opening position, and fully opens the recirculating damper. The DDC controller also starts supply and return f an motors with speed control via variable speed drives. 3. The dischar ge air temperature Tdis is sensed by temperature sensor T4. During cooling mode operation, if the of f-coil temperature Tcc is 55 °F (12.8 °C), the set point of Tdis  55 2 (fan power)  57°F (13.9 °C). When Tdis  57°F (13.9 °C) is sensed by T4, the DDC controller opens and modulates the tw o-way valve of the w ater cooling coil so as to maintain Tdis  57°F (13.9°C). 4. If To  75°F (23.9°C), using the free cooling of outdoor air is al ways benef cial in order to save ener gy. The DDC controller fully opens the outdoor and e xhaust dampers and, at the same time, closes the recirculating damper . After the air economizer is ener gized, if Tdis is still higher than 57°F (13.9°C), the DDC unit controller will open and modulate the two-way valve of the water cooling coil to maintain Tdis  57°F (13.9°C). 5. Zone temperature Tr is sensed by a temperature sensor T1. This signal is compared with the zone temperature set point Tr,s, such as 75 °F (23.9°C) in the DDC controller . If Tr drops below or rises above 75°F (23.9°C), an output from the DDC controller actuates and modulates the supply fan speed, and thus the supply v olume f ow rate through an adjustable-frequenc y v ariable-speed drive, to maintain Tr  75°F (23.9 °C). F or single-zone VAV system serving a lar ge conditioned space, several temperature sensors can be located in dif ferent areas. Either the mean v alue or the highest or lowest temperature can be taken as sensed input to the DDC controller. 6. A pressure sensor P1 is used to sense the zone pressure pr in the conditioned space. If pr  0.03 in. WC (7.5 Pa), the DDC controller starts the return f an and modulates its f an speed to maintain a zone pressure pr  0.03 in. WC (7.5 Pa).

21.18

CHAPTER TWENTY-ONE

7. When To  55°F (12.8 °C) and the system is operated in cooling mode operation, controller ●

●

●

the DDC

Modulates the outdoor and recirculating damper to maintain a mixing temperature of 55 °F (12.8°C) and Tdis  57°F (13.9°C) with a volume f ow of outdoor air not less than the v entilation requirement Modulates the fan speed to maintain a zone temperature of 75°F (23.9°C) Closes the two-way valve of the water cooling coil

8. If the zone temperature sensed by sensor T1 is 72 °F  Tr  74°F (22.2°C  Tr  23.3°C), the system is operated in dead-band operation. Dead-band operation is neither a cooling nor heating mode operation. In dead-band mode, the DDC controller ●

●

Keeps both of the two-way valves of the cooling and heating coils closed Opens the outdoor damper to its minimum opening position and fully opens the recirculating damper to provide necessary outdoor ventilation air for the occupants

9. If the zone temperature Tr drops below 72°F (22.2°C), the DDC controller calls for heating, (heating mode operation). Also if To  55°F (12.8°C), then ●

●

The outdoor damper is opened to its minimum opening position to e xtract only the necessary amount of outdoor ventilation air. The DDC controller opens and modulates the tw o-way v alve of the heating coil to maintain a discharge air temperature during heating mode not higher than 15 °F (8.3°C) above the space temperature, say 72 15  87°F (30.6°C).

For the zone temperature control in heating mode operation, a reverse-acting relay is ener gized so that the control action is re versed when the cold air supply is changed o ver to w arm air supply . The DDC controller also modulates the supply fan speed to maintain a zone temperature Tr  72°F (22.2°C). During the abo ve zone controls for a single-zone VAV system, to determine whether the DDC controller calls for cooling or heating, either a zone temperature set point Tr,s or a speci f c outdoor temperature set point To,s can be used as the comparing criterion. Zone temperature set point Tr,s directly relates to the zone load and the actual zone conditions and is recommended. Because of the high internal load, a zone may still need a cold air supply at a rather low outdoor temperature. When a cooling mode is automatically changed to heating mode or vice v ersa, the width of the dead band greater than 3 °F (1.7°C) is not appropriate for thermal comfort. When two separate set points are used and heating and cooling mode operations are not automatically changed o ver, only then can a dead-band width of 5°F (2.8°C) or sometimes greater be used. To save energy as well as to increase the zone supply volume f ow rate, the discharge air temperature is often reset to a higher v alue during cooling mode part-load operation, such as from 57 to 60°F or to 63°F (13.9 to 15.6°C or to 17.2°C).

21.3 VAV COOLING, VAV REHEAT, AND PERIMETER HEATING VAV SYSTEMS VAV Cooling Systems A VAV cooling system is a multizone air system with VAV boxes to modulate the cold supply air volume f ow rate to match the v ariation of zone load to maintain a preset zone temperature, as shown by the air system that serv es the interior zone in Fig. 21.5 a. VAV cooling systems with VAV boxes only are widely used to serve the interior zone in commercial and industrial buildings.

FIGURE 21.5 A VAV reheat system: (a) schematic diagram; (b) air conditioning cycle.

21.19

21.20

CHAPTER TWENTY-ONE

VAV Reheat Systems A VAV reheat system is a multizone VAV system with either VAV boxes or reheating VAV boxes to modulate its cold supply v olume f ow rate for each of the control zones, to match the v ariation of the zone load to maintain a preset zone temperature. F or the interior zone in a b uilding, VAV boxes are used (VAV cooling); and for the perimeter zone in a b uilding, reheating VAV boxes are used (VAV reheat) in which a reheating coil is added to a cooling VAV box, as shown in Fig. 21.5. To minimize ener gy consumption, as co vered in Sec. 20.20, ASHRAE / IESNA Standard 90.11999 specif es the follo wing criteria: VAV reheat systems shall be capable of operating in sequence the supply of heating and cooling ener gy to the zone to minimize simultaneous cooling and heating except as follows: 1. 300 cfm (142 L / s), or a zone peak f ow rate total of no more than 10 percent of the total fan system f ow rate 2. The minimum required to meet ventilation requirements 3. 0.4 cfm / ft2 (2.03 L / sm2) of zone conditioned f oor area VAV reheating is a simple and ef fective means of maintaining preset space temperature with a lower relative humidity at cooling mode part-load operation and winter heating mode. F or details refer to Standard 90.1-1999.

System Description The mixture of outdoor and recirculating air is conditioned either in an AHU or in a PU. Both contain a supply f an, a return or relief f an, cooling coils, and medium-efficiency filters and are distributed to various control zones via supply ducts, runouts, reheating VAV boxes, VAV boxes, and outlets. During summer cooling mode operation, both the reheating VAV boxes in the perimeter zone and the VAV boxes in the interior zone modulate the zone supply v olume flow rate to maintain a predetermined zone temperature Tm. When the space sensible load in the perimeter zone drops during summer cooling mode part-load operation, the VAV box closes its opening until its minimum setting is reached. If the zone temperature Tm drops further and f alls below a certain limit, the reheating coil is ener gized and reheats the supply air to maintain a predetermined zone temperature. During winter heating mode operation, when one of the zone temperatures in the perimeter zone falls below a predetermined limit, the reheating coil in the reheating VAV box in this control zone is energized to heat the air to offset the space heating load to maintain a preset zone temperature. For a control zone in the interior zone, the zone sensible cooling load seldom drops belo w 30 percent of its peak load while the space is occupied, except during morning w arm-ups and weekends. Therefore, reheating VAV boxes are not used in the interior zone. In the interior zone, zone temperature Tm is also maintained by modulating the supply v olume f ow rate of the VAV box in each control zone. If an AHU or PU serv es both the perimeter and the interior zone, space air from the perimeter and interior zone is often returned to the same AHU or PU in the f an room through return slots, ceiling plenum, and connecting return ducts.

Perimeter Heating VAV System A perimeter heating VAV system consists of a hot w ater perimeter heating system to offset the winter transmission and in f ltration loss in the perimeter zone and a VAV cooling system with VAV

AIR SYSTEMS: VARIABLE-AIR-VOLUME SYSTEMS

21.21

FIGURE 21.6 A perimeter heating VAV system.

boxes to provide cooling and outdoor ventilation air for both perimeter and interior zones, as shown in Fig. 21.6. A perimeter heating system consists of hydronic hot w ater f nned-tube baseboard heaters with a zone control system. The baseboard heaters must be carefully zoned and controlled in coordination with the VAV boxes. Each side of the b uilding should have at least a separate control zone to match the variation of the solar load and the inf ltration loss as the wind direction shifts. Perimeter heating systems are suitable for commercial b uildings in locations with long, cold winters. During the early de velopment of VAV systems, air skin VAV systems were used in of f ce buildings. An air skin VAV system uses a constant-volume variable supply temperature air skin system to offset the transmission loss or transmission gain in the perimeter zone; it also has a VAV cooling system with VAV boxes to provide cooling and outdoor air for the perimeter and interior zones. The supply temperature of the air skin system is reset according to the outdoor temperature. If se veral air skin systems are not used for v arious zones of dif ferent sides of a b uilding, f eld surveys show that during a cold day in winter , rooms facing south are suf f ciently warmed by the incoming solar radiation. The additional warm air supply from the constant-v olume air skin system raises the zone temperature to an unacceptably high le vel. Simultaneous cooling from the VAV boxes must be provided to offset the warm air supply, which results in a waste of energy.

VAV Box A variable-air-volume box or a cooling VAV box is a terminal de vice in which the supply v olume f ow rate is modulated by v arying the opening of the air passage by means of a single-blade b utterf y damper, a multiblade damper, or an air valve, as shown in Fig. 21.7. A VAV box may have a single outlet or multiple round outlets. A single-blade damper VAV box, as shown in Fig. 21.7a and b, has a simple construction and is simple to operate. A typical damper closes at an angle 30 ° from v ertical and rotates in counterclockwise direction to an angle of 60° in the fully open position.

21.22

CHAPTER TWENTY-ONE

FIGURE 21.7 VAV box es: (a) single-blade, pressure-dependent; ( b) single-blade, pressure-independent; ( c) air valve.

An air v alve, as shown in Fig. 21.7 c, is a piston damper mo ving horizontally inside a hollo w cylinder. The opening of the air passage can be adjusted. The main advantage of an air v alve is its almost linear relationship between the modulated air v olume and the displacement of the piston damper. Control Mechanisms. The control mechanism of a VAV box can be classif ed as pneumatic or direct digital control. In a pneumatic control VAV box, the zone temperature is sensed by a pneumatic thermostat, and the rotation of the single-blade damper is actuated by pneumatic po wer. When the thermostat senses an increase in zone temperature, the increase of compressed air pressure on a diaphragm moves the actuator at a stroke of 1 to 1.5 in. (25 to 38 mm) and rotates the damper through the linkages, as shown in Fig. 21.7 b. Therefore, the air passage opens wider. The spring pushes the actuator back when the air pressure decreases. In a DDC VAV box, the temperature sensor sends a signal to a DDC terminal controller . It actuates the motorized operator , moves the actuator to a certain displacement, and opens the singleblade damper wider . The damper is closed by either spring force or the re verse rotation of the motorized operator. Usually, a PI control mode is used by such a DDC terminal controller.

AIR SYSTEMS: VARIABLE-AIR-VOLUME SYSTEMS

21.23

The pneumatic or direct digital control of an air valve is similar to that of a single-blade damper. A pneumatic control VAV box has a f aster response actuator and a lo wer initial cost than those of a DDC VAV box. However, pneumatic systems require more maintenance to pro vide clean, dry compressed air. More importantly , DDC VAV boxes provide many sophisticated control functions and can communicate with other DDC controllers and the central computer. There is also a type of self-po wered or system-po wered VAV box control. It requires a higher supply air pressure to actuate the single-blade damper or air v alve according to the input signal from the temperature sensor. A self-powered VAV box also has complicated control mechanisms, so its applications are very limited. Influence of Duct Static Pressure VAV boxes can be classi f ed as pressure-dependent or pressure-independent. In a pressure-dependent VAV box, the variation of the duct static pressure at the inlet of the VAV box caused by the opening and closing of the dampers connected to the same main supply duct inf uences the modulation of its supply volume f ow rate. When the pressure at the inlet varies, the air f ow often oscillates. Pressure-dependent VAV box es are least e xpensive. They are used when the duct static pressure is more stable and in places where there is no need for maximum and minimum limit control. A pressure-independent VAV box modulates its supply volume f ow rate regardless of the variation of duct static pressure at its inlet. A typical pressure-independent VAV box is shown in Fig. 21.7c. The temperature sensor resets the set point of the sensed v elocity, and the controller actuates the damper to a wider or a narro wer position based on the sensed signal from the v elocity probe and the reset set point. Even if the static pressure at the VAV box inlet v aries from 0.5 to 3 in. WG (125 to 750 P ag), the volume f ow is maintained according to the required v alue called for by the temperature sensor and the controller. Pressure-independent VAV boxes are widely used in VAV systems. The sizes of the VAV boxes made by one manuf acturer range from 04 to 20, with a corresponding volume f ow rate from 225 to about 4000 cfm (106 to 1887 L / s). The pressure drop of these VAV boxes at nominal maximum v olume f ow rate when the damper is fully open usually v aries from 0.2 to 0.5 in. WC (50 to 125 Pa). To provide the required amount of outdoor air to the conditioned space for occupants, a VAV box usually reduces its v olume f ow to a minimum setting, such as 30 percent of its design f ow, while the space is occupied. Reheating VAV Box A reheating VAV box is a VAV box with a reheating coil, as shown in Fig. 21.8. The reheating coil is usually a tw o- to three-ro w hot w ater heating coil. If an electric coil is installed, it should be a duct-mounted coil at least 4 ft (1.22 m) do wnstream from the VAV box. Federal and local safety codes must be follo wed if an electric coil is installed. To prevent excessive pressure drop at the reheating coil, the f ace v elocity of the reheating coil is usually between 200 and 400 fpm (1 and 2 m / s). For each row of the reheating coil, the additional pressure drop is usually between 0.1 and 0.2 in. WC (25 and 50 Pa). In the perimeter zone, a two-way slot diffuser should be located about 1 ft (0.3 m) from the window. Air should be discharged downward from one of the slots in a direction to ward the window, as described in Sec. 18.6. Air is discharged horizontally away from the window from another slot. Sound Power Level of a VAV Box The sound power level of a VAV box depends mainly on the following: ●

●

●

Volume f ow of supply air for a specif c size box, cfm (L / s) Difference in static pressure ps across the VAV box, in in. WC (Pa), to provide a specif c volume f ow, in cfm (L / s) The conf guration of the VAV box, f exible ducts, and diffusers

21.24

CHAPTER TWENTY-ONE

FIGURE 21.8 DDC for a reheating VAV box.

The greater ps, the larger the static pressure at the box inlet. The smaller the damper opening, the higher the sound power level of the VAV box. Noise generated by the VAV box can be transmitted to occupants in the conditioned space by a duct-borne path through f exible ducts and dif fusers, or a radiated path from its casing through the ceiling plenum. Evaluating noise levels from VAV boxes on an occupied zone has al ways been diff cult because of the man y v ariables that af fect the results. The ADC and ARI ha ve jointly de veloped Industry

AIR SYSTEMS: VARIABLE-AIR-VOLUME SYSTEMS

21.25

Standard 885, Acoustical Level Estimation Pr ocedure, as described in Chap. 19, to provide a reliable method to predict NC or RC in an occupied space. In Smith (1989), the sound pressure levels of two VAV boxes in a 29-ft 20-ft (8.8-m 6.1 m) room with an 8-ft 7-in. (2.6-m) ceiling were estimated and measured. One VAV box was operated at 500 cfm (236 L / s) at 4.0 in. WG (1000 Pa) inlet static pressure and supplied 250 cfm (118 L / s) to each of two diffusers through f exible ducts. The actual sound pressure le vel measured in the occupied space w as NC-41. Another VAV box w as operated at 600 cfm (283 L / s) with 3.0 in. WG (750 Pa) inlet static pressure and supplied 300 cfm (142 L / s) to each of tw o diffusers in the same room. The actual sound pressure le vel measured w as about NC-41. The measured sound pressure levels were about 2 to 3 dB higher or lo wer than the estimated duct-borne and radiated paths according to Industry Standard 885. If the sound po wer level at mid-frequenc y 2k and 4k Hz can be further attenuated by f exible duct and the slot dif fuser, the sound pressure le vel in the occupied space may drop to NC-35. The sound power level of a VAV box listed in a manuf acturer’s catalog must be tested in a certif ed laboratory.

VAV Reheat Zone Temperature Control and Sequence of Operations The year-round control of zone temperature serv ed by a reheating VAV VV1 box in the perimeter zone of a b uilding serv ed by a VAV reheat system can be di vided into cooling mode, dead-band mode, and heating mode operations. 1. On-off control is in the of f position. When an AHU or a PU that serv es this perimeter zone is shut off, the single-blade damper of the VAV box VV1 and the tw o-way valve V1 of the reheating coil both are closed. 2. On-off control is in the on position. When the zone temperature Trx1 sensed by the zone sensor exceeds a set point, such as 75 °F (23.9°C), the DDC system controller calls for cooling; the AHU or PU and the reheating VAV box are both in cooling mode operation. The DDC terminal controller opens the single-blade damper VV1 of the reheating VAV box to maximum opening position. The two-way valve of the reheating coil is still closed. 3. If the zone temperature Trx1p at cooling mode part-load operation drops belo w 75°F (23.9°C) because of a reduction in the zone sensible cooling load Qrsx1p, the DDC terminal controller receives a corresponding analog input from the zone sensor . This signal is compared with the set point 75 °F (23.9°C). If Trx1p  75°F, (23.9°C), the DDC terminal controller sends an output to the damper actuator and starts to reduce the opening of the f ow passage of the reheating VAV box. The total pressure loss across the VAV box is then increased, and so the v olume f ow rate of the conditioned air supply to this control zone V˙sx1p is reduced accordingly . If the reduced V˙sx1p is balanced with the zone sensible cooling load, Trx1p will remain at 75 °F (23.9°C). The DDC terminal controller then modulates the opening of the single-blade damper VV1 with a PI control mode to maintain a constant 75°F zone set point, as shown in Fig. 21.8. 4. Further drop in Qrsx1p causes damper VV1 to close more and more until it reaches the minimum setting which is necessary to supply the required amount of outdoor v entilation air for this zone, such as 30 percent of the zone peak f ow. After the damper reaches the minimum setting, if Qrsx1p further drops and causes Trx1p to drop below 75.0°F (23.3°C), the reheating VAV box will operate in dead-band mode. 5. When the zone temperature drops within the range 72.0 °F  Trx1p  75.0°F (22.2 °C  Trx1p  23.9°C), the reheating box and zone control operate in dead-band mode. In dead-band mode, the DDC terminal controller actuates as follows: ●

●

●

The damper in the reheating VAV boxes closes to the minimum setting, and the minimum cold primary air supply V˙sx1,min is at a v alue, such as 30 percent of the zone peak f ow, to provide outdoor ventilation air for the control zone. The reheating coil is deenergized. The space relative humidity will be slightly lower than that at the design condition.

21.26

CHAPTER TWENTY-ONE

In dead-band mode, the reheating box will reenter the cooling mode operation only if T  75°F (23.9°C). 6. When the zone temperature Trx1  72.0°F (22.2°C), the reheating box and the associated zone control are in heating mode operation. During heating mode operation, the DDC terminal controller actuates the reverse action relay so that the direct-acting mode will be changed to reverse-acting mode. The volume f ow rate of the cold primary air supplied from the reheating box in heating mode V˙s,cxn is still maintained in a minimum setting of 30 percent of zone peak v olume f ow and will remain constant to provide cold primary air for ventilation requirement. The DDC terminal controller opens and modulates the opening of the tw o-way valve of the reheating coil as well as the mass f ow rate of hot w ater f owing through the coil to maintain a preset zone temperature Trx1  72°F (22.2°C). The zone relati ve humidity may be at a v alue between 25 and 50 percent depending on the zone internal latent loads, outdoor air humidity ratio, and percentage of outdoor air intake. At winter design conditions, the warm supply air temperature from the reheating box (after the cold primary air is heated in the reheating coil) should not e xceed 87°F (30.6°C), in order to reduce the buoyancy effect of the warm supply air. The dead band in a year -round zone-temperature control for a perimeter zone serv ed by a VAV reheat system is dif ferent from the dead band in a dual-thermostat year -round zone temperature control system serv ed by a constant-v olume single-zone system, as discussed in Sec. 20.19. In a VAV reheat system, ●

●

The cooling mode is automatically changed over to heating mode operation and vice versa. The occupants may suffer the zone temperature v ariation from cooling mode to dead-band mode, and then to heating mode and vice versa, continuously within a time period of several hours.

The width of the dead-band mode in such a circumstance is preferably limited to 1 to 3 1.7°C).

°F (0.6 to

Stability of Zone Control Using VAV Boxes and Reheating VAV Boxes VAV boxes including reheating VAV boxes using DDC terminal controllers can pro vide the following functional controls: ●

●

●

●

Zone temperature control by modulating the damper or air valve Zone temperature reset based on outdoor temperature or zone temperature Maximum and minimum settings of airf ow for each zone Morning warm-up and cool-down control

A normally closed DDC VAV box can shut of f its supply volume f ow when the served control zone is not occupied. This is especially useful for partially occupied f oors after normal working hours. Air f owing through VAV boxes is self-balanced along the supply main duct. Ho wever, airf ow imbalance does af fect the modulation range of VAV box es. A well-designed supply duct system with minimized imbalance is recommended. One characteristic of VAV boxes that directly affects their performance is control stability. Dean and Ratzenberger (1985) described the primary dif ferences between an ef f cient and ineff cient design of a VAV box and its control. An excellent design minimizes the open-loop gain and impro ves the overall system stability. Gain is def ned as the real part of the logarithm of the transfer function of the control system. The primary factors that affect the open-loop gain and thus the stability of a VAV box are as follows: ●

Actuator size. In an eff cient design, the stroke of the actuator is small [typically 1.0 in. (25 mm)], and in a pneumatic system, the acti ve area of the actuator is typically only 2 in. 2 (1290 mm 2)

AIR SYSTEMS: VARIABLE-AIR-VOLUME SYSTEMS

●

●

21.27

instead of 8 in. 2 (5160 mm 2 ). A larger actuator needs more compressed air to dri ve a full-range stroke, so the gain is greater. Damper size. Gain is directly proportional to the size of the damper as well as the size of the VAV box. Duct pressure. A lower excess pressure pexc, in in. WC (Pa), during maximum and minimum airf ow affects open-loop gain. Excess pressure is de f ned as the amount of pressure abo ve that required to drive a nominal f ow through a fully open damper.

According to Dean and Ratzenber ger (1985), the typical measured performance of 26 ef f cient B-size VAV boxes and 27 ineff cient C-size VAV boxes is shown below:

Maximum f ow, cfm Corresponding pressure ps,max, in. WG. Minimum f ow, cfm Corresponding pressure, ps,min, in. WG Excess pressure, in. WC

B size

C size

442 1.07 64 1.24 0.08

448 1.40 66 1.59 0.62

For a step function response due to a step change in internal loads, such as turning lights on or of f, the deviation from set point in a pneumatic control system with a lar ge actuator lasts about 40 min. In a DDC system, it lasts only about 15 min. Based on their analysis, Dean and Ratzenberger (1985) recommended the following: ●

●

●

A VAV system using pneumatic control VAV box es should ha ve relatively low main duct pressures and relatively small VAV boxes and actuators. VAV boxes using DDC systems ha ve excellent operating characteristics, although their actuators are a little slower than pneumatic controls. Pressure-independent VAV box es with v elocity resetting impro ve the control stability of poorly designed pneumatic controls, make an unstable system more stable, and improve the performance of a poorly designed supply duct system.

Case Study: A VAV Reheat System Consider a multizone VAV reheat system using an AHU to serv e a typical f oor of 20,000 ft 2 (1858 m2) of conditioned area in a high-rise of f ce b uilding that is di vided nearly equally into perimeter and interior zones. Reheating VAV boxes are used for summer cooling and winter heating in perimeter zones. VAV boxes are used for summer and winter cooling in the interior zone when conditioned space is occupied. The perimeter and interior zones may each be divided into several to more than 20 control zones. The summer and winter design conditions are shown below:

Summer indoor temperature, °F Space relative humidity, percent Summer outdoor temperature, °F Summer outdoor wet-bulb temperature, °F Winter indoor temperature, °F Winter outdoor temperature, °F Winter indoor relative humidity, percent

Perimeter

Interior

75 43 95 78 72 20 30

75 45 95 78 72 20 30

21.28

CHAPTER TWENTY-ONE

Winter outdoor humidity ratio, lb / lb Summer block cooling load, Btu / h Summer sensible cooling load, Btu / h Winter heating linear density, Btu / h  ft Winter interior zone latent load, Btu / h Outdoor air required, cfm Summer supply system heat gain, °F Summer return system heat gain, °F

Perimeter

Interior

0.0017 300,000 285,500 150 13,300 1,920 5 3

0.0017 160,700 135,000 — 13,300 1,920 5 3

For the perimeter zone in a typical f oor of this of f ce building, the average space cooling load density at summer design condition is 30 Btu / h ft2 (94.6 W/ m2 or 8.8 W / ft2). The sensible heat ratio of the space conditioning line SHRs for the perimeter zone at design load is usually about 0.95. For the interior zone of a typical f oor of this off ce building, the summer design load density is 15.5 Btu / h ft2 (4.5 W / ft2 or 48.4 W / m2) depending on the equipment load density . Occupant density is assumed to be 150 ft 2 / person (14 m 2 / person). The sensible heat ratio of the space conditioning line SHRs for the interior zone may vary from 0.80 to 0.90. Conditioned Air Off-Coil and Supply Temperature Differential For a supply temperature differential of 20°F (11.1°C), if the zone supply temperature Ts  55°F (12.8°C), for a typical AHU with a supply f an po wer temperature rise of 2 °F (1.1 °C) and a supply duct temperature rise of 3°F (1.7°C) at summer design conditions, the temperature of conditioned air lea ving the cooling coil Tcc should be 55 (2 3)  50°F (10 °C). If the relati ve humidity of conditioned air off the coil cc is 95 percent, and hcc  20 Btu / lb (46.5 kJ / kg), then, from the psychrometric chart sho wn in Fig. 21.5 b, the relative humidity of the supply air s is 81 percent. Psychrometric analysis shows that to maintain a zone temperature of 75 °F (23.9°C) with an offcoil temperature of Tcc  50°F (10 °C) and cc  95 percent, a supply temperature Ts  55°F (12.8°C), and a sensible heat ratio of space conditioning line in the perimeter zone SHR sx of 0.95, or of interior zone SHR si of 0.84, and a supply temperature dif ferential Ts  20°F (11.1°C), the relative humidity of the perimeter zone must be rx  43 percent and the relati ve humidity of the interior zone ri  45 percent, as shown in Fig. 21.5b. System Volume Flow Rate and Coil Load For any control zone in the perimeter and interior zones, the VAV box, reheating VAV box, f exible ducts, and slot dif fusers should be sized according to the zone peak supply v olume f ow rate. The zone peak supply v olume f ow rate of an y control zone serv ed by a VAV box or a reheating VAV box V˙sn, in cfm [m3 / (60 s)], at summer design conditions can be calculated as V˙sn 

Q rsn 60 scpa(Tr Ts)

(21.4)

where Qrsn  zone sensible cooling load at summer design conditions, Btu / h (W). In a multizone VAV system, the total supply v olume f ow rate for the perimeter zone V˙sx, in cfm [m 3 / (60 s)], at summer design load is the maximum possible coincident total supply v olume f ow rate of v arious control zones at the same instant, or the block supply v olume f ow rate, which corresponds to the block load of the perimeter zone Qrsx, in Btu / h (W). The block supply v olume f ow rate in the

AIR SYSTEMS: VARIABLE-AIR-VOLUME SYSTEMS

21.29

perimeter zone at summer design conditions can be calculated as V˙sx 

Q rsx 60 scpa(Trx Ts)

285,500   13,054 cfm (6160 L / s) 60 0.075 0.243(75 55)

(21.5)

Similarly, the total block supply v olume f ow rate of the interior zone at summer design condition V˙si, in cfm [m / (60 s)], is V˙si 

Q rsi 60 scpa(Tri Ts)

135,000   6173 cfm (2913 L / s) 60 0.075 0.243 (75 55)

(21.6)

Assume that Qrsi and V˙si remain constant during occupied hours. The total supply v olume f ow rate for the AHU at summer design conditions is calculated as V˙s  V˙sx V˙si  13,054 6173  19,227 cfm (9073 L / s)

(21.7)

Trunks or main ducts that serv e multizones should accommodate the maximum possible air f ow through these ducts at a gi ven time. A diversity factor can be multiplied by the sum of indi vidual peak volume f ows to estimate the possible maximum f ow. From the psychrometric chart, the weighted mean space relati ve humidity for perimeter and interior zones rm  43.5 percent, and the weighted mean humidity ratio wrm  0.0079 lb / lb (kg / kg). Because the temperature of recirculating air entering the AHU Tru  75 3  78°F, and wru  wrm  0.0079 lb / lb, point ru can be plotted on the psychrometric chart. Dra w line ru-o. ru-m 3840   0.20 ru-o 19,227 so point m can then be determined: Tm  80.5°F (26.9°C), and hm  30.2 Btu / lb (70.2 kJ / kg). The cooling coil load Qcc in the AHU which is also a block load, in Btu / h (W), is Q cc  60V˙s s(h m h cc)  60 19,227 0.075(30.2 20)  882,519 Btu / h (258,578 W)

(21.8)

Summer Cooling Mode Part-Load Operation During summer cooling mode part-load operation, if the sensible cooling load in the perimeter zone is reduced to about 50 percent of its design load and the sensible cooling load of the interior zone is reduced to 85 percent of its design load, the total volume flow rate of the supply f an may decrease to about two-thirds of its design value. It is assumed that the supply fan power heat gain and the v olume flow are both reduced to 0.65 of the design v alues. The temperature rise at the supply fan at part load is still about 2°F (1.1°C). If the chilled water temperature entering the coil Twe is reset from 45°F (7.2°C) at design load to 48 °F (8.9°C) at about two-thirds part-load operation, the discharge air temperature increases to Tdis  50 2 48 45  55°F (12.8°C). The temperature rise from duct heat gain increases because of the reduction of the supply v olume flow rate. Because the increase of this temperature rise is proportional to the reduction of the supply volume flow rate, the supply air temperature for the perimeter zone at part load can be

21.30

CHAPTER TWENTY-ONE

calculated as Tsxp  50 2 3

3  61F (16.1C) 0.5

And the supply temperature for the interior zone at part load is Tsip  50 2 3 

3  58.5F (14.7C) 0.85

The supply volume f ow rate for the perimeter zone at part load is V˙sxp 

0.5 285,500  9325 cfm (4400 L / s) 60 0.075 0.243(75 61)

The supply volume f ow rate for the interior zone at part load is V˙sip 

0.85 35,000  6359 cfm (3001 L / s) 60 0.075 0.243(75 58.5)

At summer cooling mode part-load operation, the total supply volume f ow rate is V˙sp  V˙sxp V˙sip  9325 6359  15,684 cfm (7401 L / s) Assume the following conditions: ●

●

●

The sensible ratio of the perimeter zone at part-load SHRsxp drops to 0.91. The SHRs for the interior zone at part-load SHRsip remains the same. The temperature rise of the return system heat gain from the heat released by the light trof fer in the ceiling plenum Tretp, in °F ( °C), is proportional to the air v olume f ow; that is, Tretp  3 / (15,684 / 19,227)  3.7°F (2.1°C).

From the psychrometric chart, Trup  75 3.7  78.7°F (25.9°C). If, at summer cooling mode part-load operation, the required amount of outdoor air remains the same as at design load, the ratio of outdoor air to supply air at summer cooling mode part-load operation is therefore rup-mp 3840   0.25 rup-o 15,684 If the dry- and wet-bulb temperatures of outdoor air are the same as those at design conditions, then from the psychrometric chart Tmp  83°F (28.3°C), and hmp  31.9 Btu / lb (74.2 kJ / kg). As Tcc is raised to 53 °F (11.7°C) because of the 3 °F (1.7°C) reset in Twe, the enthalpy of conditioned air of f the coil at 53°F, 95 percent relative humidity is 21.6 Btu / lb (50.2 kJ / kg), and the coil load at summer cooling mode part-load operation is Q ccp  60V˙sp s(h mp h cc)  60 15,684 0.075(31.9 21.6)  726,953 Btu / h (212,997 W)

(21.9)

Winter Reheating in Perimeter Zone If there is only one AHU or PU to serve both the perimeter and interior zones of a typical f oor in a VAV reheat system, the air discharged from the AHU or PU at winter design conditions using an air economizer cycle must fulf ll the cold air supply requirement of the interior zone, typically at 52°F

AIR SYSTEMS: VARIABLE-AIR-VOLUME SYSTEMS

21.31

(11.1°C), as shown in Fig. 21.5 b. If the temperature rise from supply duct heat gain is 2 °F (1.1°C), the temperature of supply air entering the reheating coil in a reheating VAV box Ten is 54 °F (12.2°C). One of the primary considerations in the design of w arm air supply VAV systems (including VAV reheat system) at winter design conditions is the control of stratif cation when warm air is supplied from overhead slot and ceiling dif fusers in the perimeter zone. As discussed in Sec. 18.6, it is recommended that the supply air temperature dif ferential Tr Ts not exceed 15°F (8.3°C), to prevent e xcessive b uoyancy ef fects. A terminal v elocity of 150 fpm (0.75 m / s) is required at the 5-ft (1.5-m) level to offset the cold draft from the e xternal window. An even distribution of the volume f ow rate of supply air between the w arm airstream blo wing do wnward to ward the windo w glass and the warm airstream horizontally discharged toward the partition wall across from the window is preferable. At a warm air supply volume f ow rate equal to 30 percent of the summer peak v olume f ow rate of 1.5 cfm / ft2 (9.14 L / sm2), the space heating load per ft 2 of f oor area that can be offset by warm air supply in the perimeter zone during winter design conditions is qrh  0.3 1.5 60 0.075 0.243 15  7.4 Btu / h ft 2 (23.3 W / m2) A If the depth of the perimeter zone is 15 ft (4.6 m), the heating load linear density qh,ft is qh,ft  7.4 15  111 Btu / h ft (107 W / m) The maximum reheating coil load for any control zone in the perimeter zone to offset the zone heating load at winter design conditions Qchxn, in Btu / h (W), can be calculated as Q chxn  Q rhxn Q venxn  Q rhxn 60V˙s, cxn scpa (Trxn Ten, n)

(21.10)

where Qrhxn  space heating load for control zone n in perimeter zone, Btu / h (W) Qvenxn  ventilation load due to a 30 percent of peak volume f ow of cold supply air, Btu / h (W) V˙ s,cxn  volume f ow rate of cold primary air discharged to control zone from AHU of VAV reheating system, cfm [m3 / (60 s)] Ten,n, Trxn  cold primary air temperature and zone temperature of control zone n, °F (°C) For a summer peak volume f ow rate of 1.5 cfm / ft2 (9.14 L / sm2) in the perimeter zone, at 30 percent of peak volume f ow that means 0.3 1.5  0.45 cfm / ft2 (2.3 L / sm2) of cold supply air. If the zone heating load linear density qh,ft is 111 Btu / h  ft (107 W / m) and the zone heating load density is 7.4 Btu / h  ft2, the reheating coil load for 1 ft 2 (0.0929 m 2) of f oor area at winter design conditions can be calculated as Q chn  7.4 60 0.045 0.075 0.243(72 54)  7.4 8.9 A  16.3 Btu / h ft 2 (51.4 W / m2) If the AHU that serv es the perimeter zone in a VAV reheat system is separated from the interior zone, and a minimum outdoor air intak e of 15 percent of the peak supply v olume f ow rate can be set up for the perimeter zone during the winter heating mode operation, as shown in the air conditioning cycle in Fig. 21.1b, the reheating coil load is then Q chn  7.4 60 0.45 0.075 0.243(72 64)  7.4 3.9  11.3 Btu / h ft 2 (35.6 W / m2) A saving of about 8.9 3.9  5.0 Btu / h ft2 (15.7 W/ m2) is possible.

21.32

CHAPTER TWENTY-ONE

For the winter heating mode operation in the perimeter zone of a VAV reheat system, the following is recommended: ●

●

If there is only one AHU or PU to serve both the perimeter and interior zones in a VAV reheat system and if the zone heating load linear density qh,ft  120 Btu / hft (115 W / m), then the warm air supply v olume f ow rate from the reheating box in heating mode operation (at cold primary air minimum setting) should be proportionally increased to prevent excessive buoyancy effects. It is more ener gy-eff cient to use a perimeter heating VAV system; a dual-f an, dual-duct VAV system; or a f an-powered VAV system; or else the AHU or PU that serv es the perimeter zone is separated from another AHU or PU serving the interior zone.

Winter Cooling Mode Operation in Interior Zone At winter design conditions, the outdoor humidity ratio wo  0.0017 lb / lb (kg / kg), and the space sensible cooling load in the interior zone may still equal that of summer design conditions. Therefore, the supply temperature dif ferential for the interior zone is Tr Ts  20°F, or Ts  72

20  52°F (11.1 °C). The supply volume f ow rate for the interior zone is still V˙si  6173 cfm (2913 L / s). If the same AHU or PU is used for both perimeter and interior zones, a minimum supply volume f ow rate of cold primary air V˙sx,c that equals 0.3 of the peak supply v olume f ow rate (including the required outdoor ventilation) is supplied to the perimeter zone during winter heating mode and can be calculated as V˙sx,c  0.3 13,054  3916 cfm (1848 L / s) The difference in humidity ratio between supply and zone air is wr ws  

Q rl 60V˙s sh fg, o

13,300 13,300  0.00055 lb / lb (kg / kg) 60 (6173 3916) 0.075 1061

where wr, ws  humidity ratio of zone and supply air, lb / lb (kg / kg). The temperature rise from the return system heat gain at winter design conditions is then Tret,w  and

3  5.7F (3.2C) (6173 3916) / 19,227

Tru  Tr Tret,w  72 5.7 2  79.7F (26.5C)

If the temperature rise from the supply f an power is still about 2°F (1.1°C) and the temperature rise from supply duct heat gain is 3°F (1.7°C) for the interior zone, then Tm  Ts (2 3)  52 (2 3)  47°F (7.2°C) and

Tru Tm 79.7 47  Tru To 79.7 20 V˙o  0.55  V˙s

From the psychrometric chart, at a temperature Tr  72°F (22.2 °C) and r  30 percent, wr  0.005 lb / lb (kg / kg). Dra w line r-s from point r with SHR s  0.84, which intersects the

AIR SYSTEMS: VARIABLE-AIR-VOLUME SYSTEMS

21.33

72 20  52°F (10.0 °C) constant-temperature line at point s, and ws  0.005 0.00055  0.00445 lb / lb (kg / kg), as sho wn in Fig. 21.5 b. Because Tru  79.7°F (26.5 °C) and wru  wr  0.005 lb / lb (0.005 kg / kg), draw a line o-ru. Because om / o-ru  0.55, wm  0.0032 lb / lb (kg / kg). To maintain a space relative humidity of 30 percent at winter design conditions, a steam humidif er is required with a humidifying capacity m˙ s, in lb / h (g /s), calculated as follows: m˙s  60V˙s s(ws wm)  60(6713 3916)(0.075)(0.00445 0.0032)  56.8 lb / h (0.0072 kg / s) From the above analysis, when a VAV reheat system is used in an of f ce building, these primary problems may arise: ●

●

●

Outdoor ventilation air supply becomes insuf f cient during part-load operation. This is discussed in Chaps. 22 and 23. If an air economizer cycle is used for free cooling in the interior zone during winter cooling mode operation, for locations where the outdoor humidity ratio wo drops below 0.003 lb / lb (kg / kg), the zone relative humidity may fall below 25 percent, or even 20 percent. A steam humidif er may be used, as shown in Fig. 21.5. If the same AHU or PU serves both perimeter and interior zones, free cooling is used for the interior zone, and the mixed air in the perimeter zone contains the same percentage of outdoor air as in the interior zone, then a humidif er may be required for the perimeter zone otherwise the space relative humidity may even drop to 20 percent.

When a steam humidif er is installed in an AHU or PU, it is important that its humidi f er capacity be strictly controlled so that wet surf aces will not occur inside the AHU, PU, and supply ducts, which may cause indoor air quality problems.

21.4 DUAL-DUCT VAV SYSTEMS System Description A dual-duct VAV system uses two supply air ducts — a warm air duct and a cold air duct — to supply both warm air and cold air to various control zones in the perimeter zone. Warm air and cold air are mixed in the mixing VAV box in order to maintain a predetermined zone temperature, as shown in Fig. 21.9a. For control zones in the interior zone, only a cold air duct is used to supply cold air to offset the year-round zone cooling load through VAV boxes. A dual-duct VAV system can be a built-up system with a cold deck and a hot deck, or it can be a combination of two air-handling units. Packaged units are seldom used for dual-duct VAV systems because of their more complicated conf guration and controls. A dual-duct VAV system may condition and supply air separately to perimeter zones and interior zones. It is more f exible because the supply air temperature and the v ariation of supply v olume f ow rate can be matched better according to the zone load changes and outdoor v entilation air requirements. It is also more energy-eff cient than other VAV systems because of ●

●

Transfer of internal heat gains from the interior zone to the perimeter zone through the recirculating air Different outdoor air mixing ratio that can be used for interior and perimeter zones for free cooling

According to Warden (1996), a dual-duct VAV system is less expensive than a perimeter heating VAV system and is also less e xpensive than a VAV hot-water reheat system depending upon the design and conf gurations.

21.34

FIGURE 21.9 A dual-fan dual-duct VAV system: (a) schematic diagram; (b) air conditioning cycle.

21.35

21.36

CHAPTER TWENTY-ONE

It is possible to use the w arm air duct as a cold air duct in summer cooling mode operation so that there is a w arm / cold duct and a cold duct. Such an approach may reduce the size of the cold duct and the warm air leakages. However, changeover from cold air supply to warm air supply often produces many problems if the w arm / cold duct serv es control zones with dif ferent orientations in the perimeter zone. This approach is not advisable. Number of Supply Fans If an air economizer c ycle is used, a relief f an or a return f an must be installed to relie ve the space air when 100 percent outdoor air is used for free cooling. If the ceiling plenum is used as the return plenum, and if each typical f oor in a multistory building is served by one or two air-handling units, then the pressure drop of the return system is often less than 0.6 in. WC (150 P a) and a relief f an is a better choice than a return fan. The comparison between a return fan and a relief fan is made in Chap. 22. A dual-duct VAV system may be installed with a single supply f an that deli vers to both w arm and cold ducts, or two supply fans, one for the warm duct and one for the cold duct. Although a single supply f an is simpler and lo wer in initial cost, and the air -handling unit can be a f actory-made multizone unit rather than a f eld-assembled b uilt-up AHU, a dual-f an design has the follo wing advantages: ●

●

●

It allows the use of dif ferent ratios of outdoor air to supply air for the cold duct and the w arm duct. The warm air supply fan is operated only when it is required, such as during the warm-up period, or in spring, fall, and winter when the transmission and inf ltration loss is greater than the internal and solar loads in any control zone in the perimeter zone. A dual-fan design simplif es control and sa ves energy. Different control schemes are possible for cold and w arm air temperature resets, as well as for cold and w arm air duct static pressure controls.

Dual-fan dual-duct VAV systems are widely used in current dual-duct VAV systems design. Mixing VAV Box Figure 21.10 shows a pressure-independent mixing VAV box using a terminal DDC controller . This box consists of two separate, equal-size air passages arranged in parallel — one for warm air and the other for cold air . Each has a single-blade inlet mixing damper . These tw o air passages are then combined together, and the mixture of warm and cold air is dischar ged through a discharge volume damper, and then to the dif fusers through a f exible duct, single and multiple outlets, as shown in Fig. 21.10a. The DDC terminal controller controls the inlet mixing dampers directly according to the signal from the temperature sensor and controls the dischar ge volume damper from the signal sent by the airf ow sensor. The temperature sensor also resets the air f ow set point in the DDC controller. As in the VAV box, a pressure-independent mixing VAV box impro ves system stability and performance and is widely used. A currently a vailable DDC mixing VAV box has the follo wing controls: zone temperature and reset, minimum and maximum limit, minimum outdoor air , and warm-up and cool-do wn controls. Most DDC mixing VAV boxes use proportional plus integral (PI) control mode. The year-round operation of a mixing VAV box can be divided into cooling mode, mixing mode, and heating mode operations, as shown in Fig. 21.10 b. The cooling mode has the greatest supply volume f ow rate, and the cold air supply volume f ow rate V˙ sc, in cfm (L / s), reduces from 100 percent peak supply v olume f ow rate at summer design conditions do wn to the minimum setting. In the mixing mode operation, the reduction of the cold air supply v olume f ow rate exactly equals the increase of the w arm air supply v olume f ow rate; i.e., the total supply v olume f ow rate remains constant and is equal to the minimum setting. In the heating mode operation, the maximum w arm

1

2

3

Air flow sensor

2

1

F

Discharge volume damper

Inlet mixing dampers

T Cold

Warm

Zone temperature sensor

(a)

Peak supply volume flow rate, percentage

100 80 60 40 20 0 Heating

Mixing Tc,o

Cooling Th,o (b)

FIGURE 21.10 mode.

A pressure-independent mixing VAV box: (a) schematic diagram; (b) cooling, mixing, and heating

21.37

21.38

CHAPTER TWENTY-ONE

air supply v olume f ow rate V˙ sw, in cfm (L / s), may be v aried between 50 and 100 percent of the peak cold air supply volume f ow rate, depending on the heating load linear density qh,ft, in Btu / h ft (W / m), in the perimeter zone, so that the maximum zone supply air temperature dif ferential during heating mode operation does not e xceed 15°F (8.3°C). The warm air supply v olume f ow rate reduces as the zone heating load decreases until it is equal to zero in the mixing mode operation. As in a VAV reheat system, the minimum setting of zone supply v olume f ow from the mixing VAV box in a dual-duct VAV system should follo w the guidelines for VAV systems speci f ed in ASHRAE / IESNA Standard 90.1-1999. Mixing Mode Operation In Fig. 21.10b, the mixing mode starts at an outdoor air temperature Th,o when the warm air supply reduces to zero and ends at Tc,o when the cold air supply drops to zero, both in °F (°C). When outdoor temperature To  Th,o, the cold air supply v olume f ow V˙ scx1, p, in cfm (L/S) is reduced to a mixing-mode setting, such as, 30 percent of the cold peak v olume f ow rate. Similarly , when To  Tc,o, the warm air supply is also reduced to mixing-mode setting. Consider a mixing VAV box VXn serving a control zone rxn in the perimeter zone; its peak cold air supply v olume f ow rate is 13,054 / 10,000  1.3 cfm / ft2 (6.6 L / sm2), excluding the in f uence of w arm air leakage. If the maximum warm air supply volume f ow rate is equal to 70 percent of the peak cold air supply v olume f ow rate, also when the outdoor air temperature To  Th,o, if the cold deck dischar ge air temperature Tc,dis has been reset to 58 °F (14.4°C), for a duct heat gain equal to 1 °F (0.6°C), the zone sensible cooling load is then Qrsxn  0.3 1.3 60 0.0075 0.243(75 59)  6.82 Btu / h  ft2 (21.5 W/ m2) When outdoor air temperature To  Tc,o, the control zone rxn that the mixing VAV box serv es has the following characteristics: ●

●

●

There is only warm air supply except cold air leakage. Zone temperature should be reset to a v alue for heating mode operation, such as 72 °F (22.2°C). Heating from hot water, electric heater or gas heater is usually required.

When To  Tc,o, if the hot deck dischar ge air temperature Th,dis  80°F (26.7°C), the warm air supply can offset a zone heating load Qrhxn  0.3 1.3 60 0.075 0.243(80 72)  3.4 Btu / h ft2 (10.7 W/ m2) For a dual-duct VAV system with a mixing VAV box, there may or may not be a dead band between the changeover from the cooling mode to heating mode through mixing mode operation, as shown in Fig. 21.10b. Zone Controls and Sequence of Operations of Dual-Fan Dual-Duct VAV System Consider a mixing VAV box VX1 and a VAV box VI1 in a dual-fan dual-duct VAV system, as shown in Fig. 21.9, that serves the perimeter zone and interior zone of a typical f oor in an off ce building with a summer indoor temperature of 75 °F (23.9°C) and a winter indoor temperature of 72 °F (22.2°C). The following are the zone controls and the sequence of operations of this dual-f an dual-duct VAV system: 1. When time-of-day clock signals the AHUs of the dual-fan dual-duct VAV system to “off,” the mixing VAV box VX1 and the VAV box VI1 are closed. 2. When the clock signals the AHU in the on position, the zone temperature sensor T1 measures the zone temperature Trxn in the perimeter zone. If Trxn exceeds a predetermined limit, say,

AIR SYSTEMS: VARIABLE-AIR-VOLUME SYSTEMS

21.39

77°F (24.4°C), the AHUs and the mixing VAV box VX1 and VAV box VI1 are in cooling mode operation. The DDC terminal controller signals VX1 and VI1 to peak air f ow supply to the serv ed control zones, and the DDC system controller signals the AHU to reset discharge air temperature to maximum cooling set point, say, 52°F (11.1°C). 3. If the outdoor air temperature To exceeds a temperature economizer set point, that is, To  75°F (23.9°C), the DDC system controller closes the outdoor and e xhaust dampers to a minimum opening to e xtract a minimum amount of outdoor v entilation air , fully opens the recirculating damper in both hot and cold decks, and opens and modulates the tw o-way valve of the w ater cooling coil in the cold deck to maintain a required discharge air temperature Tdis. At summer design conditions, the refrigeration capacity needed to produce chilled w ater is 100 percent full capacity . At the same time, the zone temperature sensors sense the zone temperature Trx1 and Tri1, and the airf ow sensors measure the air f ow at the dischar ge outlet of the mixing VAV box in the perimeter zone and the VAV box in the interior zone. In the control zone X1 of the perimeter zone, the DDC terminal controller modulates and controls the inlet mixing dampers according to the sensed temperature signal and the reset signal from the air f ow sensor to maintain a required zone temperature of 75 °F (23.9°C). The cold air supply v olume f ow rate is gradually reduced as the zone sensible cooling load decreases. In the control zone I1 of the interior zone, the DDC terminal controller modulates the damper opening of the VAV box VI1 to maintain a zone temperature of 75 °F (23.9°C). The cold primary air supplied from the VAV box is gradually reduced as the zone sensible cooling load decreases at cooling mode part-load operation. In the cold deck, the dischar ge air temperature Tc,dis  52°F (11.1°C); and in the hot deck, the discharge air temperature Th,dis  81°F, which is the mixing temperature of recirculating air (85 percent of air f ow) at 78°F (25.6°C) and an outdoor air temperature (15 percent of air f ow) of 95 °F (35 °C) at summer design conditions. The heating coil in the hot deck is not energized. 4. When 50 °F  To  75°F (10.0 °C  To  23.9°C), the DDC system controller fully opens the outdoor and e xhaust dampers and closes the recirulating damper to e xtract 100 percent outdoor air for free cooling in the cold deck. Refrigeration is at a capacity signi f cantly less than the fullload design capacity and is still required to maintain a cold deck dischar ge air temperature. In the hot deck, the outdoor damper remains in the minimum opening position, and the recirculaing damper is fully opened to provide a warm air supply. In the mixing VAV box, the terminal DDC controller modulates the inlet mixing dampers and the discharge volume damper according to the sensed zone temperature and air f ow, to maintain a required zone temperature of 75 °F (23.9 °C). The volume f ow rate of cold air supply further reduces as the space cooling load decreases. The cold-deck discharge air temperature remains at 52 °F (11.1°C) in order to meet the requirement of a 15 to 17 °F (8.3 to 9.4 °C) supply temperature dif ferential in the interior zone. Ho wever, the cold supply air temperature from the mixing box may be 3 to 5 °F (1.7 to 2.8°C) higher because of the duct and f an power heat gains at part load. The hot deck dischar ge air temperature may be varied from 80 to 76°F (26.1 to 24.4°C) because of the variation of the outdoor air temperature. The heating coil in the hot deck is not energized. 5. If To  Th,o  50°F (10.0°C), the DDC system controller positions the outdoor and recirulating dampers in both the cold deck and hot deck, f rst, to provide an outdoor air intake at a value not less than the minimum ventilation requirment and, second, to maintain a mixing temperature Tm  50°F (10°C) if it is possible. The DDC system controller also closes the tw o-way valve of the w ater cooling coil in the cold deck of the AHU, so that no refrigeration is required for cooling mode operation. At To  Th,o, when the cold air supply in the mixing VAV box V˙ s,cx1 is reduced to mixing mode setting 30 percent of the peak cold air supply v olume f ow rate V˙ scx1,d, the mixing VAV box is in mixing mode operation. The mixing operation ends at the w arm air supply V˙ shx1  0.3V˙ shx1,d. In mixing mode, the DDC terminal controller for control zone X1 in the perimeter zone ●

Gradually resets the zone temperature from Trx1  75°F (23.9°C) at V˙ s,cx1  0.3V˙ scx1,d (To  Th,o) to Trx1  72°F (22.2°C) at V˙ shx1  0.3V˙ scx1,d (To  Tc,o)

21.40

CHAPTER TWENTY-ONE ●

Modulates the dischar ge volume damper of the mixing VAV box so that the total supply v olume f ow rate is maintained at mixing mode setting 30 percent of the peak cold air supply V˙ scx1,d during the mixing mode operation

The cold deck discharge air temperature Tc,dis is still 52°F (11.1°C). If the cold air supply in the perimeter zone drops nearly to its minimum setting (30 percent of the peak supply v olume f ow rate), the duct and f an power heat gain may ha ve a temperature rise of 10 °F (5.6°C) which raises the cold supply air temperature in the perimeter zone to about 60°F (15.6°C). If the hot-deck dischar ge air temperature Th,dis cannot be maintained at a v alue 5 °F (2.8 °C) higher than the zone temperature Trxn, the DDC system controller opens the tw o-way valve of the hot water coil and maintains a Th,dis  Trxn 5. And Th,dis increases as To falls. The relationship between the outdoor air temperature To and Th,o as well as To and Tc,o depends on the characteristics of the zone load and the construction of the building shell. 6. When the outdoor air temperature To  Tc,o, the mixing VAV box is in heating mode operation. In heating mode operation, the DDC system controller positions the outdoor and recirculating dampers to pro vide required minimum v entilation as well as to maintain a mixture temperature of 50°F (10°C) if possible. The cold deck discharge air temperature remains at 52°F (11.1°C). Also the two-way valve of the water cooling coil remains closed. If the duct heat loss is about 2 °F (1.1 °C), the hot deck dischar ge air temperature is equal to Th,dis  72 15 2  89°F (31.7°C). The DDC system controller modulates the tw o-way valve of the hot water coil in the AHU to maintain Th,dis  89°F (31.7°C). In the mixing box X1 in the perimeter zone, the DDC terminal controller modulates the inlet mixing dampers and the dischar ge volume damper according to the sensed zone temperature and airf ow signal, to maintain a preset zone temperature of 72 °F (22.2°C). As the zone heating load increases, the warm air supply volume f ow rate increases accordingly to a maximum value at winter design conditions with a w arm air supply temperature dif ferential not exceeding 15°F (8.3°C).

Discharge Air Temperature Control In general, the discharge air temperature from the cold deck must meet the year -round cooling requirements for the interior zone. An air economizer c ycle using free cooling should be applied to the cold deck whene ver possible, to meet the outdoor air requirements and to minimize refrigeration in order to sa ve ener gy. In the hot deck, internal heat gains carried by the recirculating air should be fully utilized. Cold deck dischar ge air temperature Tc,dis is often reset according to zone demands at part load or outdoor air temperature by means of a DDC unit controller . Hot deck dischar ge air temperature Th,dis is often reset by outdoor temperature during winter heating mode. Figure 21.11 sho ws the cold and w arm deck dischar ge air temperature from the AHU of a typical dual-f an dual-duct VAV system at v arious outdoor temperatures. The cold deck discharge air temperature Tc,dis is maintained at 52 °F at all outdoor air temperatures, in order to meet high zone cooling loads in the interior zone. Ho wever, when the cold air supply v olume flow rate in the perimeter zone reduces to about 30 percent of its peak v olume flo w during part-load operation, the temperature rise because of the f an power and duct heat gain may increase to 10 °F (5.6 °C), which results in a supply air temperature of cold air up to 60 °F (15.6°C). When the outdoor air temperature To drops from 70 to 60°F (21.1 to 15.6°C), the mixing of minimum outdoor air and recirculating air lo wers the hot deck dischar ge air temperature Th,dis from 80 to 76°F (26.7 to 24.4 °C). If To  50°F (10.0°C), heating may be required in the perimeter zone, so Th,dis increases as To falls. If the duct heat loss is 2 °F (1.1°C), the warm air supply temperature differential Th,dis 2 Trxn should not exceed 15°F (8.3°C) even at winter design conditions, in order to avoid stratif cation.

AIR SYSTEMS: VARIABLE-AIR-VOLUME SYSTEMS

21.41

120

Discharge air temperature Tdis, F

Heating mode

Cooling and mixing modes

100

80 Hot deck discharge air temperature 60

Cold deck discharge air temperature Outdoor air mixing, refrigeration turned off

40

20

0

FIGURE 21.11

Air economizer

40

60

Refrigeration

80 100 Outdoor temperature To, F

Warm and cold deck discharge air temperature in a dual-fan dual-duct VAV system.

Zone Supply Volume Flow Rate For each control zone in the perimeter zone serv ed by a mixing VAV box, an air leakage of 0.03 to 0.07 of its peak supply v olume f ow should be considered for the shut of f damper. If the difference in air density between w arm and cold air is ignored, the additional supply v olume f ow rate V˙ lk required to compensate for the air leakage when the w arm air damper is shut, in cfm (L / s), can be calculated as V˙lk  0.05V˙sxn

Ts,h Trxn Trxn Ts,c

(21.11)

where Ts,h, Ts,c  temperature of warm and cold air supply at mixing VAV box, °F (°C) Trxn  zone temperature, °F (°C) As in a VAV reheat system, for any control zone in either the perimeter or interior zones, the mixing VAV box, VAV box, f exible ducts, and slot or ceiling diffusers should be sized according to the peak supply v olume f ow rate. If there is no air leakage, the control zone peak supply v olume f ow rate V˙rxn, in cfm [m3 / (60 s)], is V˙rxn 

Q rsxn 60 scpa(Trxn Ts,c)

(21.12)

21.42

CHAPTER TWENTY-ONE

where Qrsxn  control zone maximum sensible cooling load, Btu / h (W). The control zone peak supply volume f ow rate for cold air in the perimeter zone at summer design conditions is V˙cxn  V˙rxn V˙lk

(21.13)

For any control zone in the perimeter zone, the peak supply v olume f ow rate V˙sxn delivered to the conditioned space from the mixing VAV box, including 5 percent air leakage from the w arm air damper at summer design conditions, in cfm (L / s), is therefore V˙sxn 

V˙cxn 0.95

(21.14)

Because only cold air from the VAV box is supplied to the interior zone, for any control zone in the interior zone, the peak supply v olume f ow rate V˙sin at summer design conditions, in cfm (L / s), is then V˙sin  V˙rin

(21.15)

Peak supply volume f ow rate for control zone n, V˙sin, can be calculated by Eq. (21.12). Case Study: A Dual-Fan Dual-Duct VAV System Consider a dual-f an dual-duct VAV system serving a typical 20,000-ft 2 f oor in an of f ce building with the same operating parameters as the VAV reheat system described in Sec. 21.3, except the winter outdoor design temperature is 30 °F ( 1.1°C) and the outdoor humidity ratio is 0.003 lb / lb (kg / kg). As in the VAV reheat system, the maximum supply v olume f ow rate for the interior zone at summer design conditions based on block load is V˙si  

Q rs 60 scpa(Tr Ts,c) 135,000  6173 cfm (2913 L / s) 60 0.075 0.243(75 55)

For a control zone in the perimeter zone, if the temperature of the mixture of outdoor and recirculating air Tmx  80.5°F (26.9 °C), humidity ratio wmx  0.010 lb / lb (kg / kg), Tsx  Ts,c  55°F (12.8°C), and sx  s,c  80 percent, draw line sx-mx as shown in Fig. 21.9b. For a warm air leakage of 0.05 V˙sxn from the inlet damper , the state point of the mixture of cold air supply and w arm air leakage sxn can be determined. From the psychrometric chart, Tsxn  56.5°F (13.6°C), and wsxn  0.0083 lb / lb (kg / kg). The volume f ow rate of cold air supply to the perimeter zone at summer design conditions, based on the block load and including 5 percent air leakage from the warm air inlet damper, is therefore V˙scx 

285,500  14,113 cfm (6660 L / s) 60 0.075 0.243(75 56.5)

The total cold air supply v olume f ow rate to both the perimeter zone and the interior zone at summer design conditions is V˙s,c  V˙scx V˙sci  14,113 6713  20,826 cfm (9828 L / s)

(21.16)

The volume f ow rate of warm air supply in the perimeter zone is usually e xpressed as a percentage of peak supply v olume f ow rate of cold air in the perimeter zone, usually between 0.5 V˙scx and 1.0V˙scx. It depends mainly on the space heating load linear density qh,ft, in Btu / h ft (W / m), and the

AIR SYSTEMS: VARIABLE-AIR-VOLUME SYSTEMS

21.43

supply temperature dif ferential Ts Tr). If the v olume f ow rate of the w arm air supply in the perimeter zone at winter design condition V˙shx, in cfm (L / s), is 0.7V˙scx, then V˙shx  0.7 14,113  9879 cfm (4661 L / s) Other parameters such as cooling and heating coil loads can be calculated as for VAV reheat systems.

Winter Heating and Winter Cooling Mode Operation In a dual-fan dual-duct VAV system, the cold deck and the hot deck ha ve their own outdoor and recirculating air supply. In the hot deck during winter heating mode operation, it is possible to ha ve an outdoor air ratio of only about 22 percent and a mixing temperature of outdoor and recirculating air Tmx  62°F (16.7°C) for the perimeter zone, so that only necessary heating ener gy is pro vided for a hot deck dischar ge air temperature of 87 °F (30.6°C), as shown in Fig. 21.9 b. In the cold deck winter cooling mode operation, nearly 60 percent outdoor air is required to form a mixture of outdoor and recirculating air with a temperature Tmi of about 50°F (10.0°C) and a cold deck dischar ge air temperature of 52 °F (11.1°C), so that outdoor air free cooling can be fully utilized to of fset the zone cooling load in the interior zone in winter.

Part-Load Operation Consider a typical control zone in the perimeter zone of this dual-f an, dual-duct VAV system whose operating parameters at summer design condition and summer part-load operation are as follows:

Zone sensible cooling load, Btu / h Leaving cooling-coil temperature, °F Relative humidity, percent Supply system heat gain, °F (°C) Warm air supply temperature, °F (°C)

Full load

Part load

14,200 50 95 5 (2.8) 87 (30.6)

4,260 50 95 10 (5.6) 80 (26.7)

If the supply air temperature dischar ged from the mixing VAV box Tsxn  Tm2  56.5°F (13.6°C), the volume f ow rate of conditioning air supplied to this control zone V˙sxn, in cfm (L / s), including 5 percent warm air leakage from the damper at summer design conditions, is therefore V˙sxn 

14,200  702 cfm (331 L / s) 60 0.075 0.243(75 56.5)

During part-load operation, as the zone sensible cooling load is reduced to 30 percent of the design load, Tsxnp  50 10  60°F (15.6°C). Assuming that the percentage of w arm air leakage and the temperature rise due to this w arm air leakage are the same at design condition, the supply temperature of the mixture of cold and warm air of control zone xn in the perimeter zone at part-load Tsxnp is now equal to 50 10 1.5  61.5°F (16.4°C). The supply volume f ow rate from the mixing VAV box in the perimeter zone at part-load can be calculated as V˙sxnp  

Q rsxnp 60 cpa (Trxn Tsxnp) 4260  289 cfm (136 L / s) 60 0.075 0.243(75 61.5)

(21.17)

21.44

CHAPTER TWENTY-ONE

FIGURE 21.12 Air conditioning cycle of a typical control zone at summer part-load operation served by a mixing VAV box of a dual-fan dual-duct VAV system.

From Eq. (21.14), the volume f ow rate of cold air supply at part load can be calculated as V˙s,cn  0.95 289  275 cfm (130 L / s) Also the warm air supply is V˙s,hn  289 275  14 cfm (6 L / s) Figure 21.12 sho ws the air conditioning c ycle of a typical control zone in the perimeter zone served by a dual-f an dual-duct VAV system at summer cooling mode part-load operation. From the psychrometric chart, zone relative humidity at summer part-load operation is higher than at fullload condition. This is mainly because of the increase of the cold deck dischar ge air temperature due to the duct heat gain and the possible chilled water reset at part load. At winter heating mode part-load operation, the state point of supply air will mo ve along the horizontal line mx-hc as shown in Fig. 21.9b, to maintain a predetermined zone temperature.

21.5 FAN-POWERED VAV SYSTEMS System Description A f an-powered VAV system, as sho wn in Fig. 21.13, often emplo ys f an-powered box es in each control zone of the perimeter zone and VAV boxes in each control zone of the interior zone in a

FIGURE 21.13 A parallel fan-powered VAV system: (a) schematic diagram; (b) air conditioning cycle; (c) details of winter heating mode air conditioning c ycle; (d) cold primary air and heating coil output diagram.

21.45

FIGURE 21.13

21.46

(Continued)

AIR SYSTEMS: VARIABLE-AIR-VOLUME SYSTEMS

. Vs, 100% cfm

Heating

Dead band

21.47

Cooling

Minimum 0

75 Cold primary air

Tr , F

Output 100%

0

72 Heating coil (d)

Tr , F

FIGURE 21.13 (Continued)

building. Both fan-powered boxes and VAV boxes can provide cooling for the perimeter and interior zones. The function and advantages of a fan-powered box include the following: ●

●

●

●

It e xtracts w arm recirculating air from the ceiling plenum to maintain a preset zone temperature when the volume f ow of the cold primary air supplied from the fan-powered box has dropped to the minimum setting or other preset v alue at summer cooling part-load operation. This is the f rst step. If the zone temperature drops further in an y of the control zones in the perimeter zone, a reheating coil in the f an-powered box is ener gized to maintain predetermined zone temperature during heating mode operation. This is the second step. It can increase the w arm air supply v olume f ow rate to meet the requirement that the w arm air supply temperature differential does not exceed 15°F (8.3°C), in order to prevent excessive buoyancy effects. It mixes the low-temperature primary air with w arm plenum air to pre vent the dump of a cold air jet and possible surface condensation in cold air distribution.

If the fan-powered box is used for mixing lo w-temperature primary air with warm plenum air in cold air distrib ution, fan-powered box es can also be installed in the control zones of the interior zone. The disadvantages of using a f an-powered box instead of a VAV box include higher ener gy use, higher noise level, and greater initial cost and maintenance.

21.48

CHAPTER TWENTY-ONE

Fan-Powered VAV Boxes A fan-powered VAV box, or simply a fan-powered box, consists of a small forward-curved centrifugal fan, a VAV box, a heating coil (hot w ater or electric heater), a low-eff ciency disposable f lter, dampers, an outer casing, and corresponding controls, as shown in Fig. 21.14. F an-powered VAV boxes can be classi f ed as series and parallel box es according to their construction and operating characteristics. Series Fan-Powered Box. In a series fan-powered box, a small centrifugal fan is connected in series with the cold primary airstream f owing through the VAV box, as shown in Fig. 21.14 a. The volume f ow of the cold primary air is varied at part-load operation, but after this air mixes with the induced plenum recirculating air , the fan delivers a continuously operated, nearly constant-volume f ow rate of the mixture of cold primary air and plenum recirculating air . The mixture is then supplied to the control zones, either heated at the heating coil or without heating depending upon the zone load. Because the airf ow from the series fan-powered box is constant, it is easier to select ceiling diffusers to achie ve good space air dif fusion. In a series f an-powered box, cold primary air is thoroughly mixed with plenum recirculating air . The main disadvantage of a series f an-powered box is its larger fan, which consumes more energy than a parallel fan-powered box. Parallel F an-Powered Bo x. In a parallel f an-powered box, the induced recirculating plenum airstream dischar ged from the centrifugal f an is connected in parallel with the cold primary airstream discharged from the VAV box in the fan-powered box. These two airstreams are mixed either before or after the heating coil and then are supplied to the control zones. A parallel f an-powered box with a heating coil located upstream from the f an is called a dra wthrough parallel fan-powered box, as shown in Fig. 21.14b. When the heating coil is located do wnstream from the f an, the box is called a blo w-through parallel f an-powered box, shown in Fig. 21.14c. The air temperature entering the heating coil of a dra w-through unit is the temperature of recirculating plenum air, which is higher than that of the mixture of cold primary air and recirculating plenum air in a blow-through box. A parallel fan-powered box operates only when there is a need to mix the w arm plenum air with the cold primary air at summer cooling mode part-load operation, when the volume f ow of the cold primary air is at its minimum setting or other preset v alue, or when the mixture of cold primary air and warm plenum air is heated at the heating coil to of fset the zone winter heating load. When a parallel f an-powered box is used to mix 40 to 50 °F (4.4 to 10.0 °C) low-temperature primary air with warm plenum air in a cold air distribution system, it is also operated continuously during summer cooling mode. In a parallel fan-powered box when the cold primary airstream is reduced to its minimum setting during cooling mode part-load operation, the volume f ow rate of the extracted plenum recirculating air plus the cold primary air is usually equal to 0.80 to 1.0 of the zone summer peak supply v olume f ow rate V˙sn,d . The f an in a parallel f an-powered box e xtracts only recirculating plenum air , and therefore it has a smaller v olume f ow rate and a lo wer fan total pressure than that in a series f anpowered box. Because of its smaller f an size, it consumes less f an po wer in cooling mode operation. Current parallel f an-powered box designs also pro vide a good mixture of cold primary air and extracted plenum air. Some of the recent models use a variable-speed small parallel fan. It energizes as soon as the supply v olume f ow of the cold primary air drops at cooling mode part-load operation. The extracted recirculating plenum air gradually increases as the v olume f ow of cold primary air decreases up to a maximum limit between 0.4 V˙sn,d and 0.5 V˙sn,d . Under this condition, the zone supply volume f ow rate is k ept nearly constant. P arallel fan-powered boxes are often used in cold air distribution because of their lower initial cost and lower fan energy consumption. In a parallel f an-powered box, a backdraft damper should be installed upstream from the f an. This damper is normally closed to pre vent the backward f ow of cold primary air in case the f an is deenergized. When the fan is turned on, the backdraft damper is fully opened.

FIGURE 21.14 Fan-powered box es: (a) series box; ( b) parallel box, draw-through; (c) parallel box, blow-through.

21.49

21.50

CHAPTER TWENTY-ONE

The drawbacks of a parallel f an-powered box include the need for a careful balance of f ow and total pressure of two airstreams, the diff culty in selection of the size of the box, and its more complicated controls. The size and capacity of a f an-powered box is determined by the capacity of the f an, the VAV box, and the heating coil. The heating coil in a fan-powered box can be a hot water coil or an electric duct heater located do wnstream from the f an-powered box. F an-powered box es are usually built in sizes corresponding to the associated VAV box primary air v olume f ow rates from 300 to 3200 cfm (150 to 1600 L / s) and f an supply v olume f ow rates of 250 to 4000 cfm (125 to 2000 L / s). Most fan-powered boxes have a fan total pressure between 0.5 and 0.8 in. WC (125 and 200 Pa) and an external pressure loss of about 0.25 in. WC (63 Pa) to offset the pressure drop of downstream f exible ducts and diffusers. The pressure drop of the VAV box and the corresponding pressure drop of f exible ducts and diffusers due to cold primary air should be of fset by the supply f an in the primary air AHU or PU. The speed of the centrifugal fan in the fan-powered box can be adjusted to high, medium, or low or even variable-speed drive. Both radiated and duct-borne sound paths should be check ed to determine whether the y can meet the space air required NC criteria. Because the f an total pressure is higher in a fan-powered box than in a fan-coil unit, a temperature rise due to fan power heat gain of 0.75°F (0.4°C) should be considered in psychrometric analysis. The VAV box a in f an-powered box can be either pressure-dependent or pressure-independent. Pressure-independent fan-powered boxes are widely used.

Fan Characteristics in Parallel Fan-Powered Boxes In a f an-powered VAV system using a parallel f an-powered box, two fan-duct systems are connected in parallel: the cold primary airstream dischar ged from the supply f an F1 in the AHU or PU and the recirculating plenum airstream e xtracted by f an F2, as shown in Fig. 21.15. At the combining point A, the pressures of these two airstreams must be equal, and the volume flow rate is the sum of the v olume flow rates of the cold primary air V˙sn,c and the recirculated plenum air V˙sn,r . For a control zone in the perimeter zone using a parallel f an-powered box at its zone peak supply volume f ow rate with con ventional air distrib ution V˙sn,d (a supply air temperature of 55 °F, or 12.8°C), in cfm (L / s), the distribution is as follows: Zone supply volume f ow rate of control zones at summer design conditions Zone supply volume f ow of recirculating plenum air from parallel fan-powered box

1.0V˙sn,d 0

According to Yu et al. (1993), the zone supply v olume f ow rates of control zones in the interior zone using a parallel fan-powered box with cold air distribution at summer design conditions are as follows: Supply volume f ow rate of 44°F (6.7°C) cold primary air, cfm (L / s) Supply volume f ow rate from fan-powered box (recirculating plenum air), cfm (L / s)

0.6V˙sn,d 0.4V˙sn,d

Zone supply volume f ow rate for control zones in the perimeter zone at winter heating design conditions must meet the following requirements: Cold primary air for ventilation Supply temperature differential Ts Tr

0.3V˙sn,d  15°F (8.3°C)

AIR SYSTEMS: VARIABLE-AIR-VOLUME SYSTEMS

FIGURE 21.15 Fan operating characteristics of a parallel f (b) fan-duct system characteristics.

21.51

an-powered VAV system: (a) pressure characteristics;

In a f an-powered VAV system, the fan total pressure of F2 in Fig. 21.15 a is about 0.5 in. WC (125 Pa), and the pressure drop across the f exible duct and the slot diffuser may be around 0.25 in. WC (63 Pa). During summer cooling mode operation, when the zone sensible cooling load is reduced to about one-half of the design load, the volume f ow rate of the cold primary air may be reduced from 0.6V˙sn,d to 0.3 V˙sn,d . As the supply volume f ow through the f exible duct and slot diffuser is reduced accordingly, the pressure drop between points A and o also decreases. The system pressure drop of the recirculating plenum air passage becomes smaller , and a ne w system curv e S2 is formed, as shown in Fig. 21.15 b. The operating point of f an-duct system F2 and S2 mo ves from point P to point Q. The result is a greater v olume f ow rate of the recirculating plenum air V˙sxn,r than that at summer design conditions, and a smaller zone supply volume f ow rate f owing through the f exible duct and slot diffusers.

21.52

CHAPTER TWENTY-ONE

Zone Control and Sequence of Operations of a Fan-Powered VAV System with Parallel Fan-Powered Box Consider a control zone rx1 in the perimeter zone serv ed by a parallel f an-powered box FBX1 that mixes low-temperature cold primary air with warm plenum air, and a control zone ri1 in the interior zone in which the cold primary air is directly supplied from the VAV box, as shown in Fig. 21.13. Because the temperature of cold primary air from the AHU that serves fan-powered box FBX1 and VAV box VI1 Tc,p  44°F (6.7 °C), the summer indoor space relati ve humidity drops to about 40 percent. The summer indoor design temperature is still maintained at 75 °F (23.9°C). At winter design conditions, the zone temperature is maintained at 72 °F (22.2°C) with a relati ve humidity of 20 to 30 percent. 1. When the time-of-day clock signals the AHU to “off,” the AHU and the f an-powered box FBX1 and VAV box VI1, as shown in Fig. 21.13a, are deenergized. 2. When the clock signals the AHU in the on position, if the zone temperature sensed by the temperature sensor T11, Trxn  75°F (23.9°C), then the AHU and fan-powered box FBX1 and VAV box VI1 call for cooling. The DDC system controller then opens the tw o-way valve of the w ater cooling coil and resets the dischar ge air temperature from the AHU Tdis  44°F (6.7°C). The DDC terminal controllers also open the single-blade v olume dampers in the f an-powered boxes and VAV boxes to the maximum cooling position. If the outdoor temperature To  75°F (set point of the temperature economizer c ycle), the DDC system controller also positions the outdoor damper to a minimum opening position and fully opens the recirculating damper. 3. In the perimeter zone, the control zone temperature Trx1 is sensed by sensor T11. The DDC terminal controller modulates the volume damper and thus the volume f ow of cold primary air supply in the f an-powered box FBX1 to maintain a preset zone temperature of Trx1  75°F (23.9°C) with a PI control mode. The fan in the fan-powered box FBX1 is not energized. The control zone temperature in the interior zone Trin is sensed by sensor T21. Another DDC terminal controller modulates the v olume damper in the VAV box VI1 to maintain a preset zone temperature Tri1  75°F (23.9°C). 4. When 52 °F  To  75°F (11.1 °C  To  23.9°C), the DDC system controller fully opens the outdoor air damper and closes the recirculating damper to use 100 percent outdoor air free cooling as the f rst-stage cooling in the AHU. Based on the measured discharge air temperature by sensor T1, the DDC system controller modulates the two-way valve of the water cooling coil as the secondstage cooling simultaneously to maintain a preset discharge air temperature Tdis  44°F (6.7°C). When the outdoor temperature To drops from 65 to 52 °F (18.3 to 11.1 °C), Tdis is reset inversely proportional by the DDC system controller from 44 to 52 °F (6.7 to 11.1 °C). At the same time, the DDC terminal controllers modulate the v olume damper as well as the v olume f ow of cold primary air supply to maintain a preset zone temperature of 75 °F (23.9°C) in the control zones rx1 and ri1. The fan in the fan-powered box FBX1 is not energized. 5. When 52 °F  To  75°F (11.1 °C  To  23.9°C) and the supply v olume f ow rate of the cold primary air from the f an-powered box FBX1 in the perimeter zone has dropped to 0.40 V˙sn,d , the fan in the fan-powered box FBX1 is energized in low-speed operation and extracts recirculating ceiling plenum air to the fan-powered box, and the supply volume f ow rate of recirculating plenum air V˙sn,r to the control zone rx1 is roughly increased to about 0.40V˙sn,d . If the v olume f ow rate of the cold primary air supplied from the f an-powered box FBX1 has dropped to 0.35 V˙sn,d , the fan in the f an-powered box is ener gized in high-speed operation, and its supply volume f ow of recirculating plenum air V˙sn,r is further increased to 0.45V˙sn,d . 6. When 52 °F  To  75°F (11.1 °C  To  23.9°C) and the supply v olume f ow rate of the cold primary air from the f an-powered box FBX1 has dropped to the minimum setting, that is, 0.3V˙sn,d , if the control zone temperature Trx1 sensed by sensor T11 drops below 75°F (23.9°C), that is, 72°F  Trx1  75°F (22.2°C  Trx1  23.9°C), then the fan-powered box FBX1 is in dead-band operation, as shown in Fig. 21.13d. In dead-band operation

AIR SYSTEMS: VARIABLE-AIR-VOLUME SYSTEMS ●

●

●

21.53

The supply volume f ow rate of the cold primary air is at the minimum setting. The fan in the fan-powered box FBX1 is still energized. The heating coil is not energized.

7. When To  75°F (23.9°C) and the supply volume f ow rate of cold primary air from the fan-powered box FBX1 has dropped to the minimum setting, if the control zone temperature Trx1 sensed by sensor T11 drops belo w 72 °F (22.2 °C), the f an-powered box is in heating mode operation. In heating mode operation, the extracted recirculating plenum air is heated at the heating coil, f ows through the centrifugal fan, mixes with the cold primary air , and is then supplied to the control zone, as shown in Fig. 21.13b. The DDC terminal controller opens and modulates the mass f ow rate of hot water entering the heating coil to maintain a control zone temperature T  72°F (22.2°C) in reverse-acting mode. The operation of the fan in a DDC fan-powered box can be arranged in different alternatives.

Supply Volume Flow Rate According to Eq. (20.69), at summer design conditions, the volume f ow rate of cold primary air supplied to a control zone with conventional air distribution V˙sn,d in cfm [m 3 / (60 s)], can be calculated as V˙sn,d 

Q rsn 60 scpa(Tr Ts)

where Qrsn  zone peak sensible cooling load, Btu / h (W) and Tr Ts  20F (11.1C) At summer design conditions, the volume f ow rate of cold primary air supplied to a control zone with cold air distribution V˙sn,c dis in cfm [m3 / (60 s)], can be calculated as V˙sn,c dis  KV˙sn,d 

Q rsn 60 scpa(Tr Ts,c dis)

where Qrsn  zone peak sensible cooling load, Btu / h (W) K  a factor Tr Ts,c dis  32°F (17.8°C) At winter design conditions, the volume f ow rate of the mixture of cold primary air and recirculating plenum air V˙sn,h, in cfm [m3 / (60 s)], supplied to a control zone in the perimeter zone can be calculated as V˙sn,h 

Q rnh 60 scpa(Ts,h Tr)

(21.18)

where Qrnh  zone heating load, Btu / h (W) Ts,h  supply temperature at winter design conditions, °F (°C) Warm air supply temperature dif ferential Ts,h Tr should not exceed 15°F (8.3°C) in order to prevent buoyancy effects. Supply volume f ow rate of a control zone in the perimeter zone serv ed by a f an-powered VAV system with cold air distrib ution is sho wn in Fig. 21.16. In Fig. 21.16: V˙sn,c, V˙sn,r indicate supply volume f ow rate of cold primary air and recirculated plenum air to the perimeter control zone, in cfm (L / s), and Ts,n the supply air temperature to the perimeter control zone xn, in °F (°C). As in the VAV reheat system, the cooling and heating coil load can be similarly calculated by Eqs. (21.8), (21.9), and (21.10).

CHAPTER TWENTY-ONE

Percentage of zone peak (design) supply flow

80 Ts n

60 40

Supply temp Tsxn, F

21.54

. . Vsn, c Vsn, r

80 60

. Vsn, r . Vsn, c

40 20 0

Heating

Cooling Dead band

FIGURE 21.16

Supply volume f ow rate of a control zone in the perimeter zone served by a fan-powered VAV system.

Fan Energy Use In a f an-powered VAV system, there are tw o kinds of f an energy use: the supply f an energy use in an AHU or a PU and the ener gy use of the small centrifugal f ans in the fan-powered boxes. Typical supply f an ef f ciency is about 60 percent, and the motor ef f ciency is around 90 percent, which gives a combined fan and motor eff ciency of about 55 percent. According to Elleson (1993), the eff ciency of fans in fan-powered boxes from four manufacturers ranged from 16 to 35 percent, and the motor eff ciency ranged from 45 to 55 percent. Assuming a fan eff ciency of 30 percent and a motor eff ciency of 50 percent, the combined fan and motor eff ciency for the fan in the fan-powered box is 15 percent. Elleson (1993) has collected detailed data for the 1988 cooling season from the fourth f oor of a 12-story b uilding in southern California using a f an-powered VAV system (parallel f an-powered boxes). The results are as follows: 55°F (12.8°C) Supply air

Continuous operation, kWh Intermittent operation, kWh No fan-powered boxes, kWh

45°F (7.2°C) Supply air

Supply fan

Fan-powered boxes

Total

Supply fan

Fan-powered boxes

Total

11,720 11,720 11,720

7,000 1,861 0

18,720 13,581 11,720

7,794 7,794 8,417

7,000 2,140 0

14,794 9,934 8,417

Intermittent operation indicates that fans in parallel fan-powered boxes operated only when the cold primary airf ow fell below 50 percent of design f ow.

AIR SYSTEMS: VARIABLE-AIR-VOLUME SYSTEMS

21.55

If series fan-powered boxes and parallel f an-powered boxes are both operated continuously , the energy use of the supply fan plus the fans in the parallel fan-powered boxes consume less than onehalf of the kWh used by the supply f an plus the f ans in the series f an-powered box es in 45 °F (7.2°C) cold air distribution.

Design Considerations ●

●

●

●

●

●

●

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●

Because there are two mixings (the mixing of outdoor air and recirculating air in the AHU or PU and the mixing of cold primary air and plenum air in the f an-powered box), adequate supply of outdoor air is critical in a fan-powered VAV system. This is discussed in Chap. 23. In a VAV system using cold air distrib ution, fan-powered box es can be installed in both the perimeter and interior zones, or in the perimeter zone only, for lower initial cost. Cold primary air is supplied directly to the interior zone by high-induction ceiling dif fusers, as sho wn in Fig. 18.10a and b. The mixing ratio of cold primary air to zone supply air and the reset of the cold primary air supply temperature for the perimeter zone are dif ferent from those of the interior zone. Dif ferent AHUs or PUs are preferable for the perimeter zone and interior zone separately . An analysis of energy savings and investment costs is often necessary. In a series fan-powered box, the supply fan in the AHU or PU offsets the system pressure drop up to the inlet of the VAV box. The fan in the f an-powered box offsets the pressure drop of the VAV box, elbows (if any), downstream heating coil, f exible duct, and diffuser. In a parallel f an-powered box, the supply f an in the AHU or PU of fsets the pressure drop of the VAV box, f exible duct, diffuser, and straight-through mixing loss of the cold primary airstream in the fan-powered box. At summer and winter design conditions, the ratio of the v olume f ow rate of cold primary air supplied to the f an-powered box to the zone supply v olume f ow rate V˙sn,c /V˙sn,d and the ratio of the v olume f ow rate of w arm recirculating plenum air to the zone supply v olume f ow r ate V˙s,rec /V˙sn,d depend on the follo wing f actors: (1) zone sensible cooling load and heating load, (2) zone supply temperature dif ferential at summer and winter design conditions Trn Tsn,d, or Tsn,d Trn, and (3) outdoor air requirements. As mentioned pre viously, for cold air distrib ution V˙sn,c /V˙sn,d is about 0.6 for a f an-powered box. The winter heating in the perimeter zone often needs greater supply v olume f ow to match the requirement of the w arm air supply temperature dif ferential not e xceeding 15 °F (8.3 °C); V˙sn,h /V˙sn,d for cold air distribution is about 0.7. The supply main ducts, branch ducts, cold air passage in the VAV box, f exible ducts, and diffusers must be well insulated especially for cold air distribution. Cold primary air and recirculated plenum air should be mix ed in a perpendicular f ow direction to prevent downstream temperature stratif cation. A low zone relati ve humidity and zone de w point are desirable to pre vent possible surf ace condensation during cold air distrib ution. When cold primary air at 44 °F (6.7°C) is supplied directly to the conditioned space through high induction dif fusers, the surface of the induction dif fusers can be assumed to be 3 °F (1.7°C) higher than the supply air; or the de w point of zone air should be lower than 44 3  47°F (8.3°C), that is, a zone temperature of 75 °F (23.9°C) and a space relative humidity r  37 percent. At the start of cool-do wn period in a hot and humid summer , direct supply of 40 to 50 °F (4.4 to 10.0°C) cold primary air to the f an-powered boxes may cause condensation because of the higher dew point of stagnant air in the ceiling plenum before the cool-down period. The AHU or PU should supply 55°F (12.8°C) air to the fan-powered boxes f rst, to reduce moisture and lower the dew point. If humidif cation is required to maintain an acceptable zone relative humidity in winter, a humidif er can be installed in the AHU or PU.

21.56

CHAPTER TWENTY-ONE ●

●

During the unoccupied period in winter , the AHU or PU can be shut of f while the f an-powered boxes provide the necessary heating. Fan-powered boxes installed inside the ceiling plenum must meet local safety code requirements.

21.6 COMPARISON BETWEEN VARIOUS VAV SYSTEMS VAV systems are used in applications where space load is v aried in order to maintain a required indoor environment and to save fan energy. Single-zone VAV systems are mainly used for single-zone conditioned spaces such as arenas, indoor stadiums, assembly halls, and many factories. VAV reheat systems using reheating VAV boxes are simple and effective for perimeter heating in winter and with VAV boxes provide cooling only in the interior zone in a b uilding. Simultaneous cooling and heating may occur in reheat systems at lo w zone cooling loads. The VAV reheat system becomes energy-ineff cient for locations where the winter heating load linear density e xceeds 120 Btu / h ft (115 W / m). Perimeter heating VAV systems are used for b uildings that need heating for their perimeter zones in locations with cold and long winters. Dual-fan dual-duct VAV systems pro vide satisf actory indoor en vironmental control for multizone high-rise buildings and health care facilities with f exible operating characteristics. If the shutoff air leakage in the mixing VAV box can be minimized, it is often an appropriate choice for more demanding applications. Fan-powered VAV systems pro vide low cooling load compensation and winter heating from a two-stage sequence operation: intake w arm plenum air and winter heating. The small centrifugal fan in the f an-powered box also pro vides the mixing of cold primary air with the plenum air . Fanpowered VAV systems are widely used in b uildings where cold air distrib ution is needed, such as off ce buildings and schools.

REFERENCES Alley, R. L., Selecting and Sizing Outside and Return Air Dampers for VAV Economizer Systems, ASHRAE Transactions, 1988, Part I, pp. 1457 – 1466. ANSI / ASHRAE Standard 62-1999, Ventilation for Acceptable Indoor Air Quality, ASHRAE Inc., Atlanta, GA, 1999. ASHRAE, ASHRAE Handbook 1996, HVAC Systems and Equipment, ASHRAE Inc., Atlanta, GA, 1996. ASHRAE, ASHRAE Handbook 1999, HVAC Applications, Atlanta, GA, 1999. ASHRAE / IESNA Standard 90.1-1999, Energy Standard for Buildings Except New Low-Rise Residential Buildings, ASHRAE Inc., Atlanta, GA, 1999. Avery, G., The Myth of Pressure-Independent VAV Terminals, ASHRAE Journal, no. 8, 1989, pp. 28 – 30. Avery, G., The Instability of VAV Systems, Heating / Piping / Air Conditioning, February 1992, pp. 47 – 50. Avery, G., Selecting and Controlling Economizer Dampers, HPAC, no. 8, 1996, pp. 73 – 78. Brothers, P. W., and Warren, M. L., Fan Energy Use in Variable Air Volume Systems, ASHRAE Transactions, 1986, Part II B, pp. 19 – 29. Cappellin, T. E., VAV Systems — What Makes Them Succeed? What Makes Them Fail? ASHRAE Transactions, 1997, Part II, pp. 814 – 822. Chen, S., and Demster, S., Variable Air Volume Systems for Environmental Quality, McGraw-Hill, New York, 1996. Coggan, D., Mixed Air Control with DDC, Heating / Piping / Air Conditioning, no. 5, 1986, pp. 113 – 115. Cox, R. L., Retrof t for a College Campus, 1996 ASHRAE Technology Award Case Study, ASHRAE Journal, no. 11 1996, pp. 59 – 64.

AIR SYSTEMS: VARIABLE-AIR-VOLUME SYSTEMS

21.57

Dean, R. H., and Ratzenberger, J., Stability of VAV Terminal Unit Controls, Heating / Piping / Air Conditioning, October 1985, pp. 79 – 90. Dorgan, C. E., and Elleson, J. S., Cold Air Distribution, ASHRAE Transactions, 1988, Part I, pp. 2008 – 2025. Elleson, J. S., Energy Use of Fan-Powered Mixing Boxes with Cold Air Distribution, ASHRAE Transactions, 1993, Part I, pp. 1349 – 1358. Ellis, R., and McKew, H., Back to Basics: VAV Terminal Building Automation — Design and Commissioning, Engineered Systems, no. 3, 1997, pp. 9– 10. Englander, S. L., and Norford, L. K., Saving Fan Energy in VAV Systems — Part 1: Analysis of a VariableSpeed-Drive Retrof t, ASHRAE Transactions, 1992, Part I, pp. 3– 18. Grimm, N. R., and Rosaler, R. C., HVAC Systems and Components Handbook, 2d ed., McGraw-Hill, New York, 1998. Haines, R. W., Supply Fan Volume Control in a VAV System, Heating / Piping / Air Conditioning, August 1983, pp. 107 – 111. Int-Hout, III, D., Analysis of Three Perimeter Heating Systems by Air Diffusion Methods, ASHRAE Transactions, 1983, Part I B, pp. 101 – 112. Int-Hout, III, D, Stand-Alone Microprocessor Control of Dual-Duct Terminals, ASHRAE Transactions, 1987, Part II, pp. 1722 – 1733. Kettler, J. P., Eff cient Design and Control of Dual-Duct Variable-Volume Systems, ASHRAE Transactions, 1987, Part II, pp. 1734 – 1741. Krajnovich, L., and Hittle, D. C., Measured Performance of Variable Air Volume Boxes, ASHRAE Transactions, 1986, Part II A, pp. 203 – 214. Linder, R., and Dorgan, C. B., VAV Systems Work Despite Some Design and Application Problems, ASHRAE Transactions, 1997, Part II, pp. 807 – 813. Linford, R. G., Dual-Duct Variable Air Volume — Design / Build Viewpoint, ASHRAE Transactions, 1987, Part II, pp. 1742 – 1748. Lizardos, E. J., Economizer Cycle and Damper Sizing, Engineered Systems, no. 1, 1997, pp. 126 – 131. Lynn, M., Balancing DDC — Controlled Boxes, Heating / Piping / Air Conditioning, July 1989, pp. 79 – 84. Mutammara, A. W., and Hittle, D. C., Energy Effects of Various Control Strategies for Variable-Air-Volume Systems, ASHRAE Transactions, 1990, Part I, pp. 98 – 102. Muxen, S. A., and Chapman, W. F., Versatile Application-Specif c Controllers for Hotel Guest Rooms, ASHRAE Transactions, 1988, Part I, pp. 1530 – 1538. Schuler, M., Dual-Fan, Dual-Duct System Meets Air Quality, Energy-Eff ciency Needs, 1996 ASHRAE Technology Awards Case Study, ASHRAE Journal, no. 3, 1996, pp. 39 – 41. Shavit, G., Retrof t of Double-Duct Fan System to a VAV System, ASHRAE Transactions, 1989, Part I, pp. 635 – 641. Smith, M. C., Industry Standard 885, Acoustic Level Estimation Procedure Compared to Actual Acoustic Levels in an Air Distribution Mock Up, ASHRAE Transactions, 1989, Part I, pp. 543 – 548. Spitler, J. D., Pedersen, C. O., Hittle, D. C., and Johnson, D. L., Fan Electricity Consumption for Variable Air Volume, ASHRAE Transactions, 1986, Part II B, pp. 5– 18. Spitler, J. D., Hittle, D. C., Johnson, D. L., and Pederson, C. O., A Comparative Study of the Performance of Temperature-Based and Enthalpy-Based Economy Cycles, ASHRAE Transactions, 1987, Part II, pp. 13 – 22. Straub, H. E., and Cooper, J. G., Space Heating with Ceiling Diffusers, Heating / Piping / Air Conditioning, May 1991, pp. 49 – 55. Tackett, R. K., Case Study: Off ce Building Use Ice Storage, Heat Recovery, and Cold Air Distribution, ASHRAE Transactions, 1989, Part I, pp. 1113 – 1121. Taylor, S. T., Series Fan-Powered Boxes: Their Impact on Indoor Air Quality and Comfort, ASHRAE Journal, no. 7, 1996, pp. 44 – 50. The Trane Company, Variable Air Volume Systems Manual, American Standard Inc., La Crosse, WI, 1988. Wacker, P. C., Economizer Saving Study, ASHRAE Transactions, 1989, Part I, pp. 47 – 51. Wang, S. K., Air Conditioning, vol. 4, Hong Kong Polytechnic, Hong Kong, 1987. Warden, D., Dual Fan, Dual Duct Systems, Better Performance at a Low Cost, ASHRAE Journal, no. 1, 1996, pp. 36 – 41.

21.58

CHAPTER TWENTY-ONE

Wendes, H. C., Estimating VAV Retrof t Costs, Heating / Piping / Air Conditioning, August 1983, pp. 93 – 103. Wheeler, A. E., Energy Conservation and Acceptable Indoor Quality in the Classroom, ASHRAE Journal, April 1992, pp. 26 – 32. Williams, V. A., VAV System Interactive Controls, ASHRAE Transactions, 1988, Part I, pp. 1493 – 1499. Yu, H. C., Harmon, J. J., and Galway, J. M., Renovation of the Watergate 600 Building, ASHRAE Journal, no. 10, 1993, pp. 30 – 36.

CHAPTER 22

AIR SYSTEMS: VAV SYSTEMS — FAN COMBINATION, SYSTEM PRESSURE, AND SMOKE CONTROL 22.1 RETURN AND EXHAUST SYSTEMS 22.2 Types of Return and Exhaust Systems 22.2 Return Ceiling Plenum 22.2 Low-Level Return Systems and Enclosed Parking Garage Ventilation 22.2 Exhaust Hoods 22.3 22.2 FAN COMBINATIONS AND OPERATING MODES 22.4 Fan Combinations 22.4 Operating Modes 22.4 22.3 SYSTEM PRESSURE DIAGRAM 22.5 System Pressure and Duct Static Pressure Control 22.5 Fan Characteristics 22.7 Mixing-Exhaust Section and Conditioned Space 22.8 22.4 SUPPLY FAN AND EXHAUST FAN COMBINATION 22.8 System Characteristics 22.8 Operating Characteristics 22.9 Recirculation Mode and Design Volume Flow 22.9 Recirculation Mode, 50 Percent Design Flow Rate 22.12 Air Economizer Cycle 22.13 Warm-Up and Cool-Down Mode 22.13 Pressure Variation at the Mixing Box 22.13 22.5 SUPPLY FAN AND RELIEF FAN COMBINATION 22.14 Recirculation Mode 22.14 Air Economizer Mode and Design Volume Flow Rate 22.14 Air Economizer Mode, 50 Percent Design Flow 22.17 Warm-Up and Cool-Down Mode 22.17 Design Considerations and Controls 22.17 22.6 SUPPLY FAN AND RETURN FAN COMBINATION 22.18 Recirculation Mode 22.18

Air Economizer Mode 22.20 Controls 22.21 22.7 COMPARISON OF THREE FAN COMBINATION SYSTEMS 22.21 22.8 PRESSURE FLOW CHARACTERISTICS 22.22 Pressure Flow Characteristics for a Supply and Return Fan Combination System 22.22 Variation of Pressure in the Mixing Box 22.23 Field Survey of System Pressure Characteristics of a VAV System Using a Supply and Relief Fan Combination 22.23 22.9 SMOKE CONTROL AND FIRE SAFETY 22.24 Fire Safety in Buildings 22.24 Smoke Movements in Buildings 22.25 Effective Area and Flow Rate 22.27 22.10 EFFECT OF AUTOMATIC SPRINKLER ON FIRE PROTECTION 22.27 22.11 SMOKE CONTROL IN ATRIA 22.28 ANSI / NFPA 92A and 92B 22.28 Smoke Management in Atria, Malls, and Large Areas 22.28 22.12 STAIRWELL PRESSURIZATION AND ZONE SMOKE CONTROL 22.29 Stairwell Pressurization 22.29 Characteristics of Stairwell Pressurization 22.29 Overpressure Relief and Feedback Control 22.30 Stair and Shaft Vents 22.31 Zone Smoke Control 22.31 Design Considerations 22.32 Volume Flow Rate 22.32 System Pressure Loss for Stairwell Pressurization System 22.33 REFERENCES 22.38

22.1

22.2

CHAPTER TWENTY-TWO

22.1 RETURN AND EXHAUST SYSTEMS Types of Return and Exhaust Systems As described in Sec. 20.3, for any enclosed conditioned space in a b uilding, if the difference of air densities is ignored, the total v olume fl w rate of air entering the space must be equal to the total volume fl w rate of air leaving the space. Among the three main categories of air systems described in Sec. 20.9 (Fig. 20.16): ●

●

For most of the constant-v olume systems, return air is often returned to the packaged unit through door undercuts and a collecti ve return grille in the closet, under the attic, or at the high level in the basement. An e xhaust f an is optional and is usually installed in the bathroom or toilet. In many dedicated outdoor air and space recirculating systems, an exhaust system with a v olume fl w rate less than the supply v olume fl w rate of the outdoor v entilation air is often emplo yed in restrooms and service rooms. This is discussed in Chap. 28.

Return Ceiling Plenum Among the v ariable-air-volume and dedicated minimum outdoor f an systems, use of the ceiling plenum as a return air plenum, as sho wn in Fig. 20.16 b and c, is the most common return air system. In a return ceiling air plenum system, return air is often returned to the ceiling plenum through the return slots or grilles first. Usually a connected return duct connects the ceiling plenu and the fan room. Part of the return air becomes recirculating air and enters the AHU or PU, and the remaining air is exhausted outside. For a return ceiling plenum, the following must be considered during design: ●

●

●

Fire protection. Because the return air fl ws through the ceiling plenum, refer to federal and local fire codes for the requirements materials, and equipment installed inside the ceiling plenum, as well as the fire all and partitions that separate the ceiling plenum and adjacent fire compart ments. Noise contr ol. F an noise that has been transmitted into the ceiling plenum is v ery dif ficult t attenuate. Do not install silencers inside the ceiling plenum. Install silencers in the f an room. Check both duct-borne noise and radiated noise; also check noise transmitted from both supply and return air sides. Air leaka ge. Because most of the ceiling plenums use acoustical ceiling tile and T-bar construction, the airtightness between the conditioned space and the ceiling plenum is hard to maintain. Field measurement sho wed that the pressure dif ferential between the conditioned space and the ceiling plenum in a VAV system serving a small public library pr  pl  0.004 in. WC (1 Pa). The critical air leakage from a ceiling plenum is the air leakage to outdoors, or vice versa.

Low-Level Return Systems and Enclosed Parking Garage Ventilation Low-level return systems have been widely used in the following applications: ●

In indoor stadiums and assembly halls, where there are many spectator seats. It is better to extract the return air from the return grilles located at lo w-level seats to induce part of the dif fused

AIR SYSTEMS: FAN COMBINATION AND SMOKE CONTROL

●

22.3

airstreams fl wing from the upper le vel to the lo w level, as shown in Fig. 18.20. The return air then fl ws through the return air duct, pulled by a return fan to the fan room, and is either recirculated or exhausted to outdoors. In clean rooms or clean spaces, in order to form a unidirectional airfl w from the ceiling to the floo , so that dust will not ascend from the low level to contaminate the working area, as shown in Figs. 18.23 and 18.24.

Fully enclosed under ground parking garages require mechanical v entilation systems including e xhaust systems, makeup air supply systems, or both to maintain a CO concentration not e xceeding 35 ppm for up to 1 hour e xposure. ASHRAE sponsored a study that found a ventilation requirement of 1.5 cfm / ft2 (7.5 L / s  m2) or 6 air changes per hour (ach) is satisfactory. ASHRAE / IESNA Standard 90.1– 1999 mandatorily specifies that garage entilation air systems with a total volume fl w rate greater than 30,000 cfm (14,200 L / s) shall be permited to ( a) have at least an automatic control to stage fans or modulating fan volume to maintain CO concentration below levels specified in ASHRAE Standard 62, 35 ppm up to 1 hour e xposure, or (b) be capable of shutting off fans or reducing fan volume when the garage is not in use.

Exhaust Hoods ASHRAE / IESNA Standard 90.1– 1999 specifies that ind vidual kitchen e xhaust hoods lar ger than 5000 cfm (2360 L / s) shall be pro vided with a mak eup air system with a capacity equal to at least 50 percent of e xhaust v olume fl w rate that is ( a) unheated or heated to no higher than 60°F (10.6°C) and ( b) uncooled or cooled without the use of mechanical cooling (refrigeration). Exceptions include ●

●

Exhaust hoods used to exhaust ventilation air which otherwise would exfiltrate or be xhausted by other air systems Certified grease xtractor hoods that require a face velocity no greater than 60 fpm (0.3 m / s)

Standard 90.1– 1999 also specifies that uildings with fume hood exhaust systems having a total exhaust volume fl w rate greater than 15,000 cfm (7080 L / s) shall include at least one of the following features: ●

●

●

Variable-air-volume hood e xhaust and room supply systems capable of reducing e xhaust and makeup volume fl w to 50 percent or less of design fl ws. Direct mak eup air supply equal to no less than 75 percent of the e xhaust rate, heated to no warmer than 2°F (1.1°C) belo w room set point, cooled to no cooler than 3°F (1.7°C) abo ve room set point, no humidification all wed, and no simultaneous heating and cooling for dehumidification Heat recovery system for mak eup air from fume hood e xhaust with same e xceptions as in Sec. 12.5. Refer to Standard for details.

Standard 90.1– 1999 also mandatorily specifies that all outdoor air supply and xhaust hoods, vents, and ventilators shall be installed with motorized dampers that will automatically shut when the spaces serv ed are not in use. Gra vity (nonmotorized) dampers are acceptable in b uildings less than three stories in height above grade and for buildings of any height located in climates with less than 2700 HDD65. Exceptions are for v entilation systems serving unconditioned spaces as well as during preoccupancy building warmup, colddown, and setback, except when ventilation reduces energy costs (such as, night pur ge). Both outdoor air supply and e xhaust air dampers shall ha ve a maxium leakage rate of 3 cfm/ft2 at 1.0 in. WC (15 L / s  m2 at 250 Pa).

22.4

CHAPTER TWENTY-TWO

22.2 FAN COMBINATIONS AND OPERATING MODES Fan Combinations In commercial b uildings, constant-volume systems with optional e xhaust are widely used in small air systems, and the VAV systems with return ceiling plenum are the most widely used in mediumsize and lar ge air systems. As discussed in Sec. 20.9, among these air systems, the following fan combinations are often used: ●

●

●

Supply and exhaust fan combination Supply and relief fan combination Supply and return (including unhoused plug / plenum) fan combination

Supply, exhaust, relief, and return f ans can either be a single f an or multiple f ans. Recently, some manufacturers use unhoused plug / plenum f an as the return f an to form a supply and return f an combination. As discussed in Sec. 15.2 an unhoused plug / plenum f an is simpler in construction and has the fl xibility of location of the dischar ge outlet. Ho wever, it requires more po wer than a centrifugal fan of the same capacity and fan total pressure. Air systems of various fan combinations are installed to meet the following requirements: ●

●

●

●

●

●

To pro vide required outdoor v entilation air for occupants and conditioned space according to ASHRAE Standard 62– 1999 To supply conditioned air to the space to offset space load To recirculate space air for energy conservation To operate in an air economizer cycle for energy saving To exhaust objectionable contaminated air from the space To maintain a proper indoor environment

Operating Modes Constant-volume systems with optional e xhaust and VAV systems with return ceiling plenum may be operated in the following modes: ●

●

●

Cooling mode is one of the basic operating modes in air systems. In cooling mode operation, outdoor air or recirculating air, or a mixture of the two is cooled or is cooled and dehumidified in th water cooling coil or DX coil, and the conditioned air is supplied to the conditioned space to of fset the cooling load. Cooling modes can again be subdi vided into summer cooling mode and winter cooling mode. Man y interior zones that ha ve heavy internal cooling loads still need cold air supply in the winter cooling mode. Heating mode is another basic operating mode in air systems. In heating mode operation, outdoor air, recirculating air, or their mixture is heated in the direct-fired furnace electric heater or w ater heating coil. Warm air is supplied to the conditioned space to of fset the heating load. In heating mode, humidification is optional In part-load operation, the space cooling or heating load is less than the summer or winter design load. This may be due to the variation of the space load, or change of the outdoor climate, or both. Air systems operate at part load most of the time. During part-load operation, the supply volume fl w rate as well as the cooling and heating capacities of the air system reduces accordingly . It is assumed that the supply v olume fl w rate may be reduced to 50 percent, and even dropped to 40 percent of design fl ws.

AIR SYSTEMS: FAN COMBINATION AND SMOKE CONTROL ●

●

●

●

22.5

In minimum outdoor air recirculating mode or simply recirculating mode, only minimum outdoor air contained in the mixture of outdoor and recirculating air is supplied to the conditioned space for the occupant and for dilution of contaminated air . Minimum outdoor air circulation mode can be a subsidiary operating mode in a cooling or heating mode operation. In an air economizer mode, only the outdoor air or a mixture of outdoor and recirculating air is supplied to the conditioned space to replace part of or all the mechanical cooling (cooling from refrigeration). Air economizer mode is an outdoor air free-cooling mode and is often used as the first stage of cooling to s ve energy. In purge mode, 100 percent outdoor air is e xtracted during the nighttime or in the morn-ing before the space is occupied, to cool do wn the space with outdoor air or to flush the conditione space with cleaner outdoor air while at the same time squeezing out the indoor contaminated air. In warm-up and cool-down mode, the space air temperature is w armed up or cooled down to predetermined limits after the air system is started in the morning while the space is not occupied. During warm-up or cool-down mode, the recirculating damper is fully open and the outdoor and exhaust dampers are closed, except when the w arm-up or cool-down mode is combined with the purge mode. Warm-up and cool-down mode operation should not take place prior to the purge mode.

●

●

In nighttime setback mode, the space temperature is set back when the conditioned space is unoccupied in winter, typically from 70 to 55°F (21.1 to 12.8°C), to prevent freezing indoors and to reduce the warm-up temperature differential the next morning before the space is occupied. Occupied mode is for the operating hours when the conditioned space is occupied by the occupants, and the unoccupied mode is for the operating hours when the conditioned space is not occupied by the occupants.

22.3 SYSTEM PRESSURE DIAGRAM The system pressure diagram of a VAV system during minimum outdoor air recirculating mode is shown in Fig. 22.1. The system characteristics of a VAV system not only af fect its f an energy use, but also are closely related to the v ariation in its supply volume fl w rate and operating characteristics. System pressure is usually e xpressed in total pressure, in in. WG (Pa), in the form of a system pressure diagram. In conditioned space and in places where v elocity pressure is so small, it is often ignored. The total pressure is then equal to static pressure, that is, pt  ps.

System Pressure and Duct Static Pressure Control In a VAV system with duct static pressure control, the total pressure loss of the VAV system can be di vided into tw o parts: variable part pvar and fixed part pfix, both in in. WC (Pa). The total pressure loss between the centerline of the recirculating damper in the PU and the static pressure sensor of the duct static pressure control is e xpressed as psf,var, in in. WC (P a). As the supply v olume flo w rate in a VAV system v aries, pvar varies accordingly. The f ixed part pfix is independent of the v ariation in the supply flo w rate and remains constant as the v olume damper in the VAV box increases its pressure loss at reduced v olume flow rate (Figs. 20.13 and 22.1). For a VAV system installed with a supply f an and a return f an connected in series, as shown in Figs. 22.1 and 22.4, its system pressure loss psys, in in. WC (Pa), is also divided into variable part

CHAPTER TWENTY-TWO

22.6

Mixing exhaust section

Recirculating damper Exhaust damper

ru O

Return fan

2.0 pt, in WG 1.0 pro

po

0 1.0 2.0

m O

Return duct rt

rec

Minimum outdoor air damper Mixed plenum Supply fan

AFD

sf

cc

VAV box

Economizer air damper

Recirculating/ exhaust chamber

P1

sfo

rfo pret

Return air

rfi

Outdoor air

d

Supply duct

AFD

pr,s Exhaust air

Space

s

psf,var Supply air

psf,fix pt

P2

pom

psf,fix

pr

pro  0.03 in WG po

psf

sfi

3.0 4.0 prf

psf

FIGURE 22.1 System pressure diagram for a supply-return f an combination air system (connected in series) (r f  return fan inlet; rfo  return fan outlet; sf  supply fan inlet; sfo  supply fan outlet; prf  return fan total pressure; psf  supply fan total pressure; AFD  adjustable-frequency, variable-speed drive).

pvar and fi ed part pfi , in in. WC (Pa), and it can be calculated as psys  pvar  pfix  R varV˙ 2var  pfix  (pr, s  pPU  psd,var)  pfix

(22.1)

 (pr, s  psf,var)  pfix where pr,s  pressure loss of return air system, in. WC (Pa) pPU  pressure drop across PU, in. WC (Pa) psd, var  pressure loss of supply duct system before duct static pressure sensor, in. WC (Pa) psf, var  pressure loss of variable part of supply fan, in. WC (Pa) Rvar  fl w resistance representing variable part of VAV system, in. WC / (cfm)2 (Pa  s2 / m6) V˙var  supply volume fl w rate of VAV system, cfm (m3 / s) In a VAV system with a supply and return f an combination connected in series, the supply volume fl w rate fl wing through the supply f an is dif ferent from the return v olume fl w rate fl wing through the return air f an. This makes the pressure-volume operating characterirstics more complicated. If the pressure sensor is installed closer to the supply f an, the fan power savings at reduced volume fl w decreases because of the increase in pfi . The set point of the static pressure sensor may vary from 0.7 to 1.5 in. WC (125 to 375 Pa). In a VAV system serving a single floor in a multistory uilding, if the maximum design air v elocity in the main duct vmax is 3000 fpm (15 m / s), and if the length of main duct is 150 ft (45.7 m),

AIR SYSTEMS: FAN COMBINATION AND SMOKE CONTROL

22.7

then the design total pressure loss of this supply duct system psd, var may vary from 0.75 to 1.5 in. WC (187 to 375 P a), typically 1 in. WC (250 P a), depending on the duct f ttings in the main duct.

Fan Characteristics A VAV system with a supply and return f an or supply and relief fan combination is often connected in series in a f an-duct system. The reasons to use a supply f an and a return f an instead of a single supply fan are as follows: ●

●

●

To meet the requirement of maintaining a slightly positi ve pressure in the conditioned space as well as at the inlet of the AHU or PU point ru, and a negative pressure in the mixing box (mix ed plenum) to extract outdoor air To exhaust 100 percent free-cooled outdoor air during an air economizer cycle To set up dif ferent volume f ow rates of supply air and return air in order to pro vide a required amount of outdoor ventilation air as well as to maintain a desirable space pressure

During the design of a series-connected supply and return f an combination system, as shown in Fig. 22.1, it is essential that the following parameters be carefully calculated and analyzed: Clearly di vide the total pressure loss that should be undertak en by the supply f an, or by the return f an during minimum outdoor air recirculating mode at summer design conditions. The return system pressure loss undertak en by the return f an start from the inlet of return slot rt includes the return slot, airf ow inside the ceiling plenum which is often ignored because of its lower air v elocity, the return grille if an y, the return duct and duct f ttings, up to part of the recirculating damper, as shown in Figs. 22.1 and 22.4. F or the PU with a con f guration shown in Fig. 22.1, in minimum outdoor air recirculating mode, a 90° turn plus one-half of the total pressure loss of the recirculating damper at full opening design f ow is always greater than a 90° turn plus the total pressure loss of the exhaust damper and louvers at minimum outdoor air recirculating mode. In a VAV system, the f an total pressure of the return f an pt,rf must be carefully calculated, neither overestimated nor underestimated. ●

●

Based on the principle of conserv ation of mass, from Eq. (20.8), the relationship between the mass f ow rates of supply, return, and exhaust f ow rate can be expressed as m˙s  m˙rt  m˙ex  m˙exf

(22.2)

where ,m˙s m˙rt, m˙ex  mass f ow rate of supply air, return air, and exhaust air, lb / min (kg / min) m˙exf  mass f ow rate of exf ltrated air, lb / min (kg / s) If the difference in air densities is ignored and a positi ve pressure is maintained in the conditioned space, from Eq. (20.9), the relationship between the supply , return, and exhaust volume f ow rates can be expressed as V˙s  V˙rt  V˙ex  V˙exf

(22.3)

where V˙s, V˙rt, V˙ex  volume f ow rates of supply air, return air, and exhaust air, cfm (L / s) V˙exf  volume f ow rate of exf ltrated air, cfm (L / s) If the supply air is at 55 °F (12.8°C) dry-bulb and 54 °F (12.2°C) wet-bulb temperature and the return air is at 78 °F (25.6°C) dry-bulb and 65°F (18.3°C) wet-bulb temperature, when the mass f ow rates of the supply and return air are the same, the volume f ow rate of return air is about 4 percent greater than that of the supply air . During cold air distrib ution, if the supply air is at 44 °F (6.7°C) dry-bulb and 43 °F (6.1°C) wet-bulb temperature, and the return air remains the same, the volume

22.8

CHAPTER TWENTY-TWO

f ow rate of the return air is 7 percent greater taken into consideration.

. F or accurate analysis, these differences should be

Mixing-Exhaust Section and Conditioned Space In an air system, there are two places where the air is connected to the outside atmosphere: (1) the mixing box – exhaust chamber (simply the mixing-e xhaust section), where it is connected with the outside atmosphere through the outdoor air intak e and the exhaust passage, and (2) the conditioned space, where it connects with the outside atmosphere through the openings in the b uilding shell, as shown in Fig. 22.1. In a mixing box of a VAV system, the variation of the ne gative pressure af fects the amount of outdoor air intake and, therefore, the indoor air quality of the conditioned space. Within the conditioned space, the v ariation of the supply air , return air , exhaust air , inf ltration, and e xf ltration makes the space airf ow balance and associated space pressure complicated.

22.4 SUPPLY FAN AND EXHAUST FAN COMBINATION System Characteristics Air systems equipped with a single supply f an and an e xhaust f an are often used in b uildings in which the total pressure loss of the return duct system is low. A certain volume f ow of air is usually exhausted from rooms such as restrooms. A barometric relief damper is often installed in the conditioned space or in the return plenum to avoid excessively high space pressure. When the space positive pressure on the damper is greater than the weight of the damper , the damper opens and the space pressure is relie ved and is thus maintained belo w a predetermined v alue. Such an air system may be equipped with a water economizer. To analyze the pressure characteristics of the airstream at tw o cross-sectional planes along the airf ow in an air system, the total pressure is determined from Eq. (17.12): pt1  pt2  pf . To determine whether air is f owing from one enclosed space to another , or from an enclosed space to the outdoors, the static pressure dif ference between these tw o places ps  ps1  ps2 should be calculated. Ho wever, the static pressure inside an enclosure of uniform pressure, or the mean static pressure on a cross-sectional plane along air f ow, must be determined from Eq. (17.11): ps  pt  pv. Here pv represents velocity pressure. As the atmospheric pressure is assumed po  0 for the con venience of analysis, static pressure at the mixing box pm must be negative in order to e xtract outdoor air. The total pressure dif ference between the outdoor air and the air in the mix ed plenum (mixing box) pom  po  pm consists mainly of the pressure loss of the outdoor dampers and louv er. If a space is maintained at a positi ve pressure, the volume f ow rate of e xf ltration from the space V˙exf , in cfm (L / s), can be calculated from Eq. (20.11). If space is maintained at a ne gative pressure, the volume f ow rate of in f ltration into the space V˙inf , in cfm (L / s), can be similarly calculated. From Eq. (20.11), the pressure loss across the building shell pro, in in. WC (Pa), can be calculated as pro  pr  po 

 40051 A  V˙ 2

e,l

2 exf

 R exfV˙ 2exf

(22.4)

The f ow resistance against exf ltrated air across building shell Rexf, in in. WC / (cfm)2 (Pa s2 / m6), is R exf 

 40051 A 

2

e,l

AIR SYSTEMS: FAN COMBINATION AND SMOKE CONTROL

22.9

Because the v olume f ow rates of the supply air and recirculating air are dif ferent, the f ow resistance of the supply system and return system as well as the f ow resistance of the f xed part and variable part of the air system pressure loss should be calculated separately. Operating Characteristics Consider a VAV rooftop packaged system with a supply f an and an e xhaust fan serving a typical f oor in a commercial b uilding, as shown in Fig. 22.2. This air system has the follo wing operating characteristics at design conditions: Supply volume f ow rate Total pressure losses: Across recirculating damper at design f ow Filters and coils Supply main duct, pd, var Recirculating system VAV box, branch duct, and diffuser Effective leakage area on building shell Volume f ow of exhaust fan Minimum outdoor ventilation air required Space pressure

20,000 cfm (9438 L / s) 0.4 in. WC (100 Pa) 2.5 in. WC (625 Pa) 0.85 in. WC (212 Pa) 0.15 in. WC (37 Pa) 0.75 in. WC (187 Pa) 0.5 ft2 (0.46 m2) 3000 cfm (1416 L / s) 3350 cfm (1581 L / s)  0.03 in. WG (7.5 Pag)

For simplicity, the difference in air densities between supply, return, and exhaust air is ignored. In minimum outdoor air recirculating mode, the pressure losses of the outdoor louver can be neglected. A barometric relief damper is mounted in the conditioned space and is opened when the space positive pressure is greater than 0.2 in. WG (50 P ag). The supply f an has the same ptV˙ characteristics as in Example 20.3. The following control systems are installed to maintain the required operating parameters: ●

●

●

●

●

Zone temperature control Discharge temperature control Duct static pressure control Minimum outdoor ventilation air control High pressure limit control

All these control systems are described in Chap. 23. In summer, this air system can be operated in cooling mode, recirculating mode, air economizer mode, or warm-up or cool-down mode. Recirculation Mode and Design Volume Flow Minimum outdoor air is extracted through the outdoor intake and is mixed with the recirculating air in the mixing box in a rooftop packaged unit. The mixture is conditioned in the packaged unit. After conditioning, the supply air f ows through the supply f an and the supply duct and is dischar ged to the conditioned space. In the conditioned space, a certain volume f ow rate is exhausted through an exhaust fan, and another small portion is e xf ltrated through the gaps and openings on the b uilding shell if the space is maintained at a positi ve pressure between  0.005 and  0.03 in. WG (1.25 and 7.5 P ag). The major portion of the space air is returned to the rooftop unit, where it is mix ed with the outdoor air for recirculation.

22.10

CHAPTER TWENTY-TWO

FIGURE 22.2 Air system of supply and e xhaust fan combination: (a) schematic diagram; ( b) system characteristics on diagram. pt  V˙

AIR SYSTEMS: FAN COMBINATION AND SMOKE CONTROL

22.11

When the air system of a single supply f an and exhaust fan combination is operated in recirculating mode with design volume f ow, the inlet vanes at the supply fan inlet are fully opened and the outdoor air damper is at minimum opening position. The minimum outdoor ventilation air required V˙o is about 3350 cfm (1581 L / s). The f xed part of the pressure loss of such an air system is pf x  0.75 in. WC (187 Pa). The variable part of pressure loss of the air system consists of the follo wing two sections: ●

Mixed plenum point m to the static pressure sensor , point d, at a supply v olume f ow rate of 20,000 cfm (9438 L / s): pmd  2.5  0.85  3.35 in. WC (835 Pa)

Its corresponding f ow resistance can be calculated as R md  ●

pmr 3.35   8.38  109 in. WC / (cfm)2 (4.42  103 Pa  s2 / m6) (20,000)2 V˙ 2s

Recirculating system from point r to m: prm  R rmV˙ 2ru  0.15  0.4  0.55 in. WC (137 Pa)

The variable part of pressure loss of the air system is pmd,rm  pmd  prm  R mdV˙ 2s  R rmK 2rmV˙ 2s  R md, rmV˙ 2s

(22.5)

Because of a positi ve space pressure of  0.03 in. WG ( 7.5 Pag) and an ef fective leakage area Ae, l of 0.5 ft2, the volume f ow rate of exf ltration from the conditioned space V˙exf is calculated as V˙exf  4005 Ae, l√pro  4005  0.5√0.03  350 cfm (165 L / s) And the f ow resistance of the opening in the building shell is R exf 

pro 0.03   2.45  107 in. WC / (cfm)2 (0.129 Pas2 / m6) (350)2 V˙ 2exf

The recirculating volume f ow rate V˙ru through the recirculating system is then equal to V˙ru  V˙s  (V˙exf  V˙ex)  20,000  (350  3000)  16,650 cfm (7857 L / s) The f ow resistance of the recirculating system based on pressure loss between points r and m is R rm 

prm 0.55   1.98  109 in. WC / (cfm)2 (1.04  103 Pa  s2 / m6) (16,650)2 V˙ 2rm

And the f ow resistance of the variable part of the pressure loss of the air system is R md,rm  R md  R rmK 2rm  8.38  109  1.98  109

 16,650 20,000 

2

 9.75  109 in. WC / (cfm)2 (5.14  103 Pa s2 / m6) The supply fan total pressure at design volume f ow rate is psf  0.75  2.5  0.85  0.15  0.4  4.65 in. WC (1156 Pa)

22.12

CHAPTER TWENTY-TWO

The system operating point of this supply and e xhaust fan combination P is plotted in Fig. 22.2 b. The pressure-volume characteristics for the key points in this system are shown below:

p, in. WG V·, cfm

Point o

Point m

Point r

Point ru

0 3350

 0.52 20,000

 0.03 20,000

 0.12 16,650

Recirculation Mode, 50 Percent Design Flow Rate When the space sensible cooling load is reduced at part-load operation, the supply volume f ow rate may be reduced to 50 percent of the design v olume f ow rate, that is, 0.5  20,000  10,000 cfm (4719 L / s). As the dampers in the VAV boxes close to smaller openings than at design v olume f ow, the static pressure in the supply main duct rises. The sensor senses the rise, and the DDC unit instructs the inlet vanes to close to a smaller opening, until the duct static pressure is maintained approximately at its set point at the location where the pressure sensor is mounted. The static pressure is still equal to the pressure loss of the VAV box, f exible ducts, diffuser in the branch takeoff, and last section of the main duct between the pressure sensor and end of the main duct plus the space static pressure: 0.75  0.03  0.78 in. WC (194 Pa). The fan performance curve is now a new fan curve Ft50, as shown in Fig. 22.2b. The variable-part pressure loss between the mixing box and point d is now decreased to pmd50  R mdV˙ 2s50  8.38  109(10,000)2  0.84 in. WC (209 Pa) Assume that the space pressure at 50 percent design f ow is  0.03 in. WG. The volume f ow rate of inf ltrated air into the space is V˙inf  4005 Ae, l√p  4005  0.5√0.03  347 cfm (164 L / s) because the v olume f ow rate of an e xhaust fan is closely related to its pressure. According to the catalog of a f an manufacturer, within the operating range of a 24-in. (600-mm) diameter e xhaust fan, for an increase of 0.01 in. WC (2.5 P a) in f an total pressure, there is a corresponding drop in supply volume f ow rate of 135 cfm (64 L / s). Compared with the design conditions, there is a need for an increase of  0.03  ( 0.03)  0.06 in. WC (15 P a) of f an total pressure of the e xhaust fan, i.e., a decrease of 0.06 / 0.01  135  810 cfm of v olume f ow rate. The volume f ow rate of recirculating air is then V˙ru50  V˙s50  V˙ex  V˙inf50  10,000  (3000  810)  347  8157 cfm (3849 L / s) The pressure loss between the space (point r) and the mixed plenum (point m) becomes prm  R rmV˙ 2rm50  1.98  109(8157)2  0.132 in. WC (33 Pa) Therefore, the total pressure in the mixing box is pm   0.03  0.132   0.162 in. WG ( 40 Pag) The f ow resistance of the outdoor passage is R om 

pom  V˙ 2o

0.52  3350   4.63  10 2

8

in. WC / (cfm)2 (0.0244 Pas2 / m6)

Because of a pressure difference pom  0.162 in. WC, outdoor intake is therefore V˙o 

√

pom  R om

√

0.162  1870 cfm (883 L / s) 4.63  108

AIR SYSTEMS: FAN COMBINATION AND SMOKE CONTROL

22.13

If the difference in air density is ignored, the volume f ow balance between the incoming in f ltrated air and outward exhaust air in this air system becomes V˙o  V˙inf  V˙ex  1870  347  (3000  810)   27 cfm (13 L / s) The v olume f ow of incoming air and e xhaust air entering and lea ving the system is nearly balanced. From the above analysis, the system operating point at 50 percent design volume f ow rate, point Q, has a supply volume f ow of 10,000 cfm (4719 L / s) and a fan total pressure pt50  0.75  0.84  0.13  1.72 in. WC (428 Pa) During minimum outdoor air recirculating mode at 50 percent design v olume f ow, the volume f ow rate of outdoor air e xtracted through the outdoor intak e passage decreases to 1870 cfm (883 L / s). Space is maintained at a negative pressure of  0.03 in. WG ( 7.5 Pag). An inf ltration rate of 347 cfm (164 L / s) enters the space through openings and cracks in the building shell. The ptV˙ characteristics at key points during 50 percent design volume f ow are as follows:

p, in. WG V·, cfm

Point o

Point m

Point r

Point ru

0 1870

 0.162 10,000

 0.03 10,000

 0.065 8157

Air Economizer Cycle When an air economizer cycle is operated at a design f ow rate of 100 percent of outdoor air in this air system with a single supply f an and an e xhaust f an combination, the recirculating damper is closed, and both the minimum outdoor damper and the other, large outdoor damper for the air economizer cycle are fully opened. Outdoor air is e xtracted to the mixed plenum, through the f lters and coils, and is supplied to the conditioned space. Because only a small portion of supply air is e xhausted and e xf ltrated, the space positi ve pressure increases to about 0.2 in. WG (50 P ag). The barometric relief damper then opens. Most of the supply air is dischar ged through the relief damper outdoors. When the supply and e xhaust fan combination is operated at 50 percent design v olume f ow in the air economizer cycle, space pressure is still limited to 0.2 in. WG (50 Pa), and most of the supply air is discharged through the relief damper outdoors, as in the design volume f ow.

Warm-Up and Cool-Down Mode During the w arm-up and cool-do wn mode, the outdoor damper is closed and the recirculating damper is fully open, while the exhaust fan is turned of f. Recirculated air from the space is conditioned in the packaged unit and is then supplied to the space. Because of the ne gative pressure in the mixing box, outdoor air is leak ed into the box through the closed outdoor dampers and is e xf ltrated from the space because of the higher positive space pressure.

Pressure Variation at the Mixing Box During the recirculating mode, when the supply volume f ow rate is reduced from design f ow to 50 percent of design f ow, the decrease in the recirculating v olume f ow rate causes a corresponding drop in the pressure loss of the recirculation system. Therefore, the pressure at the mix ed plenum, pm increases from  0.52 to  0.16 in. WG ( 129 to  40 Pag). The greater the pressure loss of the return system at the design f ow rate, the higher the f uctuation of pom. Because of the reduced volume f ow rate at a higher f an total pressure of the e xhaust fan at 50 percent design f ow, in order

22.14

CHAPTER TWENTY-TWO

to balance the airf ow, the space pressure is changed from  0.03 to  0.03 in. WG ( 7.5 to  7.5 Pag), and the exf ltration changes to in f ltration. The pressure loss at the outdoor air intak e passage (excluding the damper), like the intak e louv er and duct friction loss plou, should be tak en into account only in the outdoor air economizer c ycle. During minimum outdoor air recirculating mode, plou should be negligible. Pressure f uctuation pom at design f ow and at 50 percent design f ow causes insuf f cient outdoor air intak e. The effect of this pressure f uctuation can be reduced by using minimum outdoor ventilation air control, as discussed in Chap. 23.

22.5 SUPPLY FAN AND RELIEF FAN COMBINATION Consider an air system with a supply f an, relief f an, and e xhaust f an combination, as sho wn in Fig. 22.3a, with the same operating characteristics as mentioned in Sec. 22.4. A relief f an is a f an that is installed in the relief f ow passage (exhaust f ow passage), adjacent to the recirculating f ow passage leading to the mixing box, to relieve the undesirably high positi ve space pressure. A relief fan is different from an exhaust fan, which is often used to exhaust fumes and contaminated or toxic gases and to maintain a negative pressure in the space or in a localized enclosure. An axial fan is often used as a relief f an because of its lar ge volume f ow and smaller f an total pressure. The control systems used in such a combination are similar to those of a supply f an and e xhaust fan combination, except that a relief fan control is added. The relief fan has the following pressure-volume characteristics: V·, cfm pt, in. WC

10,000

15,000

16,640

17,250

4.0

2.10

1.08

0.43

Recirculation Mode During minimum outdoor air recirculating mode operation, the relief f an is not operating and the relief damper is closed. Outdoor air drawn into the air system is balanced by the e xhaust air and the exf ltration through the openings in the building shell. In this case, the system becomes a supply fan and exhaust fan combination, and its operating characteristics are the same as those described and analyzed in Sec. 22.4.

Air Economizer Mode and Design Volume Flow Rate In a VAV system, there are often tw o outdoor air dampers: a smaller minimum outdoor air damper and a larger economizer damper. Both dampers have a total pressure loss of 0.4 in. WC (100 Pa) at 100 percent outdoor air intake. During air economizer mode with 100 percent outdoor air at the design volume f ow rate, the outdoor dampers are fully opened and the recirculating damper is closed. The relief fan relieves the space air pressure and maintains it within predetermined limits. Outdoor air is extracted through the intake louver and both outdoor dampers, and f ows through the f lter and coils. The conditioned air is then supplied to the space. A small part of the space air is e xhausted, and another part is e xf ltrated. The majority of space air is relie ved by the relief f an through the ceiling plenum and relief passage. At design v olume f ow, the duct friction loss of the outdoor intak e passage is ne gligible, and the pressure loss of the intak e louver is 0.2 in. WC (50 Pa). The space pressure depends mainly on the airf ow balance in the space. The variable part of the pressure loss of the supply air system is then given as pvar  0.2  0.4  2.5  0.85  0.03  3.98 in. WC (990 Pa)

FIGURE 22.3 Supply fan and relief f an combination: (a) schematic diagram; ( b) system characteristics on pt  V˙ diagram.

22.15

22.16

CHAPTER TWENTY-TWO

The f xed part of pressure loss of the air system is still 0.75 in. WC (186 Pa). The fan total pressure of the supply fan is then equal to 0.75  3.98  4.73 in. WC (1176 Pa) The f ow resistance of the variable part of the supply air system can be calculated as R or 

por 3.98   9.95  109 in WC / (cfm)2 (5.24  103 Pa  s2 / m6) (20,000)2 V˙ 2s

The f an total pressure of the relief f an is used to o vercome the pressure loss of the return system, relief damper , and louv er in the relief flo w passage. Usually , the duct friction loss of the relief passage is ignored. At design volume flow rate, for a space pressure of  0.03 in. WG ( 7.5 P ag), an e xfiltration of about 350 cfm (165 L / s), and an e xhaust v olume flo w rate of 3000 cfm (1416 L / s), the volume flow rate of relief f an is 20,000  3000  350  16,650 cfm (7857 L / s). As in the supply f an and e xhaust fan combination, the pressure loss between points r and ru is pr,ru  R r, ruV˙ 2ru  0.15 in. WC (37 Pa) and f ow resistance is R r,ru 

pr,ru  5.41  1010 in. WC / (cfm)2 (2.85  104 Pa s2 / m6) V˙ 2ru

If the pressure loss of the relief louv er at a v olume f ow of 16,650 cfm (7857 L / s) is 0.2 in. WC (50 Pa), the pressure loss of the relief damper is 0.4 in. WC (100 P a), and the v elocity pressure of the axial fan is 0.38 in. WC (95 Pa), the f ow resistance of the relief system is R rel 

0.2  0.4  0.38  3.54  109 in. WC / (cfm)2 (1.86  103 Pa  s2 / m6) (16,650)2

The sum of f ow resistance of the return and relief systems is R ro  R r,ru  R rel  5.41  1010  3.54  109  4.08  109 in. WC / (cfm)2 (2.15  103 Pa  s2 / m6) Because a total pressure of  0.03 in. WG (7.5 P ag) is pro vided by the space positi ve pressure and is undertak en by the supply f an, therefore, the f an total pressure of the relief f an is calculated as prel  0.2  0.4  0.38  0.15  0.03  1.10 in. WC (274 Pa) The supply f an curv e Fs, the relief f an curv e Frel, the system curv e of the supply system pt  R orV˙2, and the system curve of the return and relief system pt  R roV˙2 can then be plotted. At design f ow of an air economizer c ycle, the system operating point for supply system P has V˙  19,750 cfm (9320 L / s) and pt  4.73 in. WC (1176 P a). The system operating point of the relief system has V˙  16,650 cfm (7857 L / s) and pt  1.10 in. WC (273 Pa). The pressure-v olume characteristics at k ey points for air economizer mode at design v olume f ow are as follows:

p, in. WG V·, cfm

Point o

Point m

Point r

Point ru

0 19,750

 0.6 19,750

 0.03 19,750

 0.12 16,650

AIR SYSTEMS: FAN COMBINATION AND SMOKE CONTROL

22.17

Air Economizer Mode, 50 Percent Design Flow At 50 percent design volume f ow rate in the air economizer mode, the variable part of the pressure loss of the supply air system when inlet vanes are partly closed is pvar50  9.95  10 9(10,000)2  1.0 in. WC (249 Pa) And the fan total pressure of the supply fan at 50 percent design f ow is psf50  1.0  0.75  0.03  1.78 in. WC (443 Pa) Total pressure loss of the return and relief systems is then pr, ru50  4.08  10 9(6650)2  0.18 in. WC (45 Pa) At 50 percent design f ow, the velocity pressure of the relief fan is pv,rf 

6650  16,650  (0.38)  0.06 in. WC (15 Pa) 2

The fan total pressure of the relief fan at 50 percent design f ow is therefore pref50  0.18  0.03  0.06  0.21 in. WC (52 Pa) Pressure-volume characteristics at k ey points in air economizer mode at 50 percent design v olume f ow are as follows:

p, in. WG V·, cfm

Point o

Point m

Point r

Point ru

0 10,000

 0.15 10,000

 0.03 10,000

 0.02 6650

Warm-Up and Cool-Down Mode During warm-up and cool-do wn mode, the relief f an is turned of f and the relief damper is closed. As in the supply f an and e xhaust fan combination, the exhaust fan is also turned of f, the outdoor damper is closed, and the recirculating damper is fully opened. Air is recirculated from the space to the AHU or PU and to the space again.

Design Considerations and Controls ●

●

●

As in the supply f an and exhaust fan combination system, control systems such as zone temperature, discharge temperature, duct static pressure, minimum outdoor air , and high-pressure limit control systems are discussed in Chap. 23. The operation of recirculating mode or air economizer mode is actuated by the air economizer control which is either temperature-based or enthalp y-based, as described in Sec. 21.2. Warm-up and cool-down mode is acti vated when the sensed zone temperature de viates from the set point after the AHU or PU is turned on, as described in Chap. 23. At minimum outdoor air recirculating mode, when the relief f an is not operating, the space positive pressure is maintained because of the air f ow balance in the space. Suppose that 20,000 cfm (9438 L / s) of air is supplied to the space. If the space pressure is zero, as 3000 cfm (1416 L / s) of air is e xhausted from the space by the e xhaust fan, only 16,650 cfm (7857 L / s) of recirculating

22.18

CHAPTER TWENTY-TWO

●

●

air is extracted by the supply fan to mix with the 3350 cfm (1581 L / s) of outdoor air in the mixed plenum at design v olume f ow. The excessive amount (350 cfm or 165 L / s) of supply air raises the space pressure until the 20,000 cfm (9438 L / s) supply air is balanced with the sum of 16,650 cfm (7857 L / s) recirculating air, 3000 cfm (1416 L / s) exhaust air, and 350 cfm (165 L / s) exf ltrated air at  0.03 in. WG (7.5 Pag) space positive pressure. The relief fan is often controlled by a DDC unit controller according to the input from the space pressure sensor to maintain a space positive pressure within predetermined limits. The main dra wback of the supply and relief f an combination is the increase in pressure at the mixed plenum pm when the supply volume f ow rate of a VAV system is reduced at part-load operation during minimum outdoor air recirculating mode. The higher the pressure loss of the return system, the greater the increase in pm. Higher pressure in the mix ed box and, therefore, a lower pom ( po  pm ) cause a def ciency in outdoor air intake. Various outdoor ventilation controls have been developed recently and are discussed in Chap. 23. The operation of a relief f an reduces both mixing box pressure and space pressure.

22.6 SUPPLY FAN AND RETURN FAN COMBINATION Supply fan and return f an combinations are widely used, especially in air systems that serv e large conditioned areas. A return fan is located upstream from the junction of the recirculating f ow passage and the e xhaust air passage, point ru, as shown in Fig. 22.4 a. Consider an AHU or PU with a supply fan and return f an combination and an e xhaust system. Its operating characteristics are the same as those in Sec. 22.4 except for the following: ●

A return fan is installed with the following pressure-volume characteristics: V·, cfm pt, in. WC

●

10,000

15,000

20,000

1.47

1.23

1.10

The pressure loss of the e xhaust damper after point ru in the exhaust passage is 0.4 in. WC (100 Pa) and is 0.2 in. WC (50 P a) for the air louv er at the maximum e xhaust volume f ow rate. At minimum outdoor air f ow, louver pressure losses can be ignored. As in the supply and relief f an combination, a minimum outdoor air damper and an economizer damper are installed, and both have a total pressure loss of 0.4 in. WC (100 Pa) at design airf ow.

Recirculation Mode During minimum outdoor air recirculation mode, outdoor air is dra wn through the outdoor air louver, duct, and damper and is mixed with recirculating air. The mixture then f ows through the f lters and coils and is supplied to the conditioned space. At design volume f ow, a small part is exhausted, another part is e xf ltrated, and the remaining portion is e xtracted by the return f an f owing through the return system. At point ru, almost all the return air is recirculated through the recirculating damper. During minimum outdoor air recirculating mode, at design volume f ow rate: ●

●

If the return f an is not operating and the e xhaust damper is closed, then the f an total pressure of the supply fan is higher than 4.65 in. WC (1156 Pa) because of the air f owing through the turnedoff return fan. If the return fan is operating, 3000 cfm (1416 L / s) of air is exhausted through the exhaust passage and exhaust damper instead of from the separate e xhaust fan. The total pressure loss of the return

FIGURE 22.4 Supply fan and return fan combination: (a) schematic diagram; (b) system characteristics on pt  V˙ diagram.

22.19

22.20

CHAPTER TWENTY-TWO

system pret increased from 0.15 to 0.7 in. WC (37 to 74 Pa). The fan total pressure of the supply fan is then 0.2  0.4  2.5  0.85  0.75  0.03  4.73 in. WC (1176 Pa) During the minimum outdoor air recirculating mode, the recirculating air turns 90 ° and f ows through the recirculating damper and has a greater total pressure loss than the 3000 cfm (1416 L / s) exhaust air exhausted through exhaust damper and louver. Assume that the dynamic loss of the 90 ° turn of the recirculating air is 0.2 in. WC (50 P a). Then the total pressure loss of the recirculating damper is 0.4 in. WC (100 Pa) and the fan total pressure of the return fan is 0.7  0.2  0.4  0.03  1.27 in. WC (316 Pa) The pressure-volume characteristics at the key points are shown below: pret, in. WC 0.15 0.15 0.70 0.70

p, in. WC V·, cfm p, in. WG V·, cfm

Point o

Point m

Point r

Point ru

0 3,350 0 3,350

  0.52 20,000  0.6 20,000

 0.03 20,000  0.03 20,000

  0.12 16,650 1.0 19,650

At 50 percent of the design f ow rate, if the return f an is turned of f and the e xhaust damper after point ru is closed, then the fan total pressure of the supply fan will be higher than 1.72 in. WC (428 Pa), as in the supply f an and e xhaust fan combination. The space pressure pr drops to  0.03 in. WG (  7.5 Pag). Total pressure at point ru may be lo wer than  0.07 in. WG (  17 Pag) and at point m lower than  0.26 in. WG ( 65 Pa) because of the pressure loss of air f owing through the turned-off return fan. If the total pressure loss of the return system is at 0.7 in. WC (174 Pa) and 3000 cfm (1416 L / s) of air is exhausted through the return fan instead of the exhaust fan, and also if the return f an is operating at 50 percent of design f ow rate, then the f an total pressure of the return f an is about (10,350 / 20,000)2 (0.67  0.2  0.4)  0.34 in. WC (85 P a) and the total pressure at ru is about 0.16 in. WG (40 Pag). Air Economizer Mode During air economizer mode operation at the design v olume f ow rate, the supply and e xhaust volume f ow rate and the pressure losses of various sections of the supply air system are the same as in minimum outdoor air recirculating mode. The fan total pressure of the supply f an is 4.73 in. WC (1176 Pa). Because the total pressure at point ru is raised to 1.0 in. WG (249 Pag) to overcome the pressure loss of the e xhaust f ow passage and the dischar ge velocity pressure at the outlet of 0.40 in. WC (100 Pa), the return fan total pressure is then pret.f  0.15  0.03  0.60  0.40  1.12 in. WC (278 Pa) If the total pressure loss of the return system is increased to 0.70 in. WC (174 Pa) and the return air is exhausted through the return f an and e xhaust damper, during air economizer mode operation at design f ow rate, then the return fan total pressure is the same as in the recirculating mode pret.f  (0.70  0.03)  0.60  0.40  1.67 in. WC (415 Pa) Pressure-volume characteristics at the k ey points in air economizer mode at design v olume f ow, when the total pressure loss of the return system is 0.15 in. WC (37 Pa) and space air is e xhausted through the exhaust fan, are as follows:

AIR SYSTEMS: FAN COMBINATION AND SMOKE CONTROL

p, in. WG V·, cfm

Point o

Point m

Paint r

Point ru

0 20,000

 0.6 20,000

 0.03 20,000

1.0 16,650

22.21

During air economizer mode, if the total pressure loss of the return system is increased to 0.70 in. WC (174 P a) and the return air is e xhausted through the return f an and e xhaust damper, at 50 percent of design volume f ow, then the fan total pressure of the return fan is pret.f50 

9650  20,000  (0.70  1.0)  0.03  0.37 in. WC (91 Pa) 2

and the pressure-volume characteristics at the key points are as follows:

p, in. WG V·, cfm

Point o

Point m

Point r

Point ru

0 10,000

 0.15 10,000

 0.03 10,000

0.23 9650

Controls As in the supply and relief f an combinations, the controls for supply and return f an combination systems operated at minimum outdoor air recirculating mode and air economizer mode operations are discussed in Chap. 23.

22.7 COMPARISON OF THREE FAN COMBINATION SYSTEMS An air system with supply and e xhaust fan combination is simpler and less e xpensive. Such a system is not suitable for operation in air economizer mode for a high space pressure. In a supply and exhaust fan combination, a comparatively lower pressure drop in the return system is necessary to prevent an unacceptable low pressure in the mixed plenum, such as pm   1 in. WG ( 250 Pag). Such lo w mix ed plenum pressure induces air leakage and may impede minimum outdoor air control. The operating characteristics of the supply and relief f an combination and the supply and return fan combination can be compared as follows: ●

●

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●

In a supply and relief f an combination in recirculating mode, total pressure at point ru may be negative; in a supply and return f an combination, total pressure at ru must be positive, in order to overcome the pressure loss of the exhaust passage. With a relief f an, the mixed plenum pressure pm is always negative. With a return f an, pm may be positive if the return f an is too lar ge. A positive pm means that there will be no outdoor air intak e into the mixing box. Such a condition should never occur. When a negative space pressure must be maintained, a relief fan causes the pressure in the mix ed plenum to be smaller (more ne gative) than a return f an during minimum outdoor air recirculating mode. When there is a considerable pressure drop in the e xhaust or relief f ow passage, a return f an requires high positive pressure at point ru, so energy is wasted across the recirculating damper. A supply and relief f an combination with a lo wer total pressure loss return system is often more energy-eff cient than a supply and return f an combination during recirculating mode operation, for the follo wing reasons: (1) The pressure drop across the recirculating damper is smaller .

22.22

CHAPTER TWENTY-TWO

●

(2) The relief f an heat gain does not increase coil load. (3) If the return f an is not operating and there is no bypass passage, there is an additional pressure drop across the idle return fan. An axial relief f an is sometimes simpler in layout and installation, and is therefore less e xpensive to install, than a centrifugal return fan. If an axial fan is used, noise attenuation must be considered.

Therefore, for those air systems with a smaller total pressure loss in return duct system, usually less than 0.5 in. WC (125 Pa), especially for those systems also with a considerable pressure drop in the exhaust or relief f ow passage, a supply and relief fan combination is recommended. For air systems with a greater total pressure loss in the return system, or those that require a negative space pressure, a supply and return fan combination may be more appropriate.

22.8 PRESSURE FLOW CHARACTERISTICS Pressure Flow Characterisitcs for a Supply and Return Fan Combination System During a supply and return f an combination VAV system design, it is critical to ha ve low enough negative static pressure in the mixing box to e xtract the required amount of outdoor v entilation air and a positive pressure in the return chamber (junction of the return system and e xhaust system) to exhaust the proper amount of return air to outdoors to maintain a predetermined space pressure. According to the calculations and analyses in the previous sections, to achieve these: ●

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Correctly divide the system total pressure loss into tw o parts: (1) that part undertaken by the fan total pressure of the return fan, which starts from the return slot inlet rt and includes obstructions in the ceiling plenum, if an y, return ducts and f ttings, 90° elbow in the return chamber , the positve pressure required in the return chamber to e xhaust return air through the exhaust damper and louver or the recirculating damper, as shown in Fig. 22.1; and (2) that part undertaken by the fan total pressure of the supply f an which starts from the ne gative pressure in the mixing chamber, total pressure loss of the f lter, coils, direct-f red furnace, 90° elbow, supply duct and duct f ttings, f exible ducts, VAV box, diffuser, and the positi ve pressure maintained in the conditioned space. For a rooftop unit with system components as sho wn in Fig. 22.1, the total pressure in the return chamber point ru must be high enough to e xhaust part of the return air from the e xhaust damper and louver, or to overcome the total pressure loss of a 90 ° elbow of air f ow and the recirculating damper during the recirculating mode at summer design conditions. The total pressure loss of the 90° turn and the recirculating damper is often greater. In VAV systems, they are often installed with tw o outdoor dampers: minimum outdoor damper and 100 percent outdoor damper . During the minimum outdoor air recirculating mode, only the minimum outdoor air damper is acti vated. The total pressure loss of the minimum outdoor 0.4  0.2 = 0.6 damper and louv ers is still assumed to be 0.4  0.2 = 0.6 in. WC (150 P a) or less for effective control and ener gy saving. Then a ne gative pressure of 0.6 in. WC (150 P a) or less is needed during minimum outdoor air recirculating mode. During the air economizer mode, 100 percent outdoor air is required at the design f ow rate. Both outdoor air dampers are opened. A negative pressure of 0.6 in. WC (150 P a) or less is required to extract the required amount of outdoor air. For a supply and return f an combination VAV system, accurately calculate the total pressure loss of the system components, supply and return ducts and duct f ttings, f exible ducts, VAV box, and diffusers. It is necessary to dra w a system pressure diagram for a supply and return f an combination system and ensure that the ne gative pressure, at a lo w enough magnitude, is in the mixing box and that there is adequate positive pressure at the junction of the recirculating passage and the exhaust passage, point ru (return chamber, see Fig. 22.1).

AIR SYSTEMS: FAN COMBINATION AND SMOKE CONTROL

●

22.23

System pressure diagram should be dra wn by system total pressure. During the construction of a system pressure diagram of a supply f an, the static re gain in the supply duct is obtained at the expense of the velocity pressure of the supply f an. Only during the minimum outdoor air recirculating mode is the velocity pressure of the return fan utilized. It is often cost-ef fective to install adjustable-frequenc y variable-speed drive for both supply and return fans to adjust the fan speed for large fans during operation. Set the f an total pressure of the supply and return fans as well as the damper position during commissioning to achie ve a required system pressure distribution and an optimum energy use.

Variation of Pressure in the Mixing Box The negative total pressure in the mixing box  pm is related to the upstream outside atmospheric pressure po, all in in. WG (Pa), as follows:  pm  po  pom

(22.6)

where pom  pressure dif ferential between the outside atmosphere and the air pressure in the mixing box, in. WC (Pa). The negative total pressure in the mixing box  pm is also related to the outside atmospheric pressure po through the space pressure  pr, in in. WG (Pag), by the following relationship: pm  po  por  pr,s  prec,d pm  pr  pr,s  prec,d

(22.7)

where por  pressure differential between outside atmosphere and space, in. WC (Pa) pr,s  total pressure loss of return system, in. WC (Pa) prec,d  total pressure loss of recirculating damper, in. WC (Pa) During recirculating mode, part-load operation, the volume f ow rates of the supply and return air are reduced, and the total pressure loss of the return system pr,s reduces accordingly. At the same time, the variation of the mass f ow rates of the supply , return, exhaust, and exf ltrated air forms new airf ow balances and ne w space pressures. Because of the relationship between  pm, pr, and po, as shown in Eq. (22.7), the result is often a higher  pm (less negative pm) and less extracted outdoor ventilation air through the outdoor dampers at recirculating mode during part-load operation, as analyzed in Secs. 22.4, 22.5, and 22.6. The opening of the outdoor and recirculating dampers should be adjusted in order to extract the required amount of outdoor air. Field Survey of System Pressure Characteristics of a VAV System Using a Supply and Relief Fan Combination Figure 22.5 shows the pressure characteristics of a VAV system using a supply and relief fan combination. This VAV cooling system serv es the f fth f oor of a high-rise public library with a conditioned f oor area of about 16,500 ft 2 (1534 m 2). There are two AHUs and two fan rooms, each with a volume f ow rate of about 12,650 cfm (5970 L / s) and a f an total pressure of 4.25 in. WC (1057 Pa) to serv e this f oor. Each f an room has three relief f ans and three relief dampers. The volume f ow rate of relief air is modulated by relief dampers, which are controlled by an electronic controller actuated by a signal from a pressure sensor in the ceiling plenum. In the perimeter zone, winter heating is provided by the electric heating coil located downstream from the VAV box. On September 12, 1983, the system pressure characteristics of the VAV cooling system were measured. Control of relief dampers w as deliberately deenergized. The supply volume f ow rate of AHU 2 measured w as about 12,000 cfm (5663 L / s). The pressure dif ferences between v arious points in f an room 2 in the VAV cooling system during minimum outdoor air recirculating mode and 100 percent air economizer cycle were as follows:

22.24

CHAPTER TWENTY-TWO

FIGURE 22.5 Pressure characteristics of a VAV cooling system using supply-relief fan combination.

Outdoor damper Outdoor air, cfm Recirculating damper Relief fans operated Relief damper ppl,o in. WC (Pa) ppl,m in. WC (Pa) po,m in. WC (Pa) pr,pl in. WC (Pa)

Recirculating mode

Air economizer

10% open 3330 100% open 1 fan 15% open  0.06 (15)  0.20 (50)  0.14 (35)  0.004 (1)

100% open 100% closed All 3 fans 100% open  0.018 (4.5)  0.004 (1)

During the air economizer c ycle, when the outdoor damper is fully open and the recirculating damper w as closed, the pressure dif ference between the ceiling plenum and outdoors ppl,o was  0.038 in. WC (9.5 P a) when only tw o relief f ans were operating, and ppl,o was  0.04 in. WC (10 Pa) when only one relief fan was operating. The pressure dif ference between the conditioned space and the air in the ceiling plenum pr,pl was so small because of gaps between the ceiling tiles. During recirculation mode, the space pressure is about  0.064 in. WG (16 Pag). Space pressure was lower when more relief fans were operating during the air economizer cycle.

22.9 SMOKE CONTROL AND FIRE SAFETY Fire Safety in Buildings Fire safety is a critical design f actor in high-rise b uildings. On No vember 21, 1980, in Las Vegas, the MGM Grand Hotel f re took 85 li ves. Smoke inhalation is the primary killer in b uilding f res. According to the annual surv ey tak en by the National Fire Protection Association (NFPA), there was a slight increase in ci vilian f re deaths to a total of 4585 people who lost their li ves in 1995,

AIR SYSTEMS: FAN COMBINATION AND SMOKE CONTROL

22.25

which is a 7.3 percent increase o ver a record low in 1994. In 1995, in the United States, public f re departments responded to approximately 2 million f res, and property damage due to f re was nearly $9 billion. McGreal (1997) noted that b uilding f res produce both smoke and heat. Smoke is the real killer. Smoke consists of the airborne particulates, in the form of either solid or liquid, and gases which evolve when a material undergoes pyrolytic combustion. Studies around 1980 found that more than 80 percent of f re deaths were due to the inhalation of f re combustion products. About one-half of the deaths are due to the inhalation of carbon monoxide (CO). Another 16 percent of the deaths were caused by the combined ef fect of CO and hydrogen cyanide and heart disease. Fire protection and f re safety in buildings include the following measures: ●

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Fire compartmentalization Fire-resistant construction Fire alarm system Automatic sprinkler system Smoke control system including stairwell pressurization and zone smoke control system Fire protection management and coordination

Smoke control systems, which include stairwell pressurization and zone smoke control, are features of HVAC&R system design and are discussed here. National codes and local codes must be follo wed during the design of the smok e control system in buildings.

Smoke Movements in Buildings Figure 22.6 sho ws smok e mo vements in a 10-story e xperimental b uilding. According to ASHRAE Handbook 1999, HVAC Applications, the typical leakage area ratio Aleak /Af , which is the ratio of leakage area Aleak to f oor area Af , both in ft 2 (m2) for commercial b uildings of a verage tightness, is as follows: External walls, including cracks around windows and doors Stairwell walls, construction cracks only Elevator shaft walls, construction cracks only Floors, construction cracks, and area around penetrations

0.00021 0.00011 0.00084 0.000052

The cracks around the stairwell door can be typically tak en as 0.25 ft 2 (0.023 m 2), and the cracks around each elevator on each f oor can also be taken as 0.25 ft2 (0.023 m2). A building f re can be simulated by burning wooden sticks in a second-f oor corridor with a peak energy release of 900 Btu / s (950 kJ / s) for a certain period while the b uilding is under the follo wing operating conditions: ●

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●

The stair pressurization system and smoke control system are not operated. All stairwell doors are closed e xcept that on the second f oor, which is opened 12 in. (13 mm) by the high temperature of the hot gas. All windows and exit doors are closed. Outdoor temperature is 30°F ( 1.1°C), and there is a north wind at 5 mph (2.23 m / s).

22.26

CHAPTER TWENTY-TWO

FIGURE 22.6 Smoke movements in a 10-story experimental building.

When the f re is ignited, smoke mo vements are mainly caused by the e xpansion and b uoyancy forces of hot gas at a maximum temperature between 850 and 1000 °F (454 and 538 °C), the stack effect from outdoor-indoor temperature differences, and the wind effect. Smoke mo ves from f re on the second f oor to the upper f oors through stairwells, elevator shafts, service shafts, vertical risers, and f oor cracks; and smok e discharges to the outdoor atmosphere through windo w cracks, elevator machine-room openings, and other openings in the upper f oors. Outdoor air enters the building below the neutral plane and discharges to the outdoors above the neutral plane because of the stack ef fect. Oxygen supply to the f re enters through second- f oor window cracks, openings in the building envelope, and vertical air passages from the f rst and third f oors. According to Klote (1990), the CO 2 and CO le vels on the highest f oor of the e xperimental building during tests without stairwell pressurization and smok e control are 0.15 percent CO 2 and 0.015 percent CO. In tests with stairwell pressurization and smok e control, the levels are 0.002 percent CO2 and 0.001 percent CO. In winter, the stack effect assists the stairwell pressurization in pre venting the smoke from contaminating the stairwell. This result is verif ed by experiments in Tamura (1990b).

AIR SYSTEMS: FAN COMBINATION AND SMOKE CONTROL

22.27

Effective Area and Flow Rates In several airf ow passages connected in parallel, the effective area Ae, in ft2 (m2), can be calculated as Ae  A1  A2      An

(22.8)

where A1, A2,    , An  airf ow areas for path 1, 2,    , n, ft (m ) In airf ow paths connected in series, each with a f ow area A1, A2,    , An , in ft 2 (m2), the effective area for these airf ow paths connected in series Ae can be calculated as 2

Ae 

 A1

2 1

1 1    2 A22 An



2

1 / 2



(22.9)

If the f ow coeff cient is taken as 0.65 and air density a  0.075 lb / ft3 (1.2 kg / m3), the air volume f ow rate V˙ , in cfm, f owing through a crack, gap, or opening can be calculated as V˙  2610A(p)1 / 2

(22.10)

where A  f ow area or effective area, ft (m ) p  pressure difference across f ow path or opening, in. WC (Pa) 2

2

22.10 EFFECT OF AUTOMATIC SPRINKLER ON FIRE PROTECTION Automatic sprinkler systems are ef fective and reliable f re protection systems and should be installed in b uildings to pro vide f re protection for the occupants during a b uilding f re. Based on experimental results, Mawhinney and Tamura (1994) summarized the effect of automatic sprinklers on f re protection as follows: ●

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For unshielded f res, automatic f re sprinklers reduced both smok e and f re hazard to ne gligible levels. Unshielded f re is a f re exposed to the water sprays from activated sprinklers. For shielded f res with acti vated water sprays from the sprinklers, f re continues to b urn at a reduced rate and to produce smoke until the fuel inside the shielded area has all b urned. Concentrations of CO 2 and CO in the smok e were dangerously high. Shield f res represent f res that are shielded from the water sprays delivered by the sprinklers above because of a desk, table, or other furniture. According to e xperiments, shielded crib f res produced CO concentrations as high as 8 and 9 percent and CO concentrations between 1.0 and 1.5 percent (10,000 and 15,000 ppm) in a multistory b uilding. Peacock et al. (1989) indicated that a concentration of 15,000 ppm CO w ould present a severe life-safety hazard to any one trapped on the f re f oor. Sprinklers reduced the temperature and radiant heat from shielded f res to nonthreatening le vels within the boundary of the area co vered by the sprinkler system. F or a sprinkler system design that meets NFP A Standard 13 requirements, a moderate increase in v entilation to the f re f oor during a building f re should not signif cantly increase the f re temperatures. Buoyancy forces for shielded, sprinklered f re were almost ne gligible. The recommendation for the pressure dif ference between dif ferent smok e control zones ( f re zone and adjacent zones) p  0.05 in. WC (12.5 Pa) in NFPA (1988) for zone smok e control design in sprinklered b uildings (ceiling height  9 ft or 2.7 m) is more than suff cient to prevent smoke movement, provided that the door to the f re f oor remains closed. Recommended practice for zone smok e control design should allo w for some air f ow into the stairwell to pre vent the spread of smok e into the stairwell when the door to the f re f oor is opened.

22.28

CHAPTER TWENTY-TWO ●

The assumption that smoke will never become a threat to life safety in a fully sprinklered building needs to be reexamined.

22.11 SMOKE CONTROL IN ATRIA ANSI / NFPA 92A and 92B The NFPA f rst published ANSI / NFPA 92A, Recommended Practice for Smoke Control Systems, in 1987. It co vered smok e control barriers, airf ows, and the pressure dif ferential between dif ferent control zones. It also contained the guidelines to pro vide a smok eproof enclosure using stairwell pressurization, as well as smok e control for elevator shafts by using f re f oor exhaust, and elevator lobby pressurization. The latest NFPA 92A edition, in 1996, includes smoke control system supervision and instrumentation. Because the smoke control in large zones is quite complicated, another document, ANSI / NFPA 92B, Guide for Smoke Management Systems in Malls, Atria, and Large Areas, was f rst published in 1991. The latest edition is the 1995 edition.

Smoke Management in Atria, Malls, and Large Areas Klote (1997) and ANSI / NFPA 92B recommended the following smoke management in atria: ●

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Exhaust the smok e from the top of the atrium in order to achie ve a steady , clear height for a steady f re. Steady f re is an idealization of actual unsteady f re. Steady f re has a constant heat release rate. Consider that the only f ow into the smok e layer formed at the upper layer of a atrium is from the plume, and the only f ow from the smok e layer is the smok e exhaust. Plume comprises the comb ustion products abo ve the f re. The exhaust f ow must equal the f ow f rom the plume. This method and approach ha ve many simpli f ed assumptions, such as that the simple plume mass f ow equation is v alid, there is a constant heat release rate, clear height is greater than the mean f ame height, smoke layer is adiabatic, and plume f ow and exhaust are the only signi f cant mass f ows into or out from the smoke layer. Upper-layer exhaust is not necessary for an atrium with suf f cient smoke-f lling capacity. Occupants have time for both decision making and e vacuation before the smok e f lls the place where they are located. Various approaches to smok e f lling can be used to conserv atively estimate the f lling time. Air can be supplied to the communicating space to pro vide a speci f c a verage v elocity at the opening to the atrium to prevent smoke f owing from the atrium to the communicating space. When the above basic methods are not applicable, physical modeling and computing f uid dynamics (CFD) analysis can be used. Because a layer of hot air is formed under the ceiling of an atrium, and this layer of hot air prevents the smok e from reaching the ceiling, ceiling-mounted smok e detectors are usually not recommended for atrium applications. Beam smok e detectors mounted on balconies oriented horizontally are recommended to detect smoke in the plume. The exhaust system can pull some air belo w the smok e layer into the e xhaust inlet if the smok e layers are relatively thin. This plugholing of outdoor air can lower the smoke layer and expose occupants to smoke. Klote (1999) noted the maximum f ow Qsmoke of smoke without plugholing depends on the depth of the smok e layer and the temperature of the smok e. If the total smok e exhaust needed is greater than Qsmoke, a number of inlets to compensate e xhaust are needed. These inlets should be placed far enough from each other to avoid inf uence on their f ow.

AIR SYSTEMS: FAN COMBINATION AND SMOKE CONTROL

22.29

22.12 STAIRWELL PRESSURIZATION AND ZONE SMOKE CONTROL Stairwell Pressurization A stairwell pressurization system uses f ans to pressurize the stairwells to pro vide a smok efree escape route for the occupants in case of a b uilding f re. A stairwell pressurization system is a kind of smoke control system. National and local codes require stairwell pressurization systems in high-rise buildings. In a stairwell pressurization system, all interior stairwells are pressurized to a minimum of 0.15 in. WC (37 Pa) and a maximum of 0.35 in. WC (87 Pa) when all stairwell doors are closed. City of Ne w York Local Code 1979 requires a minimum air supply f ow rate of 24,000 cfm (11,326 L / s) plus 200 cfm (94 L / s) per f oor for the stairwell pressurization system. The maximum allowable pressure dif ference between the stairwell and the f oor space is 0.40 in. WC (100 P a) whether stairwell doors are opened or closed. The minimum allo wable pressure dif ference is 0.10 in. WC (25 Pa) when all stairwell doors are closed, or 0.05 in. WC (13 Pa) when any three stairwell doors are open. An alternative is to maintain at least 0.05 in. WC (13 Pa) or a minimum average air velocity of 400 fpm (2 m / s) at the stairwell door when an y three stairwell doors are opened. The force required to open a stairwell door must not exceed 25 lbf (111 N) at the doorknob. A stairwell pressurization system consists of centrifugal or v ane-axial fans, a stairwell pressurization supply duct with se veral supply air inlets, relief vents, and a control system, as shown in Fig. 22.7. Outdoor air is e xtracted directly by the centrifugal f an. It is forced into the supply duct and then supplied to the stairwell through supply inlets. When the stairwell is pressurized to a pressure typically 0.10 to 0.40 in. WC (25 to 100 P a) higher than that of the air outside the stairwell on various f oors across the stairwell doors, the smoke will not enter the stairwell, even an open stairwell door. Air supplied into the stairwell is dischar ged through the open stairwell doors, leakage area around closed stairwell doors, relief vents, or other openings to the rest of the b uilding, and then is discharged to the outdoors. If the stairwell is an airtight enclosure or its doors ha ve very small leakage areas, it will be overpressurized when all stairwell doors are closed. The pressure dif ference across the stairwell doors may be greater than 0.3 in. WC (75 P a). Often, a large force is required to turn the stairwell doorknob to open it. The total force required to open a stairwell door should not exceed 25 to 30 lbf (111 to 134 N), or it will be too diff cult to open stairwell doors during evacuation. Methods of overpressure relief are discussed later. The air velocity at the opened stairwell door on the f re f oor required to prevent the backf ow of smoke from the f re to the stairwell is called the critical velocity vcrit, in fpm (m / s). The outward f ow air v elocity from the stairwell through the open door on the f re f oor should be greater than vcrit.

Characteristics of Stairwell Pressurization Centrifugal or vane-axial fans can be used for stairwell pressurization. F ans can be installed either at the bottom le vel of the b uilding (bottom injection) or on the rooftop (top injection). Bottom injection is preferable because it minimizes the possibility of smok e inhalation and optimizes the stack effect to assist stairwell pressurization during winter . In an y case, the fan intake must be remote from the smok e exhaust during a b uilding f re. If the f an room for stairwell pressurization is located on the rooftop, facilities must be pro vided to minimize the in f uence of wind pressure on fan performance. Multiple injections, in which air is supplied from multiple inlets into the stairwell, provide a more even pressure distrib ution along the stairwell than a single injection from the top or bottom. Typically, each supply inlet serves two or three f oors. Open-tread stairs pro vide less f ow resistance than closed-tread stairs. This dif ference becomes more prominent when occupants are w alking on the stairs during e vacuation. Compartmentalization

22.30

CHAPTER TWENTY-TWO

FIGURE 22.7 Stairwell pressurization and zone smoke control systems: (a) fan bypass overpressure relief; (b) overpressure relief vents (barometric damper).

of the stairwell into many sections, such as serving four to f ve f oors, may not provide the airf ow rate necessary for stairwell pressurization when two or three stairwell doors are open at the same time.

Overpressure Relief and Feedback Control When a stairwell pressurization system is operating and all the stairwell doors are closed, the pressure difference across the stairwell doors e ventually exceeds the maximum permissible limit. Two methods are currently used to relie ve the stairwell pressure. Ov erpressure relief is achie ved by opening f rst-f oor e xit doors (as sho wn in Fig. 22.7 a) or o verpressure relief v ents (barometric dampers, as shown in Fig. 22.7 b). Variable volume of supply air is achie ved by means of feedback control (as shown in Fig. 22.7a). Exit Door Relief. When a stairwell pressurization system is turned on, the interlocked control system automatically opens the f rst-f oor e xit door to relie ve stairwell pressure. Ov erpressurized air in the stairwell is discharged to the outdoors. Because it is necessary to open the f rst-f oor exit door to evacuate the occupants during a building f re, this is a simple and effective means of overpressure relief. Test results in Tamura (1990b) showed that when stairwell pressurization w as activated and the f rst-f oor e xit door w as used as o verpressure relief, the maximum pressure dif ference across the

AIR SYSTEMS: FAN COMBINATION AND SMOKE CONTROL

22.31

stairwell doors w as lower than 0.25 in. WC (65 P a) in summer and winter . During non f re conditions, pressure differences were between 0.05 and 0.1 in. WC (13 and 25 P a) across the stairwell doors if there w as a v ent in the e xternal wall on the second f oor. When a b uilding f re was simulated on the second f oor, there w as no smok e back f ow from the f re f oor through the stairwell door when the stairwell doors on the f rst and second f oors were opened and the vent on the second f oor e xternal w all w as also opened. Smok e back f ow occurred when stairwell doors on the f rst, second, and third f oors were open. Overpressure Relief Vents. A relief v ent is typically an assembly in which a f re damper is connected in series with a barometric damper . The barometric damper is normally closed. If the stairwell is pressurized above a predetermined limit, the vents open and relieve air to various building f oors. A counterweight in the barometric damper sets the maximum pressure limit. Fire dampers are normally closed, and open when the stairwell pressurization system is turned on in case of a building f re. According to the e xperimental results of Tamura (1990b), the performance of the o verpressure relief vents at a supply f ow rate of 28,000 cfm (13,210 L / s) was slightly better than the performance of a ground- f oor exit door at a supply f ow rate of 17,800 cfm (8400 L / s). If supply f ow rates are the same, the difference is further reduced. Fan Bypass. When the stairwell is o verpressurized, a pressure sensor located inside the stairwell signals a feedback control to open a f an bypass damper so that part of the supply air returns to the centrifugal f an inlet, as shown in Fig. 22.7 a. The air v olume f ow supplied to the stairwell is reduced until the pressure at the pressure sensor drops belo w a preset v alue. If the pressure sensor is located on the rooftop or outdoors, it should be shielded from the inf uence of wind. Variable-Speed or Controllable-Pitc h Fan. When excessive pressure is detected in the stairwell, a controller actuates an adjustable-frequenc y variable-speed drive to reduce the speed of a centrifugal fan or to v ary the blade pitch of a v ane-axial fan to maintain the required pressure inside the stairwell at the pressure sensor. The test results in Tamura (1990b) showed that both fan bypass and variable-speed controls require a response time of more than 5 min. Variable-speed control is slightly faster. Because reliability is the primary f actor in stairwell pressurization control, and because of its very short operating period, it may not be worthwhile to install an expensive adjustablefrequency ac inverter to provide variable-speed control.

Stair and Shaft Vents ASHRAE / IESNA Standard 90.1-1999 mandatorily specif es that stair and elevator shaft vents shall be equipped with motorized dampers which are automatically closed during normal b uilding operation and are interlocked to open by f re and smoke control detection systems when required.

Zone Smoke Control Zone smoke control requires that a b uilding be divided into a number of smok e control zones. The control zones are separated from one another by barriers, such as walls, f oors, ceilings, and doors. In high-rise buildings, each f oor is a separate zone, or it may be subdi vided into many smoke control zones, or a smoke control zone consists of more than one f oor. In case of a b uilding f re, the spread of smok e from the zone of f re origin (the smok e zone) to adjacent zones is limited by pressure dif ferences and airf ows. The required pressure dif ferential depends on the pressure produced by the f re gases, i.e., the temperature of the b uilding f re. To form a pressure dif ference between the smok e zone and adjacent control zones, zone smok e control provides smok e e xhaust on the f re f oor (smok e zone) by opening the smok e damper connected to the smok e e xhaust duct. At the same time, it supplies outdoor air to the f oor or f oors above and below the f re f oor (to control zones adjacent to the smok e zone) and pressurizes

22.32

CHAPTER TWENTY-TWO

them to pre vent smok e contamination by operating the air -handling units (AHUs) on these f oors and closing the smoke dampers connected to the smoke exhaust duct, as shown in Fig. 22.7a. During the smok e control mode, recirculation dampers in these AHUs are fully closed. Outdoor air should be e xtracted directly through the coils, supply f ans, and supply to the duct dif fusers to pressurize these control zones These mechanisms should be controlled by a DDC unit for f re protection management overriding all other HVAC&R controls during a building f re. Lougheed (1997) noted that the 0.05 in. WC (12.5 P a) minimum design pressure dif ference between the smoke zone and adjacent zones suggested by ANSI / NFPA 92A (1993) w ould be adequate for zone smok e control applications in sprinklered of f ce b uildings. In most cases, a lower pressure of approximately 0.028 in. WG (7 Pa) would still exceed the pressure produced by the f re gases. Lougheed (1997) recommended a minimum required pressure dif ference of 0.1 in. WC (25 Pa) for an unsprinklered building with a ceiling height of 9 ft (2.74 m). The supply and e xhaust volume f ow rates for zone smok e control are often matched with the HVAC&R system in b uildings, especially when an air economizer c ycle is used. The exhaust air from the AHU can be connected to a smok e exhaust duct, and the e xhaust fan is generally located on the rooftop. In an air system operated only in recirculation mode, a smoke exhaust system for zone smok e control should be installed. In such circumstances, the smoke exhaust volume f ow rate can be determined to equal about 6 air changes per hour (ach).

Design Considerations The object of smoke control is to provide a smokefree escape route for occupants through the stairwell to the outdoors during a b uilding f re. This is a part of the b uilding f re protection scheme. The primary considerations are safety and reliability . Because the performance of stairwell pressurization and the zone smok e control are related, they should be considered an inte grated smoke control system during system design. Operation of both stairwell pressurization and zone smok e control systems must meet the requirements of national and local codes and pro vide a smokefree escape route e ven under the following conditions: ●

●

●

●

A f re reaches a f re temperature of up to 1200 °F (650 °C) when the sprinkler system f ails to operate. Three or four stairwell doors, including the door on the f re f oor, are opened simultaneously. Fire f oor smoke exhaust is performed mainly by a zone mechanical smoke exhaust system. During summer operating conditions, the stack effect does not act as an additional assistant.

Recently, it has been recommended that smok e control technology be e xtended to ele vator shafts and lobbies used to e vacuate disabled persons during a b uilding f re. An elevator shaft and ele vator lobby pressurization system should then be installed. During a b uilding f re, HVAC&R should operate according to f re safety and smok e control requirements. Smok e control systems are automatically actuated by a w ater f ow indicator from the automatic sprinkler system or from an area smok e detector. It is important to ha ve feedback mechanisms to verify system operation and performance, such as separate alarms to signal smok e migration or b urning f re, and adequate pressure le vel at k ey points during pressurization and evacuation.

Volume Flow Rate Volume f ow rate is the primary determinant of the performance of stairwell pressurization and zone smoke control systems. The volume f ow rate of a zone smok e exhaust system from the f re f oor is

AIR SYSTEMS: FAN COMBINATION AND SMOKE CONTROL

22.33

usually 6 ach. This volume f ow rate is used by many designers and has been verif ed as appropriate in f eld tests by Tamura (1991) and Klote (1990, Part II). The volume f ow rates of air supplied to the f oors immediately abo ve and belo w normally should be the same as the supply v olume f ow rate of the AHU or packaged unit for that f oor. Air discharged through the open stairwell doors and leak ed through the closed stairwell doors and walls can be summarized into the following types: ●

●

●

●

Air discharged through an open stairwell door on the f re f oor with a critical velocity vcrit Air discharged through the open f rst-f oor exit door Air discharged through open stairwell doors on f oors other than the f re f oor and the f rst f oor Air leaked through cracks around the stairwell door and in the stairwell wall

The volume f ow rate supplied to a stairwell pressurization system then can be calculated as follows: V˙s,p  (vcrit  vexit)Adoor  V˙o,dNo,d  V˙leakNc,d

(22.11)

where vexit  average air velocity at f rst-f oor exit door, fpm (m / s) Adoor  area of opened stairwell door on f re f oor, ft2 (m2) V˙o,d  volume f ow rate discharged through open stairwell door on f oors other than f re f oor and f rst f oor, cfm (m3 /s) No,d  number of open doors in stairwell other than f re f oor and f rst f oor V˙leak  air leakage through cracks across stairwell wall on f oors with closed stairwell doors, cfm (m3 /s) Nc,d  number of closed doors in the stairwell According to the experimental results in Tamura (1991), for a f re in a building with stairwell pressurization and a smoke exhaust system to exhaust smoke from the f re f oor, the critical velocity vcrit can be tak en as 300 fpm (1.5 m / s) for a f re temperature of 1200 °F (650°C) and a mechanical e xhaust from the f re f oor of 5.5 ach. In stairwell pressurized systems serving up to 10 f oors, No,d  3, including the door on the f re f oor. In systems serving 15 or more f oors, No,d  4, including the door on the f re f oor. For safety, air velocity at the f rst-f oor stairwell exit door vexit can be assumed to be equal to vcrit. Discharge air velocity at open stairwell doors on f oors other than the f re f oor vo,d is less than vcrit, in fpm (m / s). This is because the f ow resistance of the cracks around doors and windo ws in f oors other than the f re f oor is greater than that on the f re f oor. The air v olume f ow rate dischar ged through an open stairwell door other than that on the f re f oor can be calculated from Eq. (22.10) by using the ef fective area Ae of the airf ow path instead of the area of the door or opening A. Air leakage through cracks in the stairwell wall can also be calculated from Eq. (22.10).

System Pressure Loss for Stairwell Pressurization System The frictional loss of air f ow per f oor inside the stairwell pressurization system pf,s, in in. WC (Pa), can be calculated by considering the stairwell as a rectangular duct, as follows: pf,s  K

a

˙

1  5.19  DL  2g  60AV  e

where L  vertical height of each f oor, ft (m) gc  dimensional constant, 32.2 lbm ft / lbf s2

c

2

o

(22.12)

22.34

CHAPTER TWENTY-TWO

a  air density, lb / ft3 (kg / m3) V˙  air supply volume f ow rate, cfm (m3 /s) Ao  cross-sectional area of free air passage in stairwell, ft2 (m2) and the ratio R of the area of free f ow passage (ori f ce) to the interior cross-sectional area of the stairwell As, in ft2 (m2), can be calculated as R  Ao /As

(22.13)

In Eq. (22.12), De, in ft (m), indicates the circular equivalent of the stairwell, and it can be calculated as De 

4As Ps

(22.14)

where Ps  perimeter of the cross-sectional area of the stairwell, ft (m). The pressure drop coef f cient is K. It is mainly a function of the con f guration of the stairwell and the occupant density on the stairs. According to the experimental results of Achakji and Tamura (1988), for a stairwell tested under the following conditions ●

●

●

With a cross-sectional area of 134 ft2 (12.5 m2) A f oor height of 8.5 ft (2.6 m) An occupant density of 0.18 person / ft2 (2.0 persons / m2)

the frictional loss of airf ow per f oor pf,s has the following values: vo, fpm (m / s) Open-tread stair, bottom injection Closed-tread stair, bottom injection

879 659 439 1055 848 527

(4.47) (3.35) (2.23) (5.36) (4.31) (2.68)

pf,s, in. WC (Pa) 0.128 0.070 0.032 0.200 0.116 0.05

(32) (17.5) (8) (50) (29) (12.5)

Here vo represents the air v elocity calculated based on the free f ow area, in fpm (m / s). The difference in pf,s between top and bottom injection is small if other conditions remain the same. When calculating the system pressure loss of a stairwell pressurization system with multiple injection psy, in in. WC (Pa), as shown in Fig. 22.6, is calculated as psy  pa  b  pb  o

(22.15)

where pa  b, pb  o  pressure loss between points a and b and points b and o, respectively, in. WC (Pa) Pressure loss pb  o usually varies between 0.10 and 0.40 in. WC (25 and 100 Pa) and is only a small portion of the system pressure. Ho wever, it is diff cult to calculate accurately because of complicated air f ow paths at v arious operating conditions and the in f uence of outdoor conditions. Pressure loss pa  b should be calculated between points a and b according to the procedure described in Chap. 17, including the velocity pressure at the supply inlet, and pb  o should be estimated as 0.4 in. WC (100 P a). The calculated system pressure psy should be multiplied by a safety factor 1.2. The actual pb  o can be adjusted during acceptance testing. In a stairwell pressurization system with a bottom single injection, as shown in Fig. 22.8 a, the maximum pressure dif ference between the stairwell and outdoors often occurs at the top of the stairwell pc  o because the cross section of the stairwell is constant. The system pressure can therefore be calculated as psy  pa  b  pb  c  pc  o

(22.16)

AIR SYSTEMS: FAN COMBINATION AND SMOKE CONTROL

FIGURE 22.8 Bottom injection, and bottom and top injection: injection.

22.35

(a) bottom injection; ( b) bottom and top

Here, for simplicity, pc  o again can be taken as 0.4 in. WC (100 Pa). Then pb  c  pf,sNf

(22.17)

where Nf  number of f oors. Actually, part of the velocity pressure discharged from the bottom inlet may be con verted to static pressure, so that pb  c is smaller. However, more data are needed before it can be calculated accurately. In a stairwell pressurization system with top and bottom injection, as shown in Fig. 22.8 b, the system pressure loss can be calculated as it is in bottom injection, but pb  c may be smaller. Computer-aided design software for zone smoke control systems are available to assist designers. Example 22.1. In a smoke control system serving a 20-story high-rise building, each f oor has the following construction characteristics: Floor area Area of the external wall Vertical distance between f oor and f oor Area of the stairwell wall Elevator shaft (three elevators) wall area Area of stairwell door Volume f ow rate of supply air to conditioned space through AHU on each f oor

10,000 ft2 (929 m2) 4800 ft2 (446 m2) 10 ft (3m) 600 ft2 (56 m2) 1000 ft2 (93 m2) 3  7  21 ft2 (2 m2) 12,000 cfm (5663 L / s)

22.36

CHAPTER TWENTY-TWO

The pressure difference across the closed stairwell door is 0.1 in. WC (25 Pa). The pressure differences between the stairwell and the outdoors on f oors other than the f re f oor whose stairwell doors remain open are also 0.1 in. WC (25 Pa), because the pressure drop of the open stairwell door is ignored. Determine the v olume f ow rates of the zone smok e control and stairwell pressurization systems. Solution 1. The volume f ow rate of the zone smok e exhaust system from the f re f oor V˙ex at a rate of 6 ach is 10 V˙ex  6  10,000   10,000 cfm (4719 L / s) 60 The v olume f ow rate of air supplied to the f oors immediately abo ve and belo w the f re f oor is 12,000 cfm (5663 L / s). 2. The volume f ow rate of air that is dischar ged through the open stairwell door to the f re f oor V˙fire, in cfm (L / s), and is exhausted through the zone smoke control system is V˙fire  vcrit Adoor  300  3  7  6300 cfm (2973 L/s) 3. The volume f ow rate dischar ged from the f rst-f oor exit door is assumed to be the same as that from the open stairwell door on the f re f oor (6300 cfm, or 2973 L / s). 4. From the given data, the leakage area on the stairwell w all, including cracks around the stairwell door, is Aleak  0.25  0.00011  600  0.32 ft2 If the pressure dif ference across the closed stairwell door is 0.1 in. WC, from Eq. (22.10), the air leakage rate V˙ leak through the cracks of the stairwell w all on each f oor whose stairwell door is closed can be calculated: V˙leak  2610 A(p)1/2  2610  0.32(0.1)1 / 2  264 cfm (125 L / s) 5. It is possible that the f re f oor and the pressurized f oors immediately above and below are all in the lower 10 f oors. Because of the stack ef fect in this 20-story b uilding, air mixed with hot gas from the burning f re is discharged from the upper 10 stories. In a 20-story b uilding, assume that there are four open stairwell doors. One is on the f re f oor, and another is the f rst-f oor exit door. Then the remaining tw o opened stairwell doors should each discharge a combined air path that may be a combination of six parallel paths, as sho wn in Fig. 22.9. a-b-o a-c-d-o a-c-e-f-o a-g-h-o a-g-i-j-o a-k-l-o 6. The leakage area for the external wall for each f oor is Aleak  4800  0.00021  1.0 ft2 The leakage area for a f oor area of 10,000 ft2 is Aleak  10,000  0.000052  0.52 ft2

AIR SYSTEMS: FAN COMBINATION AND SMOKE CONTROL

22.37

FIGURE 22.9 Air dischar ged from an open stairwell door through a combined air path to the outdoors.

The leakage area to the elevator shaft is Aleak  3  0.25  1000  0.00084  1.59 ft2 If the total opening at the top of the ele vator shaft is 5 ft 2, assume that only 0.5 ft 2 will be allocated for the open stairwell door. For an airf ow path a-b-o, if the pressure drop at the open door is ne gligible, the effective area is Ae 

 A1 

1 / 2

 1 ft 2

2 1

For airf ow path a-c-d-o, the effective area is Ae 

 A1

2 1



1 A22

1 / 2





 0.521

2

1 12



1 / 2



 0.46 ft 2

For airf ow path a-c-e-f-o, the effective area is Ae  

 A1

2 1





1 1  2 A22 A3

1 / 2



1 1 1   2 0.522 0.522 1

1 / 2



 0.35 ft 2

For airf ow path a-k-l-o, the effective area is Ae 

 1.591

2



1 0.52

1 / 2



 0.48 ft 2

The effective area of air f ow path a-g-h-o is the same as that of a-c-d-o, that is, 0.46 ft 2; and the effective area for a-g-i-j-o is the same as that of a-c-e-f-o, that is, 0.35 ft 2. Therefore, the total

22.38

CHAPTER TWENTY-TWO

effective area of this combined airf ow path is Ae  2  1  2  0.46  2  0.35  0.48  4.1 ft2 (0.38 m2) 7. From Eq. (22.10), air discharged from this opened stairwell door is V˙  2610A(p)1 / 2  2610  4.1(0.10)1 / 2  3384 cfm (1597 L / s) 8. From Eq. (22.11), the total supply v olume f ow rate of this stairwell pressurization system is then V˙s,p  (vcrit  vexit)Adoor  V˙o,dNo,d  V˙leakNc,d  (300  300)3  7  2  3384  264(20  4)  23,592 cfm (11,133 L / s)

REFERENCES Achakji, G. A., and Tamura, G. T., Pressure Drop Characteristics of Typical Stairshafts in High-Rise Buildings, ASHRAE Transactions, 1988, Part I, pp. 1223 – 1237. Alcorn, L. H., and Huber, P. J., Decoupling Supply and Return Fans for Increased Stability of VAV Systems, ASHRAE Transactions, 1988, Part I, pp. 1484 – 1492. Alley, R. L., Selecting and Sizing Outside and Return Air Dampers for VAV Economizer Systems, ASHRAE Transactions, 1988, Part I, pp. 1457 – 1466. ASHRAE, ASHRAE Handbook 1992, HVAC Systems and Equipment, ASHRAE Inc., Atlanta, GA, 1992. ASHRAE, ASHRAE Handbook 1997, Fundamentals, Atlanta, GA, 1997. ASHRAE, ASHRAE Handbook 1999, HVAC Applications, Atlanta, GA, 1999. Avery, G., VAV Economizer Cycle: Don’t Use a Return Fan, Heating / Piping / Air Conditioning, no. 8, 1984, pp. 91 – 94. Clark, D. R., Hurley, C. W., and Hill, C. R., Dynamic Models for HVAC System Components, ASHRAE Transactions, 1985, Part I, pp. 737 – 751. Clark, J. A., and Harris, J. W., Stairwell Pressurization in a Cold Climate, ASHRAE Transactions, 1989, Part I, pp. 847 – 851. Grimm, N. R., and Rosaler, R. C., HVAC Systems and Components Handbook, 2d ed., McGraw-Hill, New York, 1998. Kalasinsky, C. C., The Economics of Relief Fans vs. Return Fans in Variable Volume Systems with Economizer Cycles, ASHRAE Transactions, 1988, Part I, pp. 1467 – 1476. Kettler, J. P., Field Problems Associated with Return Fans on VAV Systems, ASHRAE Transactions, 1988, Part I, pp. 1477 – 1483. Klote, J. H., An Overview of Smoke Control Technology, ASHRAE Transactions, 1988, Part I, pp. 1211 – 1222. Klote, J. H., Fire Experiments of Zoned Smoke Control at the Plaza Hotel in Washington, D.C., ASHRAE Transactions, 1990, Part II, pp. 399 – 416. Klote, J. H., Prediction of Smoke Movement in Atria: Part I — Physical Concepts and Part II — Application to Smoke Management, ASHRAE Transactions, 1997, Part II, pp. 534 – 553. Klote, J. H., What’s New in Atrium Smoke Management, HPAC, no. 4, 1999, pp. 28 – 31. Klote, J. H., and Tamura, G. T., Design of Elevator Control Systems for Fire Evacuation, ASHRAE Transactions, 1991, Part II, pp. 634 – 642. Kukla, M. D., Situations to Consider When Variable Air Volume Is an Option, ASHRAE Transactions, 1997, Part II, pp. 823 – 829. Lehr, V. A., Life Safety in Tall Buildings, Heating / Piping / Air Conditioning, April 1990, pp. 41 – 46. Lougheed, G. D., Expected Size of Shielded Fires in Sprinklered Off ce Buildings, ASHRAE Transactions, 1997, Part I, pp. 395 – 410.

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22.39

Mawhinney, J. R., and Tamura, G. T., Effect of Automatic Sprinkler Protection on Smoke Control Systems, ASHRAE Transactions, 1994, Part I, pp. 494 – 513. McGreal, M. P., Engineered Smoke Control Systems, HPAC, no. 4, 1997, pp. 69 – 72. NFPA, Guide for Smoke and Heat Venting, ANSI / NFPA Standard, 204, Quincy, MA, 1988. Peacock, R. D., Bukowski, R. W., Jones, W. W., and Forney, C. L., Technical Reference Guide for HAZARD I Fire Hazard Assessment Method, NIST Handbook, National Institute of Standards and Technology, Gaithersburg, MD, 1989, chap. 7, p. 146. Schwartz, K. J., Jensen, R. H., and Antell, J., The Role of Dampers in a Total Fire Protection System Analysis, ASHRAE Transactions, 1986, Part I B, pp. 566 – 576. Shavit, G., Information-Based Smoke Control Systems, ASHRAE Transactions, 1988, Part I, pp. 1238 – 1252. Tamura, G. T., Stair Pressurization Systems for Smoke Control: Design Considerations, ASHRAE Transactions, 1989, Part II, pp. 184 – 192. Tamura, G. T., Field Tests of Stair Pressurization Systems with Overpressure Relief, ASHRAE Transactions, 1990a, Part I, pp. 951 – 958. Tamura, G. T., Fire Tower Tests of Stair Pressurization Systems with Overpressure Relief, ASHRAE Transactions, 1990b, Part II, pp. 373 – 383. Tamura, G. T., Fire Tower Tests of Stair Pressurization Systems with Mechanical Venting of the Fire Floor, ASHRAE Transactions, 1990c, Part II, pp. 384 – 392. Tamura, G. T., Determination of Critical Air Velocities to Prevent Smoke Backf ow at a Stair Door Opening on the Fire Floor, ASHRAE Transactions, 1991, Part II, pp. 627 – 633. The Trane Company, Air Conditioning Fans, The Trane Company, La Crosse, WI, 1985. Wang, S. K., Air Conditioning, vols. 2 and 4, Hong Kong Polytechnic, Hong Kong, 1987.

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AIR SYSTEMS: MINIMUM VENTILATION AND VAV SYSTEM CONTROLS 23.1 MINIMUM VENTILATION CONTROL 23.2 Basic Approach 23.2 Specific Controls in VAV Systems 23.2 Types of Ventilation Control 23.2 ASHRAE Standard 62-1999 23.3 Recirculation of Unused Outdoor Air in Multizone Systems 23.4 Ventilation Controls for High-Occupancy Areas 23.5 23.2 CO2-BASED DEMAND-CONTROLLED VENTILATION 23.5 CO2-Based Demand-Controlled Ventilation or Time-Based Constant-Volume Control 23.5 CO2 Sensor or Mixed-Gases Sensor 23.7 Location of CO2 Sensor 23.7 Substantial Lag Time in Space CO2 Concentration Dilution Process 23.8 Base Ventilation and Purge Mode 23.9 A CO2-Based Demand-Controlled Ventilation System 23.10 Application of CO2-Based DemandControlled Ventilation System 23.11 23.3 MIXED PLENUM PRESSURE CONTROLS 23.12 Basics 23.12 Case Study: Mixed Plenum Pressure Control Monitoring Plenum Pressure 23.12 Monitoring Pressure Drop of Louver and Damper Controlling Mixed Plenum Pressure 23.13 Applications 23.14 23.4 OUTDOOR AIR INJECTION FAN, DIRECT MEASUREMENT, AND FAN TRACKING SYSTEMS 23.14 Outdoor Air Injection Fan 23.14 Direct Measurement of Minimum Outdoor Air Intake 23.15 Fan Tracking Systems 23.15 23.5 CONFERENCE ROOMS 23.16 23.6 SPACE PRESSURIZATION AND RETURN VOLUME FLOW CONTROLS 23.16 Characteristics of Space Pressure Control 23.16

VAV Systems Return/Relief Fan Volume Flow Control 23.17 23.7 DISCHARGE AIR TEMPERATURE CONTROLS 23.18 Basics 23.18 System Description 23.19 Operation of Air Economizer and Outdoor Air Intake — Case Study 23.21 Discharge Air Temperature Reset 23.22 23.8 DUCT STATIC PRESSURE AND FAN CONTROLS 23.23 Duct Static Pressure Control 23.23 Set Point of Duct Static Pressure and Sensor’s Location 23.24 Comparison between Adjustable-Frequency Drives and Inlet Vanes 23.24 23.9 FUNCTIONAL CONTROLS FOR VAV SYSTEMS 23.26 Nighttime Setback, Warm-Up, and Cool-Down Control 23.26 Steam Humidifier Control 23.27 Dew Point Control 23.27 Diagnostics 23.28 23.10 RECOMMENDATIONS AND INTERACTION BETWEEN CONTROLS 23.28 Recommendations for VAV Controls 23.28 Interaction between Controls 23.29 Stability Problems 23.30 23.11 SEQUENCE OF OPERATIONS OF A VAV REHEAT SYSTEM WITH MINIMUM VENTILATION, DISCHARGE AIR TEMPERATURE, AND DUCT STATIC PRESSURE CONTROLS — CASE STUDY 23.30 HVAC&R System 23.30 Sequence of Operations 23.30 Primary Considerations 23.34 REFERENCES 23.35

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23.1 MINIMUM VENTILATION CONTROL Basic Approach As discussed in Sec. 4.10, indoor air quality (IAQ) is defined as an indication of harmful concentra tion of indoor air contaminants, such as particulates, combustion products, volatile or ganic compounds, nicotine, radon, and biological compounds that affects the health of the occupants or the degree of satisfaction of a substantial majority of occupants e xposed to such an indoor en vironment. Contaminated source control, removal by air cleaners, and using outdoor ventilation air to dilute the concentrations of indoor air contaminants are three basic strate gies to improve IAQ. Emissions and odors from the human body , volatile organic compounds, and many gaseous contaminants of 0.003- to 1- m minute particles can only be remo ved by e xpensive, high-efficien y air filters an activated carbon filters. Use of adequate outdoor air to dilute the air contaminants has been pr ved an essential, practical, and cost-effective means to improve IAQ. Among the air systems, ventilation controls for constant-v olume systems and dedicated v entilation systems can be achie ved either by fixing the outdoor and recirculating damper position durin the operating period or modulating the f an speed to match the variation of outdoor air supply. Also, during the air economizer c ycle, outdoor air intak e in the AHU or PU is al ways greater than the specified amount of outdoor air for entilation required. Ho wever, the reduction in supply v olume fl w rate at part-load operation of v ariable-air-volume (VAV) systems, a tighter building shell, and the inef fective v entilation control may reduce the amount of outdoor air intak e to less than the requirement for ventilation. Minimum ventilation control (control in minimum outdoor air recirculating mode) becomes one of the critical problems in VAV systems in commercial b uildings. As discussed in Sec. 4.10, ventilation means supplying and removing ventilation air, and ventilation air consists of outdoor air plus treated recirculated air for acceptable indoor air quality. Specific Controls in VAV Systems In a multizone VAV system, zone temperature control directly affects the thermal comfort of the occupants and is the basis of the VAV system control. During the minimum v entilation control, only the amount of minimum outdoor air supply to a conditioned space is controlled; the space airfl w balance and space pressurization v ary accordingly. There are interactions between v entilation and space pressurization controls. Zone temperature control is achie ved by modulating the v olume fl w rate supplied to each control zone. As the v ariation of zone load causes the change of damper positions and the airfl w passages as well as the zone v olume fl w in VAV boxes, the duct static pressure in the main supply duct may rise or f all. Duct static pressure control then modulates the supply f an speed and thus the system supply volume fl w rate and fan total pressure to match the variation in zone loads. Discharge air temperature control adjusts the cooling or heating capacity of a VAV system to match the system load. It also resets the discharge air temperature and optimizes energy use. Therefore, specific controls for a VAV system include ●

●

●

●

SH__ ST__ LG__ DF

Zone temperature controls Minimum ventilation control and space pressurization control Discharge air temperature control, including economizer, cooling capacity, mixed air, and heating capacity controls in sequence, and low-temperature limit control Duct static pressure control

Types of Ventilation Control In VAV systems, ventilation control includes air economizer control and minimum outdoor v entilation air control during the recirculation mode. During the air economizer c ycle including mixed air

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control, a volume fl w of outdoor air that exceeds the amount of minimum outdoor ventilation air is always supplied to the conditioned space. Ventilation problems often occur during minimum outdoor air recirculating mode, i.e., minimum ventilation controls. Currently used minimum v entilation controls in VAV systems include the following: CO2-based demand-controlled ventilation Mixed plenum pressure controls Outdoor air injection fan control Fan tracking control

ASHRAE Standard 62-1999 As discussed in Sec. 4.10, ASHRAE Standard 62-1999, Ventilation for Acceptable Indoor Air Quality, specifies t o alternative procedures to obtain acceptable IAQ: 1. Ventilation rate procedure. In the v entilation rate procedure, acceptable indoor air quality is achieved by pro viding v entilation air of specified quality and quantity to the conditioned space This procedure prescribes the following: ●

●

●

●

●

●

●

The quality of the outdoor air shall be acceptable for v entilation, i.e., contaminants in outdoor air do not e xceed the concentrations listed in the Table of National Primary Ambient-Air Quality Standards by EPA in Sec. 4.10. Outdoor air treatments are required when necessary. Ventilation rates for typical spaces are listed in Table 4.5. For details, refer to ASHRAE Standard 62-1999. Ventilation systems for spaces with intermittent or v ariable occupancy may have their outdoor air quantity adjusted by damper or c ycling of fan operation to pro vide sufficient dilution to maintai contaminant concentrations within acceptable levels at all times. For intermittent occupanc y, when contaminants are associated only with occupants or occupant activities, do not result a health hazard because contaminants are dissipated during unoccupied periods. When contaminants are independent of occupants or occupant aci vities, the supply of outdoor air should lead occupancy lag or lead time depending on the v entilation rate and air capacity per person in the space. Where peak occupancies are less than 3 h in duration, the outdoor airfl w rate may be determined according to average occupancy for the duration of operation of the system pro vided the average occupancy used is not less than half of the maximum.

2. Indoor air quality pr ocedure. In the indoor air quality procedure, acceptable air quality is achieved within the conditioned space by restricting the concentration of all known contaminants of concern to some specified acceptable l vels. The acceptable concentration level in Table of National Primary Ambient-Air Quality Standards, as listed in Sec. 4.10, also applies indoors for the same exposure time. The following limits apply for four other indoor contaminants according to Standard 62-1999: Contaminant

Concentration

ppm

Exposure time

Carbon dioxide CO2 Chlordane Ozone Radon

1.8 g / m 5 g / m3 100 g / m3 4 pCi / L

1000 0.0003 0.05

Continuous Continuous Continuous Annual average

3

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CHAPTER TWENTY-THREE

The new Standard 62-1999 contains the entire 1989 v ersion, unchanged, along with the follo wing four new addenda: ●

●

●

●

Addenda 62e remo ves the statement that v entilation rates in Table 2 accommodate a moderate amount of smoking because of the dangerousness of second-hand tobacco smoke. Addenda 62c remo ves the requirement that the temperature and humidity conditions specified i Standard 55 be maintained in all ventilated space, such as even in garages. Addenda 62d re words the standard scope to state that compliance with the standard will not ensure acceptable indoor air quality. Addenda 62f clarifies that carbon dioxide (C 2) is simply a useful indicator of the concentration of human bioeffluents

For details, refer to ASHRAE Standard 62-1999. Standard 62-1999 is under continuous maintenance by a Standing Standard Project Committee for which the Standards Committee has established a documented program for regular publication of addenda or revisions.

Recirculation of Unused Outdoor Air in Multizone Systems For a multizone VAV system, the ratio of outdoor air to supply air required to meet the v entilation and thermal load requirements is often dif ferent from zone to zone. The system v entilation air required, in most cases, is actually the amount of outdoor air intak e at the AHU or PU. The problem is how the system-fi ed ratio of outdoor air to supply air can satisfy the dif ferent ratios of outdoor air to supply air in various control zones. In a specific control zone if the ratio of outdoor air to supply air is higher than is needed in this zone, the result is an increase in unused outdoor air in the recirculating air (more truly , the concentration of air contaminants is lo wer) which raises the dilution po wer of the supply air (including outdoor and recirculating air) for other control zones. Methods of calculating ventilation requirements using the unused outdoor air contained in recirculating air were de veloped in Australia and are included in ASHRAE Standard 62-1999. The required corrected fraction of outdoor air supply in system supply air Y, in cfm (L / s), considering the outdoor requirement of the critical zone and the unused outdoor air contained in the recirculating air can be calculated as Y

X 1XZ

(23.1)

and Y  V˙o,cor / V˙s X  V˙o,sys / V˙s

(23.2)

Z  V˙o,cr / V˙s,cr

SH__ ST__ LG__ DF

where V˙s  supply volume fl w rate of air system, cfm (L / s) V˙o,cor  corrected outdoor air supply volume fl w rate, considering critical zone and unused outdoor air in the recirculating air, cfm (L / s) V˙ o,sys  calculated system outdoor air volume fl w rate, cfm (L / s) V˙ o,cr  outdoor air volume fl w rate required in critical zone, cfm (L / s) V˙ s,cr  critical zone supply volume fl w rate, cfm (L / s) X  uncorrected fraction of outdoor air supply Y  corrected fraction of outdoor air supply Z  fraction of outdoor air in supply air in critical zone

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The system outdoor air volume f ow rate can be evaluated as V˙o,sys  NocV˙o,req

(23.3)

where Noc  number of occupants in conditioned area served by air system V˙ o,req  ventilation air requirements as specif ed by ASHRAE Standard 62-1999, cfm / person (L / s person) Outdoor air v olume f ow rates V˙ o,cor, V˙ o,sys, and V˙ o,cr and supply v olume f ow rates V˙s, and V˙ s,cr all must occur at the same time. Ventilation Controls for High-Occupancy Areas ASHRAE / IESNA Standard 90.1-1999 mandates that air systems with design outdoor v entilation air intake greater than 3000 cfm (1420 L / s) serving areas ha ving a design occupanc y density e xceeding 100 people per 1000 ft 2 (93 m 2) shall include means which automatically reduce outdoor air intake below design v olume f ow rate when spaces are partially occupied. Outdoor v entilation air control shall be in compliance with ASHRAE Standard 62 and local codes e xcept for systems with heat recovery.

23.2 CO2-BASED DEMAND-CONTROLLED VENTILATION CO2-Based Demand-Controlled Ventilation or Time-Based Constant-Volume Control As discussed in Sec. 4.10, among the indoor contaminants, particulates should be remo ved by medium- and high-eff ciency f lters; vented combustion products including CO do not normally e xceed the allowable indoor concentration le vels in a conditioned space; v olatile organic compounds including formaldehyde may have above-allowable indoor concentration levels when the building is new or after remodeling; smoking is prohibited in man y public places and is often limited in a specif c area in many buildings; only 6 percent of U.S. homes in their annual a verage radon concentration exceeds the EPA specif ed level and needs subslab depression system and others; and airborne virus may become dangerous when any one of the occupants is sick and carries such a virus. Under many circumstances, occupant-generated CO 2, as a useful indicator of the concentration of body bioeff uents, emissions, and related contaminants, may be used as one of the indicators of poor indoor air quality. Many buildings, such as auditoriums, assembly halls, airport terminals, retail stores, off ces, and classrooms, have a v ariable occupancy during w orking hours and thus a f uctuation of the amount of indoor contaminants released. Figure 23.1 sho ws the CO 2 concentration prof les measured in the f oor space and in the return air duct on the 18th f oor with a 20 percent f xed outdoor air supply of a 22-story building in Ottawa, Canada. On a weekday, the concentration of CO2 gradually rose from 400 ppm in the morning to about 800 ppm before the of f ce w orkers went out for lunch. After lunch, the concentration of CO 2 f uctuated and reached a daily maximum at about 3 p.m. Then the CO2 concentration dropped gradually to 450 ppm at 8 p.m. If the amount of v entilation air required to dilute the concentration le vel of the indoor contaminants v aried accordingly , energy can be saved. More than the required amount of outdoor v entilation air intak e means more than the required amount of energy is used to cool and dehumidify the outdoor air during summer and to heat it during winter. Demand-controlled ventilation (DCV) means that when the number of occupants (occupant density) in the conditioned space during the occupied period, and thus the demand, drops, less outdoor air is taken into the AHU or PU and supplied to the conditioned space to meet the demand of dilution of concentration le vels of contaminants with nearly the same speci f ed v entilation rate

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CHAPTER

TWENTY- THREE

cfm/person (L/s-person). Time-based constant-volume ventilation control intends to supply constant-volume conditioned outdoor air to the conditioned space for dilution during the occupied period even if the occupant density drops. Compared with the time-based constant-volume ventilation control, demand-controlled ventilation has the following advantages: .DCV supplies the demanded amount of outdoor air to the conditioned space to maintain an acceptable air quality and, at the same time, saves heating and cooling energy to condition the excessive amount of supplied outdoor air in the time-based constant-volume ventilation control. According to the simulation results in Carpenter (1996), compared to a system without DCV, l000-ppm CO2-based DCV reduces 25 to 30 percent of space heating energy use. With 1000-ppm CO2-based demand-controlled ventilation, the average ventilation rate is 0.1 to 0.3 ach (air changes per hour), 10 to 30 percent lower than that of the constant-volume ventilation control.DCV actually has taken into account the dilution power of the infiltrated outdoor air, which resuIts in a greater energy saving. .DCV often uses CO2 or a group of mixed gaseous contaminants as an indicator (controlled variable). The space concentration of CO2 itself provides feedback of the control system, whereas in a constant-volume ventilation control system, the volume flow rate of the outdoor air is difficult to measure exactly. In a VAV system, DCV is nonnally activated only in minimum outdoor air recirculating mode. During the air economizer cycle, the amount of outdoor air intake in the AHU or PU is always greater than the system outdoor air volume flow rate ~.sys. The amount of outdoor air intake is determined by the free-cooling requirement of the VAV system.

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CO2 Sensor or Mixed-Gases Sensor In DCV systems, two kinds of air sensors (air transmitter) are most often used as an indicator: CO2 sensors and a v olatile organic compound (V OC, or mixed-gases) sensor, as discussed in Sec. 5.4. Compared to a CO 2 sensor, a mixed-gases sensor senses a variety of mixed gases, is less expensive, and needs less maintenance. Ho wever, according to Stonier (1995), a CO 2 sensor has the following advantages: ●

●

●

The concentration le vel of CO 2 directly indicates the concentration of occupant-generated bioeff uents in the occupied zone. The concentration of CO 2 of outdoor air is usually rather steady , between 300 to 350 ppm in many locations in the United States. If each occupant e xhales 0.0106 cfm (0.3 L / min) of CO 2, from the balance of CO 2 by using Eq. (4.26), the ventilation rate supplied to the occupied zone can be roughly estimated. Or if the ventilation rate is known, then the number of occupants can be approximately estimated. If the v entilation rate of outdoor air supplied to a speci f c control zone and its number of occupants are known, then a higher-than-normal concentration level of CO 2 may be due to a poor ventilation effectiveness, because of the short circuit of supply air or improper location of the CO2 sensor.

ASHRAE Standard 62-1999 recommends that “Where only dilution v entilation is used to control indoor air quality , an indoor to outdoor dif ferential concentration not greater than about 700 ppm of CO 2 indicates comfort (odor) criteria related to body bioef f uents are likely to be satis f ed.” Standard 62-1999 also points out that using CO 2 as an indicator of bioef f uents does not eliminate the need for consideration of other contaminants. Steven Taylor, chair of the committee re vising Standard 62-1989, was asked in IAQ and Energy ‘98 New Orleans, October 24 to 27, if CO 2 is an appropriate indicator of indoor air quality . His response: “The use of CO 2 as a reliable indicator of IA Q has been o verblown, . . . CO2 is not itself a contaminant of concern at the concentrations found in most buildings. It can provide an indication of the amount of outdoor air being supplied to the space per person, but it cannot be used to accurately determine that rate unless the space is in nearly steady-state conditions. In an of f ce, it typically takes several hours to reach steady-state conditions, and people in most of f ces do not stay in one place that long.” Although CO2 is not itself an indoor contaminant or is a comprehensive indicator of IAQ, the indoor CO 2 concentration can pro vide an indication of outdoor air supplied to a space per person. Reardon (1994) and many other f eld measurements and investigations had clearly shown ●

●

The relationship between the concentration of space CO 2 and the air change rate (as sho wn in Fig. 23.2) The f oor-space concentration prof les of CO 2 or average concentration prof les of CO 2 in a return duct (as shown in Fig. 23.1)

The use of a CO 2 sensor to measure the space CO 2 concentration to control the amount of outdoor air intake in a v entilation control system is often an appropriate choice unless there is a still better sensor. As the outdoor ventilation air dilution is the primary issue to impro ve IAQ, the space concentration of CO2, to a certain degree, most often is an indirect indicator of IAQ. Location of CO2 Sensor Test results of Reardon (1994) sho wed that the CO 2 concentrations measured in the occupied f oor spaces are more variable than data measured in the openings of the return air shaft on the same f oor. The concentration of CO 2 of return air f airly well represents the space a verage CO 2 concentration

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FIGURE 23.2 Measured daily peak and daily average CO2 concentrations in the return duct versus air changes per hour. (Source: Reardon et al., ASHRAE Transactions, Part II, 1994)

data measured of that particular f oor after air mixes in the return air plenum. On the other hand, the f oor space concentration data were consistently higher than the return air data. This may be due to some degree of short-circuiting of outdoor air supply to the occupied zone in that f oor. Emmerich and Persily (1997) noted that air mo vements due to open doors or open windo ws cause differences and instabilities in CO 2 concentrations. Enermodal Engineering (1995) reported that an air system with CO 2 sensors in the return duct on each f oor showed no signi f cant difference in IAQ compared to a system with a CO2 sensor in the central return. Central control with a set point of 800 ppm provided similar performance but at a much reduced installation cost compared to individual zone control with a set point of 1000 ppm. Ruud et al. (1991) found that concentrations of CO2 measured at the wall and in the exhaust air were nearly identical with the wall-mounted sensor having an additional 2-min delay compared to that located in the exhaust air. Therefore, for most VAV systems, the suitable location to install a CO 2 sensor lies in the return main duct (or return air intake) after the return ceiling plenum and close to the AHU or PU for sensing an average CO 2 concentration. The operator should check the dif ference in CO 2 concentrations between the f oor space and the return duct periodically during operation so that the CO2 set point can be properly adjusted; e.g., the set point in the return duct is 100 to 200 ppm lower than the 1000-ppm target concentration level. In a VAV system, a single centrally-controlled sensor is recommended. For a dedicated ventilation and space recirculating system, a better location of a CO2 sensor is in the exhaust air main duct if there is an e xhaust air system for the v entilation air supply. A critical zone or a representative f oor space is also a candidate for the location of a CO2 sensor. Substantial Lag Time in Space CO2 Concentration Dilution Process SH__ ST__ LG__ DF

When a certain amount of outdoor air is e xtracted into an AHU or a PU, it should thoroughly mix with the recirculating air f rst and form an outdoor -recirculating mixture. The mixture is then

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supplied to a control zone in a VAV system, where it needs again to mix thoroughly with the space air in that control zone and to distribute over the entire conditioned space, in order to dilute the concentration of occupant-generated and other contaminants. It tak es a substantial lag time to perform such a mixing and dilution process. Elovitz (1995) reported that there is a substantial lag time in the space CO 2 concentration dilution process. If CO 2 rises, an increase in the outdoor air v entilation rate will not mak e it drop instantly. The control continues calling for more and more dilution and opens the outdoor damper wider and wider , until the drop in the CO 2 concentration is sensed by the CO 2. Recently a CO 2 minimum ventilation control system in an insurance of f ce building retrof t showed such a phenomenon. As the concentration of CO 2 raised above the set point, the control system k ept cranking the system up to 100 percent outdoor air intak e until e ventually the dilution ef fect showed up in the return air; at that point, the control crank ed the outdoor damper do wn until it w as fully closed, as the concentration of CO2 continued to fall long after the damper started to close. The following are two methods that will solve the time-lag problem: 1. Determine the relationship between the sensed CO 2 concentration and the required air changes per hour of minimum outdoor air intak e expressed in minimum outdoor damper setting in the AHU or PU, as shown in Fig. 23.2 during commissioning. For each sensed CO2 concentration in the return duct or f oor space, the DDC controller resets the required minimum outdoor air intake, in air changes per hour , as shown in Fig. 23.2, and the corresponding outdoor damper setting according to the determined relationship between them during commissioning. 2. When the sensed CO 2 concentration exceeds the set point, the outdoor damper will open an additional 5 percent. If, after a prespecif ed time interval, the sensed CO 2 level still exceeds the set point, the outdoor damper will open an additional 5 percent until the outdoor air intak e in the AHU or PU reaches the upper limit V˙ o,sys. The prespecif ed time interv al should be determined during commissioning.

Base Ventilation and Purge Mode In a CO 2-based demand-controlled v entilation system serving a typical of f ce building, even if the number of occupants drops to zero during the occupied period, the outdoor air intak e should not drop below a base v entilation. Base v entilation V˙ bv, in cfm (L / s), is the lo wer limit of minimum outdoor ventilation air required to dilute mainly the nonoccupant-generated air contaminants, such as building-generated volatile organic compounds (VOCs), to provide exhaust air in restrooms, and to maintain a positive space pressurization. For a typical off ce building, the amount of base ventilation required depends on interior b uilding materials, building construction, and the tightness of the building shell. The exhaust air in restrooms is often pro vided by the transfer air that also dilutes the building-generated VOCs. The base v entilation V˙ bv is the sum of the outdoor air required for restroom exhaust and exf ltration to maintain a space pressurization. Consider a one-story of f ce building of 20,000 ft2 (1859 m2) of f oor area with an exterior wall area of 10,000 ft2 (929 m2): ●

●

Two bathrooms ha ve typically 12 w ater closets or urinals. According to ASHRAE Standard 62-1999, these two restrooms need about 50  12  600 cfm or 0.03 cfm / ft2 (0.15 L / sm2). As discussed in Sec. 20.2, the f ow coeff cient of an a verage leaky building is 1.0 cfm / ft2 (5.0 L / s m2) of exterior wall at 0.3 in. WC (75 Pa). From Eq. (20.2), the overall leakage rate is V˙leak  Cflow Aw (p)0.65  1.0(10,000)(0.3)0.65  4572 cfm (2157 L / s) From Eq. (20.1) the effective leakage area of this average leaky building is: Ae,l  

V˙inf 4005√p



4572 4005√0.3

 2.08 ft 2 (0.194 m2)

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Under base ventilation, the amount of e xf ltration required to maintain a space positi ve pressure of  0.005 in. WG (1.25 Pa) can be calculated as  V˙inf  4005Ae,l √p  4005  2.08 √0.005  589 cfm (278 L / s) The volume f ow rate of base ventilation is then V˙bv  600  589  1189 cfm or roughly 1200 cfm (566 L / s), that is, 1200 / 20,000  0.06 cfm / ft2 (0.30 L / s m2) of f oor area, which is about 43 percent of the value of the minimum outdoor air required (V˙bv  0.43 V˙o, sys). Formaldehyde and other possible indoor air contaminants in the space should be measured. If their concentrations are above allowable levels, or if the building shell is too leaky, the base ventilation should be increased accordingly. The amount of outdoor v entilation air provided for CO 2-based demand-controlled ventilation in a VAV system in an of f ce building with a ceiling height of 9 ft (2.7 m) during v arious operating modes is as follows: cfm / ft2 (L / s m2) of f oor area Air economizer cycle, free-cooling mode Minimum outdoor air recirculating mode V·o,sys (at design occupancy) Base ventilation V·bv (lower limit)

ach

0.14 – 1.3 (0.7 – 6.5)

0.93 – 10

0.14 (0.7) 0.06 (0.25)

0.93 0.40

The volume f ow rate of minimum outdoor air during recirculating mode at design occupanc y is the calculated system outdoor air intak e in the AHU or PU at summer design conditions V˙ o,sys by Eq. (23.3). It is recommended that a purge operation take place every weekday morning if the concentration of the volatile organic compounds or other indoor contaminants is abo ve the allowable level before the building is occupied or if the dif ference between the indoor and outdoor CO 2 concentrations is greater than 200 or 300 ppm.

A CO2-Based Demand-Controlled Ventilation System Figure 23.3 shows a CO 2-based demand-controlled ventilation system in a VAV system. It consists of a CO 2 sensor, a DDC unit controller , a minimum outdoor damper , an economizer damper, a recirculating damper, an exhaust damper, and damper motors. The CO2 sensor is located in the return duct that connects the return ceiling plenum and the return f an inlet. During minimum outdoor air recirculating mode: ●

●

●

●

SH__ ST__ LG__ DF

The supply fan and the minimum outdoor air damper of this VAV system must ha ve the capacity to provide the required amount of outdoor air at a v olume f ow rate of V˙ o,sys, in cfm (L / s), calculated by Eq. (23.3). Outdoor air should be supplied to the occupied zone with a high v entilation effectiveness for the dilution of the concentration of CO2 and other indoor contaminants. During commissioning, damper settings of required minimum outdoor air at design occupanc y V˙o,sys, and base v entilation V˙ bv should be clearly mark ed, and the relationship between the measured CO 2 concentration and the minimum outdoor damper opening setting from V˙ bv up to V˙ o,sys should be determined after tests. When a VAV system is in the occupied, minimum outdoor air recirculating mode, after supply and return f ans start, the minimum outdoor damper should be opened to an opening setting of base ventilation.

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DDC controller

Economizer damper

Minimum outdoor damper Recirculating damper

Exhaust damper

Return fan

Supply fan

Supply duct

CO2 Return duct

CO2

CO2 sensor

FIGURE 23.3 A CO2-based demand-controlled ventilation control in a VAV system.

●

●

●

If the signal of CO 2 concentration sensed by the CO 2 sensor located in the return duct or a representative f oor space is sent to the DDC controller , the DDC controller resets the minimum outdoor air intak e V˙ o,sys to a corresponding minimum outdoor damper opening setting according to the relationship between the measured return duct or space CO 2 concentration and the damper opening setting. If the required amount of minimum outdoor air corresponding to the sensed CO 2 level is less than the base v entilation, the DDC controller resets the minimum outdoor air intak e down to the base ventilation V˙ bv. The recirculating damper is opened 100 percent. When the VAV system is in the of f position, the minimum outdoor , economizer, and e xhaust dampers are closed, and the recirculating damper is opened fully.

During 100 percent outdoor air free-cooling mode, the air economizer control o verrides the CO 2based DCV control.

Application of CO2-Based Demand-Controlled Ventilation System Chen and Demster (1996) and Emmerich and Persily (1997) recommended the use of CO DCV systems under the following circumstances: ●

●

●

2-based

The existence of high occupancy density with considerable f uctuations in occupancy Low pollutant emissions from nonoccupant-generated sources In location with climate where heating or refrigeration equipment has a signi f cant number of working hours annually

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23.12

For a multizone VAV system that serv es control zones with an ob vious occupancy variation schedule, such as in an of f ce building where of f ces and conference rooms use a common AHU or PU, when many occupants in the of f ces may go out for lunch, but a meeting is still taking place in a conference room, remedies should be considered during design to supply additional outdoor air to the conference room under such circumstances.

23.3 MIXED PLENUM PRESSURE CONTROLS Basics As discussed in Secs. 22.5 and 22.6, for VAV systems operating at minimum outdoor air recirculating mode, when the supply v olume f ow reduces during part-load operation, the static pressure in the mixed plenum (mixing box) where outdoor air mix es with the recirculating air tends to rise and causes a small pressure difference between the mixed plenum and outdoors. The minimum outdoor air intake in the AHU or PU reduces accordingly . Mixed plenum pressure control aims to maintain a nearly constant static pressure in the mix ed plenum and thus keeps a nearly constant pressure difference between the mix ed plenum and outdoors to e xtract a nearly constant amount of minimum outdoor air in both design load and part-load operation.

Case Study: Mixed Plenum Pressure Control Monitoring Plenum Pressure Graves (1995) developed a mix ed plenum pressure control system in a VAV system by monitoring the pressure difference between the mixed plenum pressure and outdoors atmospheric pressure and modulating the recirculating damper opening to maintain a nearly constant mix ed plenum pressure, as shown in Fig. 23.4a.

DDC controllers

Recirculating Exhaust damper damper

Recirculating damper

M

M

Return fan

Supply fan P

Mixing plenum

Outdoor louver

Mixing plenum P

Minimum outdoor air damper

Economizer damper

Filters

Economizer damper

(a) P

SH__ ST__ LG__ DF

Pressure sensor

(b) Mixing temperature element

FIGURE 23.4 Mixed plenum pressure controls: (a) monitoring mix ed plenum – outdoors pressure dif ferential; (b) monitoring pressure drop across the outdoor damper and louver.

Minimum outdoor air damper

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System Description. The VAV system includes tw o rooftop packaged units. Each has a capacity of 8300 cfm (3917 L / s) with a v ariable-speed supply f an and a v ariable-speed return f an. A f xed minimum oudoor damper supplies 1800 cfm (850 L / s) of outdoor air to the conditioned space, and an economizer damper supplies the additional outdoor air during the air economizer c ycle. When the outdoor temperature is higher than 70 °F (21.1°C), the VAV system is operated in minimum outdoor air recirculating mode. A DDC unit closes the economizer damper and opens the recirculating damper. Whenever the packaged unit is operating, the exhaust damper is opened 100 percent. In the mixed plenum, outdoor air is drawn through a louver and damper on one side of the unit, and e xhaust air is dischar ged through the e xhaust damper and louv er on the opposite side. The opening of the f xed minimum outdoor damper w as set to pro vide 1800 cfm (3917 L / s) with a plenum pressure of  0.50 in. WG (  125 P ag). At design conditions, the air f ow and pressure drops across dampers and louvers are as follows:

Recirculating damper Minimum outdoor damper Economizer damper Exhaust damper Outdoor air louver Exhaust air louver

Airf ow, cfm (L / s)

p, in. WC (Pa)

6500 (3067) 1800 (850) 6500 (3067) 6500 (3067) 8300 (3917) 6500 (3067)

0.55 (137) 0.30 (75) 0.30 (75) 0.10 (25) 0.25 (63) 0.25 (63)

Minimum Ventilation. A DDC control system is used for the controls of this rooftop unit including a mix ed plenum pressure control. A static pressure sensor located in the mix ed plenum senses the plenum pressure, and a shielded outdoor air pressure reference is located abo ve the rooftop packaged unit. The DDC system modulates the recirculating damper to maintain a static pressure difference between the mixed plenum pressure and outdoors pm,p of 0.55 in. WC (137 Pa) 0.05 in. WC (12.5 Pa). If pm,p is too large, the recirculating damper is gradually opened until it reaches the set point. If pm,p is too small, the recirculating damper is gradually closed until the set point is maintained. According to f eld-measured results, the DDC mix ed plenum pressure control system has achieved a very stable control of pm,p. Supply and Return Fans. The speed of the supply fan is modulated to maintain a duct static pressure of 1.50 in. WG (375 P a) at the location of the duct static pressure sensor in the supply main duct during both design and part-load operations. The volume f ow rate of the return f an is al ways approximately 1800 cfm (850 L / s) less than the supply air v olume f ow. A linear relationship between the supply f an speed and return f an speed w as determined by testing the rooftop packaged VAV system during commissioning. Roughly , when the supply f an was at 100 percent f an speed, the return fan was at 90 percent f an speed. When the supply f an was operated at approximately 40 percent speed, the recirculating damper was closed and the return fan was turned off. Actual operation of this system sho wed that the supply f an speed w as always above 40 percent because the minimum supply volume f ow rate required by the VAV boxes was 35 percent, or about 2900 cfm (1369 L / s). Monitoring Pressure Drop of Louver and Damper Controlling Mixed Plenum Pressure Mumma and Wong (1990) f rst introduced a kind of mixed plenum pressure control to modulate the recirculating damper to maintain a predetermined pressure drop across an outdoor air louv er and damper combination, as sho wn in Fig. 23.4 b. This mix ed plenum pressure control consists of a minimum outdoor damper , an economizer damper , a recirculating damper , a pressure sensor , a DDC controller, damper motors, a variable-speed supply fan, and a variable-speed relief fan. It proceeds as follows:

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CHAPTER TWENTY-THREE ●

●

●

The minimum outdoor damper is f xed at an opening setting which deli vers minimum outdoor air V˙ o,sys. This damper setting is determined during commissioning. When the supply f an reduces its speed at part-load operation, the pressure dif ferential across the outdoor louver and minimum outdoor damper combination drops. As this reduction is sensed by the pressure transducer, a DDC unit closes the recirculating damper to maintain a preset pressure drop across the combination, and therefore ensures that the amount of minimum outdoor air V˙ o,sys intake is always extracted. If the supply f an increases its speed, the pressure drop across the outdoor louv er-damper combination rises. The DDC controller opens the recirculating damper to reduce the pressure drop across the combination and maintains the required amount of minimum outdoor air V˙ o,sys intake.

Applications In VAV systems, mixed plenum pressure controls ensure time-based constant v olume of minimum outdoor ventilation air supply to the conditioned space. The controls are simple and ef fective and can be used for buildings where there is less variation in occupancy.

23.4 OUTDOOR AIR INJECTION FAN, DIRECT MEASUREMENT, AND FAN TRACKING SYSTEMS Outdoor Air Injection Fan To provide a guaranteed amount of minimum outdoor air V˙ o,sys to a VAV system, recently an outdoor air injection f an has sometimes been installed in parallel to the outdoor air passage leading to the main supply fan, as shown in Fig. 23.5. An outdoor air injection fan has the following operating charateristics:

Ceiling plenum

ru

Supply duct Slot diffuser

Return duct

Return slot Supply fan

Return fan Outdoor air injection fan Minimum outdoor air damper

Mixed plenum

m Economizer damper

SH__ ST__ LG__ DF

FIGURE 23.5 A VAV system with an outdoor air injection fan.

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●

●

●

●

23.15

__RH

The maximum supply v olume f ow rate of this outdoor air injection f an should be equal to the amount of minimum outdoor air V˙ o,sys required. The fan total pressure of this injection fan is often equal to the total pressure loss between the outdoor air intake and the mixed plenum inlet. The ductwork of the injection fan system is sized so that the v elocity pressure is easily measured. An airf ow measuring station with a velocity sensor sends signals to a DDC controller and adjusts the fan speed as well as its v olume f ow through a v ariable-speed drive to correct for wind and backpressure effects. The injection f an system is a constant-v olume minimum outdoor air supply system and operates whenever outdoor air is required. According to Avery (1996), if an outdoor air injection fan is installed with a supply and return fan combination, the mixed plenum may be v aried from ne gative pressure when the VAV system is operating on 100 percent outdoor air or mix ed air control to positi ve pressure during minimum outdoor air cooling mode.

Outdoor injection f an systems ha ve the adv antages of easy determination and measurement of the minimum outdoor air f ow and control its required amount. They have been used in man y applications by designers. Ho wever, a small injection f an is often ener gy-ineff cient. Also, an injection fan needs e xtra f an and ductw ork and results in a more complicated system pressure distrib ution and control, especially for a retrof t in existing buildings. Direct Measurement of Minimum Outdoor Air Intake Minimum ventilation control by direct measurement of the minimum outdoor air f ow during minimum outdoor air recirculating mode needs a DDC controller to open or close the minumum outdoor damper and recirculating damper according to the directly measured air f ow signal, to provide the required amount of minimum outdoor air for v entilation. The outdoor air at the intak e is usually designed according to a v elocity of 500 fpm (2.5 m / s). During part-load operation, it may reduce to 35 percent of the design v alue, that is, 175 fpm (0.87 m / s), or a v elocity pressure of 0.002 in. WC (0.5 P a). In such a v elocity pressure, even a multiple-point Pitot-tube a veraging probe has diff culty sensing a 10 percent change in outdoor air v olume f ow rate. The direct measurement of minimum outdoor air intak e cannot accurately sense, measure, and provide minimum ventilation control. Fan Tracking Systems In a VAV system with a supply and return f an combination using a f an tracking system for minimum ventilation control, the volume f ow rate of the supply f an is v aried according to the system load while at the same time maintaining a preset duct static pressure at the static pressure sensor in the supply main duct. The volume f ow rate of the return f an is controlled to “track” the supply fan at a constant cfm (L / s) less than the supply v olume f ow rate. If the e xhaust damper is tightly closed and does not leak, if there is no inf ltration or exf ltration, and if one ignores the difference in densities between the outdoor and recirculating air , the volume f ow rate of the supply f an minus the volume f ow rate of return f an V˙ sf  V˙ rf is the amount of outdoor air intak e in the AHU or PU, designated V˙o, in cfm (L / s). Because outdoor air is often from one- f fth to one-se venth of the supply air , consider a supply fan with a design v olume f ow V˙ sf of 10,000 cfm (4719 L / s) and a return f an with a design v olume f ow V˙ rf of 8500 cfm (4011 L / s), which results in a design outdoor air intak e of 10,000  8500  1500 cfm (708 L / s). Due to the measuring and instrumental errors, V˙ sf reads 2 percent low and V˙ rf reads 2 percent high. The controlled outdoor air intake is then 9800  8670  1130 cfm (533 L / s), which is about 25 percent less than the design v alue. Field experience had proved that fan-tracking minimum v entilation control f ailed to pro vide the required amount of outdoor air correctly and energy-eff ciently.

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However, for a VAV system with a supply and return f an combination, when a supply f an reduces its supply v olume f ow rate as the space load drops, the volume f ow of the return f an must decrease accordingly, to meet the following requirements: ●

●

●

Maintain a desirable space pressurization. Exhaust a speci f ed amount of e xhaust air and recirculate a speci f ed amount of recirculating air in the AHU or PU. Form a new airf ow balance in the space.

The volume f ow of the return f an tracks the volume f ow of the supply f an in order to pro vide outdoor air that e xactly equals V˙ sf  V˙ rf but it f ailed to do so in actual practice. Ho wever, reduce the speed of the return fan approximately according to V˙ sf  V˙ rf while the outdoor air intake is guaranteed by DCV or mix ed plenum pressure control, and this gi ves a successful minimum v entilation control experience. This is discussed in a later section.

23.5 CONFERENCE ROOMS Conference rooms ha ve characteristics of high occupanc y density (20 ft 2 /person or 2 m 2 /person) and a v ariable occupancy with a schedule that may be quite dif ferent from pri vate and general off ces. Conference rooms are often critical spaces to pro vide v entilation for acceptable indoor air quality in off ce buildings. The following are recommendations of minimum v entilation control for conference rooms in off ce buildings: Use a separate AHU or PU for the conference room with its o wn outdoor air intake and minimum ventilation control, if possible. Increase the total supply air f ow rate as well as the outdoor v entilation air rate. According to ASHRAE Standard 62-1999, for a conference room with an occupanc y density of 20 ft 2 /person (2 m2 /person) instead of 140 ft 2 /person (13 m2 /person) in an off ce, the outdoor air requirement for ventilation should be increased from 0.14 cfm / ft2 (0.7 L / sm2) to about 1.0 cfm / ft2 (5 L / sm2). In a conference room, the supply air v olume f ow rate should be determined based on the outdoor air requirement for minimum v entilation. Even the VAV system load and the supply v olume f ow rate reduce during the occupied period; based on the ASHRAE Standard 62-1999 speci f ed outdoor air requirements, most of the conference rooms still have the required minimum outdoor air to provide ventilation for acceptable indoor air quality. ●

●

23.6 SPACE PRESSURIZATION AND RETURN VOLUME FLOW CONTROLS As discussed in Sec. 20.3, for low-rise buildings except on stormy days, building pressure characteristics are dominated by air systems (including mechanical ventilation systems) most of their operating time.

Characteristics of Space Pressure Control The following are characteristics of space pressure control: ●

SH__ ST__ LG__ DF

According to the air f ow balance, the mass f ow rate of the outdoor air entering air system m˙o, in lb / min (kg / min), is balanced with the mass f ow rate leaving the system, or. m˙o  m˙inf  m˙ex,r  m˙eu

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where m˙ex,r, m˙eu  mass f ow rate of air e xhausted, including air e xhausted from the restroom and from the AHU, lb / min (kg / min). In most conditioned space, a space positive pressure  pim is thus formed because of a surplus of outdoor air  m˙inf entering the system, which is then e xf ltrated through the gaps and openings on the building shell to outdoors and can be calculated as m˙inf   V˙inf r V˙inf  4005Ae,l √pro

(23.4)

where  V˙ inf  volume f ow rate of exf ltration, cfm (m3 /min) Ae,l  effective leakage area on building shell, ft2 (m2)

r  space air density, lb / ft3 (kg / m3) pro  space and outdoor pressure difference, in. WC (Pa) ●

●

As discussed in Sec. 20.3, the space and outdoor pressure dif ference often has a v alue between  0.005 and 0.03 in. WC (1.25 and 7.5 P a). If a pressure sensor in a space pressurization control system needs to sense a 10 percent change in pim, that is, typically  0.0005 in. WG (0.13 Pa), it would be diff cult and expensive to specify such a level of accuracy for most comfort systems. For an a verage leaky one-story of f ce building with an e xterior wall area of 10,000 ft 2 (929 m 2) and a f oor area of 20,000 ft2 (1858 m2), as discussed in Sec. 23.2, from Eq. (20.2) V˙leak  Cflow Aw(p)0.65  1.0(10,000)(0.3)0.65  4572 cfm (2157 L / s)

From Eq. (23.4), for this average leaky off ce building, the effective leakage area is Ae,l 

V˙inf 4005√pro



4572 4005√0.3

 2.08 ft 2 (0.193 m2)

Also from Eq. (23.4) the outdoor air intak e required to maintain v arious space positi ve pressures and to exf ltrate them in this a verage leaky off ce building with 10,000 ft 2 (929 m 2) of exterior wall is as follows: Required outdoor air intake pro, in. WC (Pa)

Exterior wall area, cfm / ft2 (L / s m2)

Floor area, cfm / ft2 (L / s m2)

 0.005  0.01  0.02  0.03  0.05  0.1

0.059 (0.30) 0.083 (0.42) 0.118 (0.60) 0.144 (0.73) 0.186 (0.95) 0.263 (1.33)

0.0295 (0.15) 0.0417 (0.21) 0.059 (0.30) 0.072 (0.36) 0.093 (0.47) 0.132 (0.67)

VAV Systems Return / Relief Fan Volume Flow Control For a VAV system with either a supply and relief fan combination as shown in Fig. 22.3, or a supply and return fan combination as shown in Figs. 22.1 and 22.4, if the volume f ow rate of supply air reduces during part-load, the amount of return air extracted along points ru should be reduced accordingly in order to provide the required outdoor ventilation air and to maintain a desirable space pressure. Modulating the speed of the return / relief f an with adjustable-frequenc y drives, and thus its volume f ow, to maintain space pressure within predetermined limits is a simple, direct, and effective means of return / relief fan volume control. Place a pressure sensor in a representative f oorspace. The location of the indoor pressure sensor must be away from the door and openings to the ouside, as well as away from the elevator lobbies. The outdoor reference pressure should be typically 10 to 15 ft (3 to 4.5 m) abo ve the building and

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CHAPTER TWENTY-THREE

well shielded from wind and air velocity effects. As a DDC unit controller reduces the speed of supply fan to maintain a preset duct static pressure during part-load, another DDC unit controller also modulates the speed of return f an to adjust the space pressure pr, f oating within a dif ferential or dead band between an upper limit of  0.03 in. WG ( 7.5 Pag) and a lo wer limit of  0.005 in. WG (  1.25 P ag) using a f oating control mode. The e xhaust / relief damper is often opened between 0 and 30 percent during minimum outdoor v entilation recirculating mode and opened 100 percent during all outdoor air economizer mode. If a VAV system using a supply and relief f an combination is operated during minimum outdoor air ventilation recirculating mode with a constant-v olume f ow of minimum outdoor v entilation air provided by mixing plenum pressure control, and if the building that this VAV system serves has an air tightness between an average leaky and leaky building, then ●

●

Either the relief f an is turned of f and the relief damper is closed, so that all the outdoor air is exhausted from the restrooms, other exhausts, and exf ltrates through the exterior leakage area. Or the relief f an is turned on at lo w speed and the relief damper e xhausts only a v olume f ow less than 30 percent of the designed volume f ow. The remaining portion of the outdoor air is exhausted outdoors through restrooms, other exhausts, and exf ltrates through the exterior leakage area.

During minimum v entilation recirculating mode, these e xhausts and e xf ltration, plus the recirculating air e xtrated along points ru (recirculating / exhaust chamber) and m (mixing plenum) by the supply f an are automatically balanced with the design and reduced supply v olume f ow rate at part-load. Another alternative of return / relief fan volume control in VAV systems is air f ow tracking. Required return / relief volume f ow corresponding to v arious reduced supply air f ow at part-load are determined during commissioning by means of multipoint Pitot tube measurements. Return / relief fan speed and v olume f ow can then be modulated at part-load accordingly . Airf ow tracking is an open-loop control. Supply f an volume f ow must be adjusted to match with the v olume f ow of the return fan during morning warm-up and cool-down mode.

23.7 DISCHARGE AIR TEMPERATURE CONTROLS Basics In a multizone VAV system, the dischar ge air temperature Tdis, in °F ( °C), is the conditioned air temperature leaving the supply outlet of the AHU or PU, where Tdis  Tcc  Tsf. Here Tcc represents the temperature of the conditioned air lea ving the coil, and Tsf indicates the temperature rise across the supply f an, both in °F ( °C). Typically, Tdis  55  2  57°F (13.9 °C). The difference between the supply temperature from the ceiling dif fuser mounted on the ceiling plenum and the discharge air temperature Ts  Tdis, in °F (°C), is the rise or drop in temperature resulting from the duct heat gain or duct heat loss Tdu. For a typical multizone VAV system, Ts  Tdis is usually a function of the temperature dif ference between the supply air and the ambient air surrounding the supply duct and the supply v olume f ow rate. Its relationship can be e xpressed as Ts  Tdis Tdu. As discussed in Secs. 20.18 and 21.3, for a summer indoor design temperature of 75 °F (23.9°C), a lower conditioned air off-coil temperature Tcc and discharge air temperature Tdis result in the following: ●

●

SH__ ST__ LG__ DF

A greater supply temperature dif ferential Ts  Tr  Ts, a lo wer space relati ve humidity , a smaller supply volume f ow rate V˙s, fan sizes, fan energy use, duct, terminal, and diffuser sizes A lo wer chilled w ater temperature entering and lea ving the cooling coil Twe and Twl, a lo wer evaporating temperature in the DX coil Tev, and therefore higher po wer input to the refrigeration compressors

Excluding the the cold air distrib ution in the ice storage system, discussed in Chap. 31, a Ts between 15 and 20°F (8.3 and 11.1°C) during summer design conditions is most widely used in air

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systems. The following list relates parameters when Ts  15 and 20 °F e xcept in an ice storage system in which Ts may increase up to 30 °F (16.7°C). All units are in °F (°C), except the space relative humidity r is in percent: Ts

Tcc

Tsf

Tdis

Tdu

Ts

Tr

r

15 (8.3) 20 (11.1)

55 (12.8) 50 (10.0)

2 (1.1) 2 (1.1)

57 (13.9) 52 (11.1)

3 (1.7) 3 (1.7)

60 (15.6) 55 (12.8)

75 (23.9) 75 (23.9)

50 45

System Description Discharge air temperature control consists of the following control loops, as shown in Fig. 23.6a: ●

●

●

●

●

Economizer control loop. For a temperature air economizer control, it consists of an outdoor temperature sensor T2, a recirculating air temperature sensor T3, a DDC controller, outdoor damper actuators D1 and D2, a recirculating damper actuator D3 and a relief damper actuator D4. Cooling coil control loop. It consists of a dischar ge air temperature sensor T1, a DDC controller, and a cooling coil valve actuator V1. Mixed air contr ol loop. It consists of a mix ed air temperature sensor T4, a DDC controller , outdoor damper actuators D1 and D2, and a recirculating damper actuator D3. Heating coil control loop. It consists of a dischar ge air temperature sensor T1, a DDC controller, and a heating coil valve actuator V2. Low-limit control loop (coil freeze protection). It consists of a temperature sensor T4, a DDC controller, and outdoor air damper actuators D1 and D2.

For VAV systems, discharge air temperature control is a kind of cooling and heating capacity control so that conditioned air is dischar ged from an AHU or PU with a predetermined temperature and relative humidity, such as 57°F (13.9°C) and 95 percent relati ve humidity, and can perform the following requirements: 1. Provide a kind of conditioned air for zone temperature control by v arying the supply v olume f ow rate. 2. Based on the conditions of air lea ving the coil sensed by sensor T1, provide economizer control and modulate the cooling or heating capacity of the AHU or PU to maintain a predetermined Tdis. 3. In VAV systems, if an AHU or a PU serv es both perimeter and interior zones, usually perimeter heating is provided by the heating coil installed in the reheating VAV boxes or fan-powered VAV boxes or by baseboard heating directly . In such a condition, Tdis should meet the free-cooling requirement in interior zones even in winter. 4. According to the outdoor temperature or system load, reset Tdis to an optimum v alue for both effective operation and energy saving. For a cooling coil or heating coil in the AHU, the DDC controller modulates the coil valves continuously. However, in a rooftop packaged unit, refrigeration capacity is often adjusted by c ycling the compressors on and off in the separate refrigerant circuits of the DX coils. The capacity control of packaged units is discussed in Sec. 11.8. Most VAV systems use the air economizer or w ater economizer as the f rst-stage cooling to sa ve energy. Cooling coil, mixed air, and heating coil control loops are operated in sequence so as to maintain a predetermined discharge air temperature while preventing simultaneous cooling and heating. For a typical AHU, the set point of the dischar ge air temperature at summer design conditions may be 57°F (13.9°C). To avoid sudden repeated changes in operating mode among the cooling coil and mixed air control loops, resulting in unstable operation, the set point for mix ed air control is better set at 55°F (12.8°C) as shown in Fig. 23.6b.

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23.20

Relief fan

Relief D4 air damper Recirculating air damper

T3

Minimum outdoor air damper D1

Heating Cooling coil coil

D3

T2 D2

Supply fan

T1

T4 Economizer damper

V2

V1

AC

BI1 AI1 2 3 4 5 6

T3

Damper actuator

F

Flow measuring station

Capacity, %

AC

100 50

AC inverter

V0 100%

0 20

(a)

w lb/lb

70 Mixing air

Cooling coil

Min 0

.

57

.

V0

62 Tdis, F

50

cc

r

s

40

40

(b)

0.012

60

m

0 20 40 60 80 100 Outdoor air temperature T0, F

0.016

50 si sx

ru

m

DF

0.008

r ru si

sx

60

70

80

Supply air for interior zone Supply air for perimeter zone (c)

FIGURE 23.6 Discharge air temperature control for a typical AHU: (a) control diagram; ( b) output diagram; ( c) air conditioning cycle.

SH__ ST__ LG__

D

AO1 2 3 DO1 AO4

Temperature sensor

D

T0

90 T, F

0.004

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The modulation of valve V1 in the cooling coil control loop of dischar ge air temperature control can be performed in parallel with the control operations in 100 percent air economizer control. Although mixed air control is part of air economizer control, it is also a part of the sequence control in a discharge air temperature control. Mix ed air control mix es the outdoor air with the recirculating air by modulating the economizer damper and the recirculating damper as to pro vide a mix ed air temperature, such as 55°F (12.8°C) for the sake of free cooling in the interior zone. Mix ed air control functions only when To 55°F (12.8°C) and ends during minimum outdoor air heating mode. For VAV reheat and fan-powered VAV systems, heating is often provided by a water heating coil or an electirc heating coil in a terminal directly abo ve the conditioned space. The modulation of the VAV box and the control valve of the heating coil in sequence forms the zone cooling and heating control. For a single-zone VAV system, a DDC unit actuates the economizer dampers, cooling coil valve, economizer and recirculating dampers, and heating coil v alves in sequence, based on the input signal from a zone temperature sensor to maintain a predetermined zone temperature. The discharge air temperature control is replaced by a zone temperature control.

Operation of Air Economizer and Outdoor Air Intake — Case Study The outdoor air intak e of a f xed dry-bulb air economizer in an AHU of a VAV system sho wn in Fig. 23.7 is from Nabinger et al. (1994). In Fig. 23.7, the abscissa of the diagram is the indoor -outdoor temperature dif ference of the conditioned space that the VAV system serv es and the outdoor

FIGURE 23.7 The outdoor air intak e of a temperature air economizer in a et al., ASHRAE Transactions, Part II, 1994. Reprinted by permission.)

VAV system. ( Modif ed from Nabinger

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air temperature, both in °F (°C). The ordinate is the measured outdoor air intak e of a f xed dry-bulb air economizer in a VAV system, expressed in supply f ow density, cfm / ft2 (L / sm2), and ach based on a ceiling height of 9 ft (2.74 m). According to Nabinger et al. (1994), the test off ce building is located in Overland, Missouri. The building consists of seven f oors with a f oor area of 378,000 ft 2 (35,100 m2). Construction began in 1988, and occupancy began in late 1990. The air conditioning system is a VAV system with fan-powered VAV boxes in the perimeter zone and VAV boxes in interior zones. There are all together 15 AHUs with a total supply v olume f ow rate of 267,000 cfm (126,000 L / s), and the design minimum outdoor air intak e for this b uilding is 50,600 cfm (23,900 L / s). Outdoor air intake in the AHUs was measured by using the trace gas decay technique with SF 6 as the trace gas. Trace gas was injected into each of the 15 AHUs and allowed a period of about 20 to 30 min to mix. Four decay tests were conducted each day. The tracer gas injection rates were calculated based on a tar get concentration of 150 ppb (parts per billion). Tracer gas concentration was monitored at 19 indoor locations and an outdoor location with the concentration measured at each location every 10 min. Operation of Air Economizer. When the outdoor air temperature To was below a set point v alue, such as To 30°F (1.1°C), the AHU or PU was set to minimum outdoor intak e to protect it from freezing and to avoid heating more outdoor air than necessary. The AHU was operated in minimum outdoor air heating mode, as shown in Fig. 23.7. When the outdoor temperature 30 °F  To  70°F ( 1.1°C  To  21.1 °C), the DDC controller in mixed air control modulated the economizer and recirculating dampers to pro vide a mixture of outdoor and recirculating air at the required mix ed air temperature, such as 55 °F (12.8°C). The air economizer was operated in mixed air control. Here the outdoor air temperature from which the mix ed air control be gan To.mix  30°F ( 1.1°C) depended on the percentage of outdoor air among the supply air, and on the load ratio. When the outdoor air temperature To > 70°F (21.1°C), it was more economical to cool the recirculating air than the outdoor air. The minimum outdoor and recirculating dampers opened, and the econimizer damper was closed. The AHU or PU was operated in minimum outdoor air cooling mode. In Fig. 23.7, the highest measured outdoor air supply f ow density w as out 0.52 cfm / ft2 (2.6 L / s ·m2), which was lower than the average building supply f ow density 267,000 / 378,000  0.71 cfm / ft2 (3.5 L / s ·m2). The measured minimum outdoor air intak e w as lo wer than the ASHRAE Standard 62-1999 specif ed minimum outdoor air intak e of 0.93 ach, or 0.14 cfm / ft2 (0.7 L / s ·m2). The a verage supply f ow density for VAV systems at design f ow for a typical f oor (including perimeter and interior zones) usually v aries between 0.8 and 1.3 cfm / ft2 (4 and 6.5 L / s ·m2). The reason for having lower f ow density was a lower space load density at part load. There were a certain number of minimum outdoor air intak es even lower than the base v entilation. It is interesting to note that the amount of minimum outdoor air intak e during heating mode was higher than the amount of minimum outdoor air intak e during cooling mode. This may be due to the higher density of cold outdoor air in cold weather than the lo wer density of outdoor air in warm weather. Carbon Dioxide Concentrations. From November 1991 to No vember 1992, CO2 measurements under full occupancy revealed that instantaneous daily peak v alues in the return ducts ranged from 450 to 750 ppm. In the space, they ranged from 500 to 850 ppm. On a fe w isolated occasions, CO2 concentrations in the return ducts were as high as 1100 ppm; in the space, they were as high as 1400 ppm when some of the AHUs were shut off. Discharge Air Temperature Reset Discharge air temperature or supply temperature reset saves energy in the following ways: SH__ ST__ LG__ DF

●

It reduces the simultaneous cooling and heating if VAV reheating and dual-duct VAV systems are used.

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●

●

●

23.23

__RH

It lowers the duct heat gain or loss. It raises the suction temperature of the refrigeration system. It increases the number of hours that the refrigeration system can be shut of f during an air or w ater economizer cycle. It increases the zone supply volume f ow rate at low zone loads.

Discharge air temperature reset also has the following drawbacks: ●

●

Lower fan energy savings because less volume f ow can be reduced A higher space relative humidity during cooling mode operation

During part-load operation, because the reduction in space load and supply v olume f ow rate in the perimeter zone is dif ferent from that in the interior zone, a separate dischar ge air temperaure reset schedule for perimeter and interior zones is bene f cial if separate AHUs or PUs for perimeter and interior zones are adopted, or a dual-duct system is used. For a conditioned space that requires strict relati ve humidity control, discharge air temperature reset is not recommended. For a VAV system that uses the same AHU or PU to serve both perimeter and interior zones, discharge air temperature reset must consider the higher space cooling load at part load in the interior zone. In VAV systems using inlet v anes for duct static pressure and system supply volume f ow control, discharge air temperature reset is preferable because it may sa ve more energy. In VAV systems an analysis is often bene f cial to determine whether dischar ge air temperature reset saves more energy and is cost-effective.

23.8 DUCT STATIC PRESSURE AND FAN CONTROLS Duct Static Pressure Control In a VAV system, if more than one half of the VAV boxes close their dampers at the same time to reduce the volume f ow supplied to various zones at part-load operation, the static pressure of the supply duct will be raised. Such a high static pressure may cause objectionable noise and v olume f ow in some of the terminals and may result in unstable operation. At part load, the volume f ow and the fan total pressure of the supply fan as well as the return fan should be reduced to achieve a balanced and optimized system operation if a supply and return f an combination is emplo yed. The reduction of the volume f ow of the return fan is discussed in Sec. 23.6. The purpose of supply duct static pressure control (or simply duct static pressure control) is to maintain a nearly constant static pressure at a specific location in the main supply duct and, at the same time, to sa ve f an ener gy use. A duct static pressure control system consists mainly of a static pressure sensor, a DDC controller, and an adjustable-frequency variable-speed drive, or an inlet vanes actuator, as shown in Fig. 23.8. When the static pressure sensed by the static pressure sensor at the predetermined location in the supply main duct is higher than a set point, the DDC controller actuates the ac in verter or positions the inlet v anes to reduce the f an v olume flow and fan total pressure, so as to maintain a nearly constant static pressure at the location of the sensor. Duct static pressure shows a small capacitance lag, or time constant. Therefore, duct static pressure responds quickly to changes in the supply f an output. Proportional-inte gral-derivative (PID) control action with a wider proportional band is more appropriate. A high-limit pressure sensor should be installed after the supply f an outlet to pre vent damaging the duct system by means of e xcessive pressure due to malfunction, or closing of the f re or smoke dampers. When these circumstances arise, the excessive static pressure buildup due to the blockage of the air passage can be sensed by the high-limit pressure sensor, and the DDC controller will send a signal to turn off the fan.

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FIGURE 23.8 Duct static pressure control.

Set Point of Duct Static Pressure and Sensor’s Location The set point of the duct static pressure should be the minimum static pressure such that when added to its associated v elocity pressure, their sum — the total pressure — should overcome the maximum total pressure loss in an y of the branch tak eoffs including VAV boxes, f exible ducts, diffusers, and dynamic losses under all operating conditions. The duct static pressure is more stable than the duct total pressure in a supply main duct when the airstream is f owing. More often, the set point of the duct static pressure in VAV systems ranges between 0.8 and 1.5 in. WG (200 and 375 Pag) The static pressure sensor should be located in the supply main duct in a position such that the required supply volume f ow rate is guaranteed for all VAV boxes with greatest energy savings during part load. The nearer to the f arthest branch takeoff, the greater the ener gy saving because pf x is nearly minimum. F or most VAV systems, the duct pressure sensor is often located in the main duct near the farthest branch takeoff and at a location where steady static pressure can be measured. Duct static pressure control is discussed also in Sec. 29.5. Warren and Norford (1993) recommended that the duct static pressure set point be reset based on messages from each VAV box collected by a poll manager control block during operation. Such a DDC control system is complicated and expensive. ASHRAE / IESNA Standard 90.1-1999 speci f es that duct static pressure sensors used to control variable-air-volume fan operating characteristics shall be placed in a position such that the controller set point is no greater than one-third of the total design f an (supply) static pressure e xcept for direct digital control systems with zone reset capability where the sensor may be located at the f an discharge. For a supply duct system with major duct splits, multiple sensors should be installed in each major branch to ensure static pressure can be maintained in each of the major branches. For air systems with direct digital control of indi vidual zone VAV boxes connecting to the central system controller, the duct static pressure set point shall be reset based on the zone requiring the most pressure, that is, the set point is reset lower until one zone damper is nearly wide open.

Comparison between Adjustable-Frequency Drives and Inlet Vanes SH__ ST__ LG__ DF

In VAV systems, both adjustable-frequency variable-speed drives and inlet vanes are used to modulate the volume f ow and fan total pressure of the centrifugal supply, return, relief, and exhaust fans, as described in Sec. 15.4. Generally , fans using adjustable-frequenc y variable-speed drive to v ary

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the f an speed are more ener gy-eff cient at reduced v olume f ow than are positioning inlet v anes. However, adjustable-frequency variable-speed drive is more expensive than inlet vanes. Lorenzetti and Norford (1992) reported that adjustable-frequenc y v ariable-speed dri ves consumed 46 to 66 percent less supply fan energy than inlet vanes, based on the measured performance of four AHUs during a retrof t project in two connected buildings. The brake horsepower of the supply fan ranged from 21 to 36 hp. Only one of the AHUs operated 24 h/day, while the rest operated 8 to 15 h /day. From the results of ener gy simulation programs, other researchers also found that adjustable-frequency variable-speed drives have an energy savings between 20 and 50 percent over inlet vanes depending on the type and size of f an, annual operating hours, and characteristics and location of the building. Field measurements also sho wed that man y fans installed with inlet v anes saved only 20 to 30 percent, even as little as 10 percent, when the v olume f ow rate w as reduced to 0.3 of its design value, if inlet v anes are mounted too f ar away from the impeller . Theoretically, the energy savings should be greater than 50 percent. Figure 23.9 compares the f an power consumption at fractions of the design v olume f ow r ate when adjustable-frequenc y v ariable-speed dri ves and inlet v anes are used. Adjustable-frequency drives are often cost-effective under the following conditions:

FIGURE 23.9 Comparison of fan power consumption between adjustable-frequency, variable-speed drives and inlet vanes.

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A variable supply volume f ow is required during operation. The fan brake horsepower is 25 hp (19 kW) and greater. Backward-inclined and airfoil fans are the best candidates. The annual operating hours exceed 2500 h.

Mays (1998) suggested that the most common adjustable-frequenc y drive problem today is ref ected w ave v oltages, which create tw o to three times o vervoltage at the motor terminals with pulse-width-modulated (PWM) dri ves as the rise time of the pulses becomes shorter . These cause noise in motors. An external low-pass f lter can be installed, or apply series reactors between the PWM output terminals and the cable to the motor to slow the rise time or to minimize overvoltages. The shorter the cable between the adjustable-frequenc y drive output and the motor , the lower will be the peak voltage. For a vane-axial supply f an, blade pitch controls are often used at reduced v olume f ow for f an energy sa vings. Controllable blade pitch v ane-axial f ans ha ve f an ener gy sa vings characteristics comparable to those of adjustable-frequency variable-speed drives. But they are also expensive.

23.9 FUNCTIONAL CONTROLS FOR VAV SYSTEMS Nighttime Setback, Warm-Up, or Cool-Down Control Many VAV systems are shut down during the nighttime unoccupied period. Only the heating system or the fan-powered VAV boxes are operated to maintain a nighttime setback temperature set point in order to sa ve ener gy. The purpose of w arm-up control is to reco ver the setback temperature and maintain an indoor environment between 70 and 72°F (21.1 and 22.2°C) during the occupied period in winter. The purpose of cool-do wn control is to cool and dehumidify the space air do wn to predetermined limits following an increase in the space air temperature and humidity when the air system is turned off during the unoccupied period in locations where the outdoor climate is hot and humid in summer. When the supply fan of the AHU or PU is turned on, if the difference between the space temperature Tr sensed by the temperature sensor and the set point Tsp (Tr  Tsp) is less than or equal to a predetermined limit, such as 3 °F (1.7 °C), then the air system enters the w arm-up mode; and if (Tr  Tsp)  a predetermined limit, such as 3°F (1.7°C), the air system enters the cool-down mode. Both the warm-up and cool-down controls close the minimum outdoor , economizer, and relief / exhaust dampers and open to 100 percent the recirculating damper . In addition, the warm-up control opens fully the two-way valve of the heating coil or the gas valve of the furnace, and the cool-down control opens fully the tw o-way valve of the cooling coil or starts all the refrigeration compressors that serve the DX coil, until the space temperature drops within predetermined limits. One problem in warm-up or cool-down control is to determine the optimum start time. The optimum start time is affected by the following factors: ●

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●

SH__ ST__ LG__ DF

Space temperature and relative humidity just prior to the warm-up or cool-down period Construction of the building envelope Capacity of the heating / cooling equipment Internal loads Outdoor temperature

These factors are used to calculate the thermal characteristics of a conditioned space serv ed by an AHU or PU, and start the equipment prior to occupanc y to reach the desirable space temperature. There are softw are packages a vailable to calculate separate sets of heating and cooling rates for

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spaces that ha ve been unoccupied less than or greater than 24 h and pro vide an optimum start or stop time of the air system. A simple and ef fective way that is w orthwhile to try is as follo ws: Record the outdoor air temperature, space temperature, relative humidity, internal load, and equipment and damper operating conditions at suitable time interv als, say, 10 min to 1 h, during the w arm-up or cool-do wn period; analyze and determine the optimum start time once suf f cient data have been accumulated. Modify the results until they are verif ed by experience. Some manuf acturers also pro vide c ycling capacity during morning w arm-up to maintain the space temperature reached before the occupied period. If both cool-down and purge mode or warm-up and purge mode operations are required, and the outdoor air is signi f cantly hot and humid, or colder than the space air during the occupied period, then the purge operation must be performed f rst, and the cool-down or warm-up operation follows until the space temperature is maintained within predetermined limits. During purge mode, the minimum outdoor air, economizer, and exhaust dampers are opened 100 percent, and the recirculating damper is closed. During w arm-up or cool-down mode, the minimum outdoor air, economizer, and exhaust dampers are closed, and the recirculating damper is opened 100 percent. Only when the outdoor air can be used for free cooling, and to squeeze out the contaminated indoor air at the same time, are the cool-do wn and pur ge operations combined into a single operation.

Steam Humidifier Control A steam humidif er is the most widely used humidi f er when a warm air supply is required in winter in an AHU or PU. A humidity sensor is often located in the supply duct just after the supply fan discharge, to sense the relati ve humidity of the dischar ge air. The DDC controller recei ves the sensed signal and positions the two-way valve to modulate the steam f ow rate entering the steam humidif er, as shown in Figs. 15.42 and 15.43, until the relative humidity of the discharge air meets the required set point. When the fan is turned on, the humidif er control is energized. The steam valve closes when the fan is turned off. If a steam humidi f er is installed, it is critical that the humidifying capacity not be e xceeded, so that a wet surf ace does not occur in the AHU as well as in the supply duct in order to maintain an acceptable indoor air quality.

Dew Point Control For air systems serving conditioned space that needed both close control of the space temperature and relative humidity with a sensible ratio of the space conditioning line SHR s above 0.9, a dew point control is often used. De w point control involves using a water cooling coil, DX coil, or air washer to cause a specif ic amount of conditioned air to cool and dehumidify or occasionally cool and humidify to point cc, such as that on Fig. 23.10. Because of the high relati ve humidity of the departing air, the temperature Tcc is approximately equal to the dew point on the saturation curve. When the indoor space temperature is to be maintained at 75 °F (23.9°C), with a relative humidity of 50 percent, Tcc  55°F (12.8°C), cc  97 percent, and SHR s  0.95, the state point of the conditioned space r must move approximately along the horizontal dew point line. Either the sensible cooling load is absorbed by the supply air or reheating is pro vided to the supply air by a reheating coil to compensate the sensible load at part load, if the space temperature is maintained at 75 °F (23.9°C), the relative humidity must be at 50 percent. A temperature sensor senses the space temperature Tr, and a DDC controller modulates the capacity of the reheating coil according to the sensed space temperature to maintain Tr  75°F (23.9°C) with cc  50 percent. Dew point control is simple and ef fective. Its main dra wback is its need for reheating which wastes energy. Dew point control is mainly used for industrial applications such as clean rooms.

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FIGURE 23.10

Dew point control.

Diagnostics The system diagnostics automatically monitors the operation of all w orkstations, modems, LAN connections, control panels, and controllers. The failure of any device should be displayed in a diagnostic display module for which a le gend is pro vided. In a recent model by a manuf acturer, 49 different diagnostics can be read at the display module, and the last 20 diagnostics can be held in an active history buffer log in the module. The diagnostic module is also capable of displaying outputs of microprocessor-controlled run tests to v erify operation of e very thermistor, potentiometer, fan, and compressor before the air system is started. A smoke detector is usually located in the return air passage. Once the smoke detector detects smoke in the return airstream, it sends a signal to the DDC controller and ener gizes the f re alarm system, as shown in Fig. 21.3b. Smoke controls and stairwell pressurization systems are discussed in Chap. 22. A pressure sensor is usually used to sense the pressure drop across the f lter. As soon as the pressure drop exceeds a predetermined limit, an alarm signal appears at the display module to indicate the need for replacement or maintenance. Safety controls are often part of the system components in HV AC&R equipment, as discussed previously.

23.10 RECOMMENDATIONS AND INTERACTION BETWEEN CONTROLS Recommendations for VAV Controls ●

SH__ ST__ LG__ DF

●

Equipment, control v alves, and air dampers should be properly selected and sized. Ov ersized equipment and components degrade the control quality. Proportional-integral (PI) control mode is the widely used control mode in HV AC&R applications. PI control eliminates the offset, resulting in a better indoor environment and a lower energy

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consumption than proportional control mode. A wider proportional band and properly selected proportional and integral gains for the controller provide a stable and effective operation. Cold discharge air temperature reset for the perimeter zone is ener gy-eff cient when PI or proportional-integral-derivative (PID) control mode is used. ●

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The use of an air economizer c ycle is ener gy-eff cient and pro vides adequate outdoor v entilation air for occupants. A sensor should be located where the v alue it senses represents the controlled v ariable of the whole zone to be controlled. A sensor should be exposed and located in a place that has suf f cient air movement. However, it should not be directly under a supply air jet, in which the air temperature and relati ve humidity are quite dif ferent from those in the occupied zone. It also should not be located in a stratif ed airstream. To measure the duct static pressure, if a long section of straight duct is a vailable (say, a length of duct section greater than or equal to 10 duct diameters), a single point, Pitot-tube type of duct static pressure sensor can be used. Otherwise, a multipoint Pitot-tube array of f ow-measuring station with airf ow straighteners should be used for v elocity pressure or static pressure measurement. The small holes used to measure static pressure in a Pitot tube should ne ver be directly opposite an airstream with a v elocity pressure that can af fect the reading. A reference pressure should be picked up at a point with low air velocity outside the duct, at a point served by the same air system, or in the ceiling plenum. A space pressure sensor should be located in an open area of the conditioned space where the air velocity is less than 40 fpm (0.2 m / s) and where its reading is not af fected by the opening of doors. The reference pressure pickup is best located at the rooftop, at a level 10 ft (3.0 m) abo ve the building and shielded from the inf uence of wind gusts. If a simpler control system can do the same job as a more comple x system, the simpler system is always the f rst choice.

A suf f ciently clear , well-followed operations manual and a well-implemented maintenance schedule are key factors for an effective control system.

Interaction between Controls In a typical multizone VAV system, each of the zone temperature control, discharge air temperature control, minimum ventilation control, duct static pressure control, space pressurization control, and warm-up or cool-down controls may consist of a single closed control loop, or several closed control loops combined together. The interaction between these control loops can be cate gorized as follows: ●

●

Sequence contr ol. F or this kind of control action, several indi vidual control loops operate in sequence in order to maintain a particular controlled v ariable within predetermined limits to pre vent simultaneous cooling and heating. For example, the discharge air temperature control in an AHU or PU consists of an air economizer control loop as the f rst-stage cooling, a cooling coil control loop, a mixed air temperature control loop, and a heating coil control loop in sequence. The two cooling control loops can be actuated simultaneously , whereas the cooling coil, mixed air, and heating coil control loops must be operated in sequence to prevent simultaneous cooling and heating. Override. If two control loops operate on the same control de vice, one control loop must override the control action of the other control loop. One must have a priority to actuate the control device. The predetermined precedence of one control signal o ver another in the DDC controller is called override. F or e xample, the minimum v entilation control and the air economizer control both actuate the minimum outdoor air damper , economizer dampers, and recirculating dampers. When the DDC controller fully opens the economizer and minimum outdoor dampers, the air economizer control o verrides the minimum v entilation control because the outdoor air intak e during the 100 percent air economizer c ycle is f ar greater than the minimum outdoor air intak e during minimum ventilation control.

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Control loops are often operated independently of one another. For example, the cooling coil control loop that modulates the tw o-way valve of the chilled w ater supply is operated independently of the space pressurization control loop that modulates the speed of the return f an to maintain the space pressure within preset limits. Control actions af fect each other. The control action of one control loop somtimes may af fect the controlled variable of another control loop. For an example, the closing of the damper in a VAV box of a zone temperature control loop reduces its v olume f ow and raises the static pressure of the duct static pressure control loop. As the speed of the supply f an is reduced, the supply volume f ow rate and the fan total pressure are reduced accordingly. The duct static pressure is then maintained within predetermined limits, and a new system balance is formed.

Stability Problems ●

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In VAV controls, instability may occur as the result of hunting because of a smaller proportional band and a fast opening-closing controlled device. Avery (1992) noted that interaction problems may occur because the strok e speeds of the v alve actuator or damper actuator may be dif ferent, or the sensor response times may not be the same for system components from different manufacturers. Control loop interaction between pneumatic operated VAV boxes and the electric inlet vane actuators of the supply fan can be serious as the slow-operating electric inlet vane actuator cannot keep pace with the changes in duct static pressure caused by fast-acting pneumatic actuators. Control loops ha ving control actions af fecting each other sometimes may destabilize the system.

23.11 SEQUENCE OF OPERATIONS OF A VAV REHEAT SYSTEM WITH MINIMUM VENTILATION, DISCHARGE AIR TEMPERATURE, AND DUCT STATIC PRESSURE CONTROLS — CASE STUDY HVAC&R System Consider a multizone VAV reheat system using an AHU to serve a typical f oor in an off ce building with perimeter and interior zones. The summer and winter design conditions are similar to those for a VAV reheat system described in Sec. 21.3, except the winter design temperature is 30°F ( 1.1°C) with a design outdoor humidity ratio of 0.003 lb / lb (0.003 kg / kg). Humidif cation is not required in the AHU during the winter heating mode. In minimum v entilation control, the sequence of operations of both CO 2-based demand-controlled v entilation and the alternati ve mix ed plenum pressure control at the AHU intak e are included. The discharge air temperature control consists of a temperature air economizer c ycle, a water cooling coil control loop, a mix ed air temperature control loop, and a w ater heating coil control loop. The set point that changes o ver from minimum outdoor air recirculating mode to 100 percent outdoor air free-cooling mode is 75 °F (23.9°C). There is a minimum outdoor damper , an economizer damper, a recirculating damper, and an exhaust damper. In duct static pressure control, supply and return f an pressure and v olume f ows are modulated by varying the speeds of the supply and return fans. Sequence of Operations The recommended year-round sequence of operations of this VAV reheat system is as follows: SH__ ST__ LG__ DF

1. During the unoccupied period, the time-of-day clock shuts do wn the AHU and its associated supply and return f ans. Outdoor minimum damper D1, economizer damper D2, recirculating

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damper D3, exhaust damper D4, cooling coil v alve V1, and heating coil v alve V2 are all closed. Supply and return fan motors are off. 2. During the unoccupied period, an outdoor air temperature sensor (temperature transmitter) T1 measures the outdoor temperature To, and an indoor temperature sensor TT11 measures the indoor temperature Tr. Also an outdoor CO 2 sensor CO 2  1 measures the outdoor CO 2 concentration CO2o, and an indoor CO 2 sensor CO 2  2 measures the concentrations of a representati ve f oor space CO2r. In hot summer , if the summer indoor design temperature Tr,des  75°F (23.9 °C) and the measured representative zone temperature Trxn  3°F, a cool-down mode is required. If the measured indoor-outdoor CO 2 concentration dif ference (CO 2r  CO2o  200 to 300 ppm, a pur ge mode is also required. If both cool-do wn and pur ge modes are required and if To  78°F (25.6 °C), the purge mode must be performed f rst, then the cool-down mode. ●

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Based on the space temperature, space relative humidity, outdoor temperature, and temperature of representative structural member, or according to the computer softw are, the DDC unit controller starts supply and return f ans at very low speed, such as 50 rpm through the adjustable-frequenc y variable-speed drives. The minimum outdoor damper , economizer damper, and exhaust damper are then opened fully , whereas the recirculating damper remains closed. The speeds of supply and return fans are gradually increased to design speeds. Once the measured CO 2r  CO2o 100 ppm, the purge mode is completed, and the DDC controller closes the minimum outdoor, economizer, and exhaust dampers, and opens the recirulating damper 100 percent. As To  75°F (23.9°C), the DDC controller sets the dischar ge air temperature set point at 52 °F (11.1°C) and opens 100 percent the cooling coil valve. When the difference between the measured zone temperature and the summer indoor design temperature Tr  75 1°F (Tr  23.9 0.56°C), the cool-down mode is completed.

If both cool-do wn and pur ge modes are required, and if the outdoor temperature 52 °F  To  75°F (11.1 °C  To  23.9°C), the DDC controller opens minimum outdoor , economizer, and exhaust dampers all to 100 percent, closes the recirculating damper , and operates the air economizer cycle in 100 percent outdoor air free-cooling mode. At the same time, it purges the conditioned space with outdoor air . When Tr  75 1°F (Tr  23.9 0.56°C), the cool-down mode is completed; and when CO2r  CO2o 100 ppm, the purge mode is completed. 3. As the temperature sensor measures the representati ve zone temperature Trxn  75°F (23.9°C), the DDC controller calls for cooling. And if the outdoor temperature To  75°F (23.9°C), the VAV system is in minimum outdoor air recirculating mode, as shown in Fig. 23.7. Based on the dischar ge air temperature reset schedule for cooling listed belo w, as the outdoor temperature To  75°F (23.9 °C), the DDC controller resets the dischar ge air temperature Tdis  52°F (11.1°C). To, °F

Tr, °F

r, %

Tcc, °F

Tmix, °F

Tdis, °F

Ts, °F

Tr  Ts, °F

76 – 95 59 – 75  58

75 75 75 or less

42 – 45 43 – 46 48 – 51

50 53 56

50 – 80 53 – 78 56 – 78

52 55 58

55 59 63

20 16 12

Here Tr indicates zone temperature, r zone relati ve humidity , Tcc off-coil temperature, Tmix mixed air temperature, Tdis discharge air temperature, and Ts supply temperature. For a VAV system using the same AHU that serv es both perimeter and interior zones, Tdis only resets to 55 °F (12.8°C) when outdoor temperature To drops belows 75° F (23.9°C) because a high space sensi-

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ble cooling load in the interior zone during part-load operation in an of fice building needs cold air supply. 4. During minimum outdoor air recirculating mode, when the sensed zone temperature Trn exceeds 75 °F (23.9 °C), DDC terminal controllers modulate the opening of single-blade dampers of the reheating VAV boxes in the perimeter zones and VAV boxes in the interior zones, to increase the zone, volume f ow of supply air to maintain a nearly constant zone temperature Trn  75°F (23.9°C). During part-load operation, when Trn drops below 75°F (23.9°C), because of a reduction in zone sensible cooling loads, DDC terminal controllers modulate the opening of single-blade dampers of the reheating VAV boxes in the perimeter zones and VAV boxes in the interior zones, to reduce the v olume f ow of the zone supply air to maintain a nearly constant zone temperature of 75°F (23.9°C), as discussed in Sec. 21.3. 5. During minimum outdoor air recirculating mode part-load operation, when reheating VAV boxes and VAV box es close their single-blade dampers to smaller openings, the pressure drop across the dampers increases. This results in a higher static pressure in the supply main duct. As such, an increase in duct static pressure is sensed by a pressure sensor located at a distance of 0.3Lmain from the farthest branch takeoff in the supply main duct, as discussed in Sec. 23.8; a DDC controller modulates and reduces the speed of the supply f an through an adjustable-frequenc y variable-speed dri ve to maintain a nearly constant duct static pressure set point, such as 1.2 in. WG (300 Pag), at the location of the duct static pressure sensor. Therefore, the volume f ow rate of the supply f an is reduced to balance the reduction of zone supply v olume f ow rate during part load. 6. During the minimum outdoor air recirculating mode, if CO2-based demand-controlled ventilation is used, the relationship between the supply f ow density of minimum outdoor air (as well as the corresponding minimum outdoor damper opening) and the CO 2 concentration in the return duct (refer to Fig. 23.2) has already been determined during commissioning. Since as the supply f an is modulated at a speed to maintain a preset duct static pressure of 1.2 in. WG (300 P ag), the DDC controller opens the minimum outdoor damper to a setting to pro vide base ventilation V˙ bv f rst. According to the measured CO 2 concentration in the return duct, the DDC controller then resets the minimum outdoor damper opening setting according to the determined relationship at a predetermined time interval (such as 2 and 5 min). Measured CO2 concentration, ppm  800  400

SH__ ST__ LG__ DF

Required minimum outdoor air, cfm / ft2 (L / s m2) V·o,sys  0.14 (0.7) V·  0.07 (0.35) bv

The recirculating damper opens 100 percent, and the exhaust dampers open 30 percent. For a VAV system with a supply and return f an combination, a pressure sensor is used to measure the representative space pressure pr, in in. WG (Pag), and controls the space pressure in f oating control mode. Only when pr  0.03 in. WG ( 7.5 Pag) does the DDC controller increase the speed of the return f an, which raises the amount of e xhaust air through the e xhaust damper V˙ eu and maintains pr   0.03 in. WG (7.5 Pag). Also only when pr  0.005 in. WG will the DDC controller reduce the speed of the return f an to increase space pressure to  0.005  pr   0.03 in. WG ( 1.25  pr  7.5 Pag). When the measured return duct CO 2 concentration is  800 ppm, the minimum outdoor air intake for a typical of f ce building should be 0.14 cfm / ft2 (0.7 L / s m2). Minus the e xhaust through the restroom of 0.03 cfm / ft2 (0.15 L / sm2), the remaining part is 0.14  0.03  0.11 cfm / ft2 (0.55 L / s m2). Such an amount of outdoor air is suf f cient to maintain a space pressure of  0.005  pr   0.03 in. WG ( 0.005  pr   7.5 Pag) and an associated exf ltration from an average leaky building as well as an amount of 0.015 cfm / ft2 (L / sm2) of e xhaust air through the e xhaust damper.

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Similarly, when the measured CO 2 concentration is  400 ppm, the minimum outdoor intak e V˙o,min  V˙bv  0.07 cfm / ft2 (0.35 L / sm2), and the remaining part is 0.07  0.03  0.04 cfm / ft2 (0.2 L / sm2) which can provide a space pressure around  0.005 in. WG (1.25 Pag). 7. During the minimum outdoor air recirculating mode, if a mix ed plenum pressure control is used, either a pressure sensor located in the mix ed plenum measures the static pressure dif ference between the mix ed plenum and outdoors pm,p, as shown in Fig. 23.4 a, or a pressure sensor measures the pressure drop of the outdoor air damper and the louver, as shown in Fig. 23.4b. If the measured static pressure difference or pressure drop decreases and is less than a set point, such as  0.4 to  0.6 in. WC (  100 to  150 P a), the DDC controller closes the recirculating damper until pm,p equals the set point and an amount of minimum outdoor air of V˙ o,sys  0.14 cfm / ft2 (0.7 L / s  m2) is extracted in the AHU. The minimum outdoor air damper is opened a f xed amount, such as 100 percent. Mixed plenum pressure control provides an approximately constant minimum outdoor air with an amount of V˙ o,sys. During minimum outdoor air recirculating mode, for VAV systems with a supply and return f an combination, the minimum outdoor damper opens 100 percent, the exhaust damper opens 30 percent, and the economizer damper is closed. A space pressure sensor measures the representative space pressure pr, in in. WG (Pag). If it exceeds  0.03 in. WG (7.5 Pag), a DDC controller increases the speed of the return f an through a v ariable-speed drive, and thus reduces the amount of minimum outdoor air intak e as well as the space pressure and maintains pr at less than  0.03 in. WG (7.5 Pag). When the zone load reduces, the speed and v olume f ow of the supply f an reduce accordingly. As the measured pr drops below  0.005 in. WG (1.25 Pag) because the return fan has not followed the supply f an to reduce its v olume f ow appropriately, the DDC controller then reduces the speed of the return fan, and therefore space pressure pr increases and is maintained at a value not less than  0.005 in. WG (1.25 Pag), as discussed in Sec. 23.6. 8. When the outdoor temperature 55 °F  To  75°F (12.8°C  To  23.9°C), the AHU is operated in 100 percent outdoor air economizer c ycle, as shown in Fig. 23.7 b. The DDC controller opens 100 percent the minimum outdoor air damper , economizer damper, and exhaust damper. The recirculating damper is closed. When 100 percent outdoor air is used for free cooling, it overrides the minimum ventilation control. The cooling and dehumidifying capacity of the chilled w ater cooling coil is also required at the same time. The DDC controller modulates the opening of the tw o-way control valve of the cooling coil to maintain a discharge air temperature Tdis  55°F (12.8°C). 9. When the outdoor temperature To,mix  To  55°F (12.8°C), the AHU is operated in the air economizer cycle with mix ed air control. Here To,mix is the outdoor air temperature at which the AHU or PU starts the mix ed air control (such as To,mix  30°F, or  1.1°C) to provide a mixture of outdoor air and recirculating air at dischar ge air temperature Tdis directly, as discussed in Sec. 23.7. (Refer to Fig. 23.7.) So To,mix depends on the percentage of outdoor air among the supply air and on the load ratio. In an air economizer cycle mixed air mode, a DDC controller opens 100 percent the minimum outdoor air damper and e xhaust damper and modulates the economizer damper and recirculating damper to maintain a mix ed air temperature equal to Tdis  55°F (12.8°C) directly. Since the amount of outdoor air intake in the AHU during mixed air mode is always greater than the amount of minimum outdoor air V˙ o,sys in recirculating mode, the mixed air control overrides the minimum ventilation control. To provide cold air supply for the interior zone e ven in winter, the DDC controller resets Tdis  55°F (12.8°C). 10. In the perimeter zone ●

If the perimeter zone temperature Trxn and interior zone temperature Trin drop belo w 75 °F (23.9°C), the terminal controller reduces the opening of the single-blade damper and thus the sup-

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ply volume f ow rate of cold conditioned air to the control zone and maintains a nearly constant zone temperature Trxn  Trin 75°F (23.9°C). Further drop in the perimeter zone space sensible cooling load causes the damper in the reheating VAV box to close to a smaller f ow passage until it reaches its minimum setting, such as 30 percent of the zone peak f ow. When the temperature of perimeter zone n is 72 °F  Trxn  75°F (22.2 °C  Trxn  23.9°C), zone n is operated in the deadband mode, as discussed in Sec. 21.3. The damper in the reheating VAV box closes to its minimum setting, and neither cooling nor heating is pro vided except minimum ventilation air. When the temperature of the perimeter zone n is Trxn  72°F (22.2 °C), the zone temperature control is operated in heating mode operation. As discussed in Sec. 21.3, the DDC terminal controller actuates the re verse-action relay, and the cold primary air supply is still in minimum setting to pro vide minimum v entilation. Also the terminal controller modulates the opening of the tw o-way v alve of the heating coil to maintain a preset zone temperature Trxn  72°F (22.2°C). The zone relative humidity may be between 25 and 50 percent. At winter design conditions, the warm air supply temperature should not e xceed 87°F (30.6°C) in order to pre vent the buoyancy effect.

Primary Considerations ●

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The sequence of operations of a VAV reheat system can be di vided into two categories: sequence of operations for AHU or PU controls and for zone controls (including perimeter and interior zones). Sequence of operations for AHU or PU is usually varied based on outdoor temperature To and can be subdivided into Minimum outdoor air recirculating mode, cooling Air economizer cycle, 100 percent outdoor air free-cooling Air economizer cycle, mixed air control Minimum outdoor air recirculating mode, heating Sequence of operations for zone control is v aried based on zone temperatures: perimeter zone temperature Trxn and interior zone temperature Trin. Zone temperature control can be subdi vided into cooling, deadband, and heating mode operations. Minimum v entilation control is only v alid to pro vide v entilation during minimum outdoor air recirculating mode between base v entilation V˙ bv  0.07 cfm / ft2 (0.35 L / sm2) and minimum outdoor air speci f ed by ASHRAE Standard 62 – 1999, V˙ o,sys  0.14 cfm / ft2 (0.7 L / sm2) for a typical off ce building. During air economizer mode, outdoor air intak e in the AHU or PU is always greater than V˙ o,sys; there is no need of minimum ventilation control. For a VAV system with supply and relief f an combination operated in minimum outdoor air recirculating mode, if minimum outdoor air is pro vided by CO 2-based demand-controlled v entilation or mixed plenum pressure control, the space pressure is maintained between  0.005 and  0.03 in. WG (1.25 and 7.5 Pag) for a properly designed average leaky off ce building. For a VAV system with supply and return fan combination, when the speed of the supply fan is reduced during part-load operation, a DDC controller should reduce the speed of the return f an to maintain a space pressure between  0.005 and  0.03 in. WG (1.25 and 7.5 Pag). If a humidif ier is used, be sure that the humidifying capacity is ef fectively controlled and the wet surf ace doesn ’t occur inside the AHU, PU, or supply ducts, to create indoor air quality problems.

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REFERENCES ASHRAE, ASHRAE Handbook 1996, HVAC Systems and Equipment, ASHRAE Inc., Atlanta, GA, 1996. ASHRAE, ASHRAE Handbook 1999, HVAC Applications, Atlanta, GA 1999. ASHRAE, ASHRAE Standard 62 – 1999, Ventilation for Acceptable Indoor Air Quality, Atlanta, GA 1990. Avery, G., The Instability of VAV Systems, Heating / Piping / Air Conditioning, February 1992, pp. 47 – 50. Avery, G., Selecting and Controlling Economizer Dampers, HPAC, no. 8, 1996, pp. 73 – 78. Becelaere, R. V., Mixing Box Damper Testing, ASHRAE Transactions 1998, Part II, pp. 1226 – 1231. Carpenter, S. C., Energy and IAQ Impacts of CO2-Based Demand-Controlled Ventilation, ASHRAE Transactions, 1996, Part II, pp. 80 – 88. Chen, S. Y. S., and Demster, S. J., Variable Air Volume Systems for Environmental Quality, McGraw-Hill, New York, 1995. Chen, S. Y. S., Yu, H. C., and Hwang, D. D. W., Ventilation Analysis for a VAV System, Heating / Piping / Air Conditioning, April 1992, pp. 36 – 41. Ellis, R., and McKew, H., Back to Basics, Test 27 — Central Air with Return Air System Advanced Energy-Eff ciency — Design, Engineered Systems, no. 8, 1997a, pp. 9– 10. Ellis, R., and McKew, H., Back to Basics, Test 29 — Central Air with Return Air System Building Automation — Design, Engineered Systems, no. 9, 1997b, pp. 11 – 12. Elovitz, K. M., Variable-Speed Drives: What They Do and When to Use Them, Heaitng / Piping / Air Conditioning, no. 12, 1993, pp. 67 – 75. Elovitz, D. M., Minimum Outside Air Control Methods for VAV Systems, ASHRAE Transactions, 1995, Part II, pp. 613 – 618. Elovitz, D. M., Minimum Outside Air Ventilation in VAV Systems, Engineered Systems, no. 3, 1997, pp. 44 – 50. Emmerich, S. J., and Persily, A. K., Literature Review on CO2-Based Demand-Controlled Ventilation, ASHRAE Transactions, 1997, Part II, pp. 229 – 243. Enermodal Engineering Ltd., An Evaluation of the Effect of CO2-Based Demand-Controlled Ventilation Strategies on Energy Use and Occupant-Source Contaminant Concentrations, ASHRAE Research Project 740-TRP, Waterloo, Ontario, Canada, 1995. Englander, S. L., and Norford, L. K., Saving Fan Energy in VAV Systems — Part 1: Analysis of a VariableSpeed-Drive Retrof t, ASHRAE Transactions, 1992, Part I, pp. 3– 18. Graves, L. R., VAV Mixed Air Plenum Pressure Control, Heating / Piping / Air Conditioning, no. 8, 1995, pp. 53 – 55. Janssen, J. E., Ventilation for Acceptable Indoor Air Quality, ASHRAE Journal, October 1989, pp. 40 – 46. Janu, G. J., Wenger, J. D., and Nesler, C. G., Strategies for Outdoor Airf ow Control from a System Perspective, ASHRAE Transactions, 1995, Part II, pp. 631 – 643. Kettler, J. P., Controlling Minimum Ventilation Volume in VAV Systems, ASHRAE Journal, no. 5, 1998, pp. 45 – 50. Lo, L., VAV System with Inverter-Driven AHU for High-Rise Off ce Building in Tropical Climates — A Case Study, ASHRAE Transactions, 1990, Part I, pp. 1209 – 1217. Lorenzetti, D. M., and Norford, L. K., Measured Energy Consumption of Variable-Air-Volume Fans under Inlet Vanes and Variable-Speed-Drive Control, ASHRAE Transactions, 1992, Part II, pp. 371 – 379. Mays, M., Identifying Noise Problems in Adjustable Speed Drives, ASHRAE Journal, no. 10, 1998, pp. 57 – 60. Mumma, S. A., and Wong, Y. M., Analytical Evaluation of Outdoor Airf ow Rate Variation vs. Supply Airf ow Rate Variation in Variable-Air-Volume Systems When the Outdoor Damper Position Is Fixed, ASHRAE Transactions, 1990, Part I, pp. 1197-1208. Nabinger, S. J., Persily, A. K., and Dois, W. S., A Study of Ventilation and Carbon Dioxide in an Off ce Building, ASHRAE Transactions, 1994, Part II, pp. 1264 – 1274. Reardon, J. T., Air Change Rates and Carbon Dioxide Concentrations in a High-Rise Off ce Building, ASHRAE Transactions, 1994, Part II, pp. 1251 – 1263. Roberts, J. W., Outdoor Air and VAV Systems, ASHRAE Journal, September 1991, pp. 26–30. Robinson, K. D., Damper Control Characteristics and Mixing Effectiveness of an Air-Handling Unit Combination Mixing / Filter Box, ASHRAE Transactions, 1998, Part I A, pp. 629 – 637.

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Ruud, S. H., Fahlen, P., and Anderson, H., Demand Controlled Ventilation — Full Scale Tests in a Conference Room, Proceeding of the 12th AIVC Conference Air Movement and Ventilation Control within Buildings, Air Inf ltration and Ventilation Center, Conventry, United Kingdom, 1991, pp. 187 – 200. Stonier, R. T., CO2: Powerful IAQ Diagnostic Tool, Heating / Piping / Air Conditioning, no. 3, 1995 pp. 88 – 102. Taylor, S., Is CO2 Appropriate Indicator of IAQ? ASHRAE Journal, no. 12, 1998, p. 8. The Trane Company, Packaged Rooftop Air Conditioners (20 to 130 Tons), The Trane Company, Clarksville, TN, February 1997. Turpin, J. R., Driving into the Future, Engineered Systems, no. 6, 1997, pp. 34 – 39. Wang, S. K., Air Conditioning, vol. 4, Hong Kong Polytechnic, Hong Kong, 1987. Waren, M., and Norford, L. K., Integrating VAV Zone Requirements with Supply Fan Operation, ASHRAE Journal, no. 4, 1993, pp. 43 – 46. Williams, V. A., VAV System Interactive Controls, ASHRAE Transactions, 1988, Part I, pp. 1493 – 1499.

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24.1 IAQ PROBLEMS 24.1 24.2 VENTILATION, DILUTION OF CONCENTRATION OF CONTAMINANTS 24.2 Outdoor Air Requirements for Ventilation 24.2 Importance of Time of Operation and Ventilation Rate 24.2 Adopting an Air Economizer and Purge Operation 24.2 Select an Appropriate Minimum Ventilation Control System 24.3 Install a Minimum Outdoor Damper and an Economizer Damper 24.3 24.3 ELIMINATING MICROBIAL GROWTH 24.4 Basics 24.4 Contagious Respiratory Microorganisms 24.4 Microbial Growth 24.4 Eliminate Water Leaks, Prevent Dampened Surface and Material 24.5 Purge, Pressurization Control, and Ultraviolet Germicidal Irradiation 24.5 24.4 FILTERS TO REMOVE CONTAMINANTS 24.6 Remove Indoor Air Contaminants by Air Filters 24.6

Filter Selection for IAQ 24.6 Service Life of Air Filters 24.7 Filter Installation 24.7 24.5 REMOVING GASEOUS CONTAMINANTS BY ADSORBERS AND CHEMISORBERS 24.8 Indoor Gaseous Contaminants 24.8 Activated Carbon Adsorbers 24.9 GAC Performance 24.9 Simulating GAC Applications 24.10 Chemisorption 24.11 Chemisorber Performance 24.11 24.6 MAINTENANCE TO GUARANTEE IAQ 24.12 Inspection, Service, and Access 24.12 Monitoring of Operating Conditions in Air Systems 24.12 Coils and Ductwork 24.12 24.7 SPACE PRESSURIZATION CONTROL 24.13 24.8 LEGAL RESPONSIBILITY FOR IAQ CASES 24.13 SBS or IAQ Cases 24.13 Who Is Legally Responsible 24.13 HVAC&R Engineer 24.14 REFERENCES 24.15

24.1 IAQ PROBLEMS The sick building syndrome (SBS) received public attention from 1970s after the ener gy crisis as a result of a tighter b uilding and a reduced amount of outdoor v entilation air. Since Americans are spending more and more of their time indoors, they need a comfortable and healthy indoor environment and an acceptable indoor air quality . In an unhealthy b uilding environment, uncomfortable employees do not perform well and producti vity declines. Worker illness due to a poor indoor environment elevates absenteeism. A significant issue that is facing building owners, operating mangers, architects, consulting engineers, and contractors today is the possibility of le gal suits filed by occupants or owners who feel that their health has been damaged by poor indoor air quality. In the 1900s, indoor air quality (IAQ) has become one of the primary concerns in air conditining (HVAC&R) system design, manufacturing, installation, and operation because of the following: 24.1

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IAQ is closely related to the health of the occupants inside a b uilding, whether a b uilding is a healthy building or a sick building. IAQ and thermal control (zone temperature and relati ve humidity control) indicate primarily the quality of the indoor environment in a building. IAQ and thermal control represent mainly the functional performance of an air conditioning (HVAC&R) system.

As described in Sec. 4.10, the indoor air contaminants are the basic pollutants that af fect IAQ. According to the results of man y field i vestigations and recent IAQ conferences, the primary IAQ problems and the appropriate mitigations and impro vements are based on the three basic strate gies: contaminated source control, removing contaminants by air cleaner , and v entilation air dilution, which are summarized and discussed in the following sections.

24.2 VENTILATION, DILUTION OF CONCENTRATION OF CONTAMINANTS Outdoor Air Requirements for Ventilation Persily (1989) took field measurements of the outdoor entilation rates of 14 of fice uildings for approximately one year in 1983 and found that the y are typically between 0.6 and 1.2 ach, with a mean value of 0.94 ach. Among these 14 office uildings 52 percent had a minimum le vel of outdoor air intake that was lower than the design le vel, 45 percent had less than 20 cfm / person (10 L / sperson), 8 percent had less than 10 cfm / person (5 L / sperson), and 1 percent had less than 5 cfm / person (2.5 L / sperson). ASHRAE Standard 62-1999 specified outdoor air requirements for entilation for each occupant in various commercial facilities. Some of these requirements are listed in Table 4.5. Outdoor air requirements for ventilation per person V˙ o,req, such as 20 cfm / person (10 L / s person) in an offic building, must be pro vided at both design fl w and part-load reduced v olume fl w in VAV systems. From Eqs. (23.2) and (23.3), the calculated system outdoor air v olume fl w rate V˙ o,sys  NocV˙ o,req, in cfm (L / s). Here Noc indicates number of persons, and V˙ o,sys is the minimum outdoor air intak e at design fl w in an AHU or a PU for the dilution of the concentrations of indoor air contaminants. F or a typical office uilding with a ceiling height of 9 ft (2.7 m) V˙ o,sys is 0.93 ach, as listed in Sec. 23.2. According to Mendell (1993), a statistically significant correlation between outdoor entilation rate in offices and sick uilding syndrome complaints was found. Ventilation has been proved to be the most important factor to improve IAQ. Importance of Time of Operation and Ventilation Rate An often discussed topic is whether the type of HV AC&R system is of an y significance that a fects the aspect of v entilation and thus the SBS and IA Q. Sundell (1996), based on a study of 210 of fic buildings in northern Sweden, noted that the type of air system — whether it w as a modern ne w HVAC&R system or an old simple system — had no significance on the aspect of entilation and IAQ. Instead, the important aspects of v entilation were the time of operation and the v entilation rate. If the time of operation of the air system to provide ventilation was less than 10 h / day, the risk of SBS w as increased. As the v entilation rate cfm / person (L / s person) was reduced, the odds of SBS symptoms were raised accordingly. Adopting an Air Economizer and Purge Operation SH__ ST__ LG__ DF

Select an AHU or a PU with an air economizer and purge mode control. As discussed in Sec. 23.11, if an air economizer is used, the amount of outdoor air intake in the AHU or PU during 100 percent

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air economizer cycle and mixed air control mode is f ar greater than the calculated minimum air intake at design f ow V˙ o,sys. A 100 percent air economizer cycle signif cantly improves the IAQ. If the difference in concentration of CO 2 between the conditioned space and the outdoors during morning time unoccupied period  200 to 300 ppm, or the difference in concentration of a speci f c indoor contaminant between outdoors and indoors is signi f cantly greater than the normal operating period, a purge mode operation to exhaust all the space air with higher concentration of indoor contaminants and extract the outdoor air of lo wer concentrations of contaminants is bene f cial to lower the concentration of the indoor contaminant.

Select an Appropriate Minimum Ventilation Control System As discussed in Chap. 23, for a multizone VAV system that serves space with an obvious occupancy variation schedule and has lo w pollutants from nonoccupant-generated source, a CO 2-based demand-controlled ventilation (DCV) system should be used. In a DCV system: ●

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The amount of outdoor air intak e during minimum outdoor air recirculating mode v aries between V˙o,sys at design occupancy and the base ventilation V˙ bv when occupancy reduces. Outdoor air requirements for v entilation per person, in cfm / person (L / s person), meet the criteria specif ed in ASHRAE Standard 62-1999. Part of the cooling and heating ener gy to condition outdoor air is sa ved during reduced occupancy.

For a VAV system that serves space with a schedule of less occupancy variation, a mixed plenum pressure (MPP) minimum ventilation control system monitoring either plenum pressure or damper louver pressure drop should be used. In an MPP control system, a f xed amount of outdoor air V˙ o,sys is extracted in an AHU or a PU for ventilation at both design f ow and part-load volume f ow during minimum outdoor air recirculating mode. During commissioning, it is important to determine the relationship between the measured concentration of CO 2 in the return air duct (or on a representati ve f oor space) and the minimum outdoor air intake expressed as the damper opening setting in a CO 2-based demand-controlled ventilation system, as discussed in Sec. 23.2.

Install a Minimum Outdoor Damper and an Economizer Damper For better minimum v entilation control in a VAV system, it is recommended to install tw o outdoor dampers: a minimum outdoor damper and an economizer damper . The minimum outdoor damper is opened during minimum outdoor air recirculating mode. Its size is designed based on a v olume f ow rate of V˙ o,sys, in cfm (L / s). The economizer damper is opened fully during 100 percent air economizer cycle. Its size is based on a volume f ow rate V˙ ec,d, in cfm (L / s), and can be calculated as V˙ec,d  V˙s  V˙o,sys

(24.1)

where V˙s  supply volume f ow rate of the AHU or PU, cfm (L / s). The face velocity of either the minimum outdoor damper or the economizer damper is usually 1000 to 3000 fpm (5 to 15 m / s), and the ratio of the area of damper to the area of the duct of f ow passage Adam /Ad is often 0.5 to 0.9, as discussed in Sec. 5.7. The pressure drop of the damper when the damper is opened 100 percent pdam can be calculated by Eq. (5.11) and often has a v alue between 0.2 and 0.5 in. WC (50 and 125 Pa) for minimum outdoor damper and economizer damper . During 100 percent air economizer cycle, both the minimum outdoor damper and the economizer damper are fully opened, and the pressure drops across the dampers are equal. Outdoor air intak es should ha ve wind shields and louv ers, away from plumber stacks, and exhausts, with air f lters or cleaners if the outdoor air quality is poor , and consider pre vailing winds, as discussed in Sec. 16.2.

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24.3 ELIMINATING MICROBIAL GROWTH Basics Microorganisms or microbes include fungi, bacteria, and viruses. Fungi and some bacteria form spores. Mold is minute fungi growing on moist organic substance. According to Kowalski and Bahnfleth (1998), disease-causing microorganisms transmit respiratory irritation and other serious health pathogens via the airborne route from tw o kinds of sources: Most contagious pathogens come from human hosts, and most noncommunicable pathogens come from the en vironment — the HVAC&R and building-related environment. At the ASHRAE Journal Health Building / IAQ 1997 Conference, Dr. Bascom reported that “In the United States and Australia, asthma rates have nearly doubled in the past 15 years.”

Contagious Respiratory Microorganisms Kowalski and Bahnf eth (1998) noted that most contagious respiratory microor ganisms induce their human host to aerosolize large quantities of infectious bioaerosols by irritation, which causes cough and sneezing. A single sneeze can generate 100,000 f oating bioaerosols, and many carry viable microorganisms. A single cough typically produces only 1 percent of these bioaerosols b ut are ten times more frequent than sneezes. K eeping a distance from the infected person is a simple and useful method to reduce the risk of being infected.

Microbial Growth Spores are characteristically lar ger and more resistant to impacts that will destro y viruses and bacteria. According to Kowalski and Bahn f eth (1998), noncommunicable diseases are almost entirely due to fungal or actinomycete spores and en vironmental and agricultural bacteria. Spores form the most important group in noncommunicable diseases. Microbes are commonly present in outdoor air. Normally, indoor air spore le vels tend to be from 10 to 100 percent of outdoor spore le vels when the species mix indoors re f ects outdoors and are mostly less than 200 colon y forming units (CFU) per m3 because of transport from outdoors to indoors due to ventilation, inf ltration, and people, and because of microbial gro wth indoors. In the California Healthy Building Study from Godish (1995), a naturally ventilated building had an indoor spore le vel of about 80 percent that of outdoors, mechanical ventilated building about 65 percent, and an air conditioned building about 15 percent. Spores germinate and gro w in the presence of w ater and nutrients — dirt and debris. The k ey factor is water. Water may exist inside a building or an HVAC&R system because of the follo wing: ●

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Water leaks. Entrained water droplets. Wet coil face velocity exceeding 550 fpm (2.75 m / s). Wetted interior surface. Negative space pressure, due to improper air f ow balance, extracts hot and humid outdoor air which contacts with the indoor cold surfaces and causes surface condensation. Clogged condensate pan and drain system. Oversized humidif er or improperly designed direct evaporative cooler.

The result is a “damp b uilding” which is characterized by dampened materials, damp surf aces, mold, and microbial gro wth. Legionnella is a kind of pathogen transmitted via airborne route because the cooling tower water system is not properly treated and maintained.

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Sundell (1996) noted that man y studies ha ve demonstrated a relationship between damp b uildings and respiratory and aller gic diseases in children and adults, and estimated that approximately 20 percent of asthma in children in Sweden can be attributed to damp housing.

Eliminate Water Leaks, Prevent Dampened Surface and Material Always remember that w ater and dirt (dust and debris) are tw o necessities for microbial gro wth. Water leaks through the b uilding envelope, especially through cracks and openings around outdoor air ducts, must be prevented. Pooling of water on the roof should be prevented. In an HVAC&R system without a humidif er, the condensate f ows over part of the outer surface of a water cooling coil or DX coil and collects in the condensate pan. This is often the only wetted surface inside the system that contacts the conditioned supply air. As discussed in Secs. 15.10 and 15.11, to prevent dampened surf aces and damp material, the following is recommended: ●

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Choose a coil f ace velocity va  550 fpm (2.75 m / s) for corrugated f ns to pre vent condensate droplets f ying over. If the coil f ace has a v ertical dimension e xceeding 42 in. (1067 mm), an intermediate drain pan must be added to prevent condensate carryover. Install an individual drain pan and drain pipes for each coil. Drain pan may e xtend 1.5 to 3 times the coil depth in the direction of airf ow. Condensate should be drained freely with a condensate trap, condensate collected in the condensate pan should be completely drained, and cleaned periodically. There should be no dirt and debris remaining in the condensate pan. Prevent surface condensation especially because of the inf ltrated hot and humid outdoor air when the space pressure is negative. In case a humidif er must be installed in an air conditioning system, the following is recommended:

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For most comfort air conditioning systems, select a steam humidif er or a heating element humidif er, instead of an atomizing or a supersonic humidif er, to prevent possible microbial growth. Modulation capacity control is essential. An accuracy   5 percent relative humidity is recommended.

Purge, Pressurization Control, and Ultraviolet Germicidal Irradiation Kowalski and Bahn f eth (1998) recommended the follo wing HVAC&R measures to eliminate microbial growth and improve IAQ: ●

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If the concentration of a specif c species of disease-causing microbe in outdoor air is substantially lower than that in the indoor air, a 100 percent outdoor air purge at a f ow rate of 1 ach lowers the indoor concentration to 40 percent of the initial concentration level after 1 h of purge operation. If the purging airf ow rate is 4 ach, the indoor concentration le vel drops to a fe w percent of the initial concentration level after 1 h of purge operation. Room pressurization control is often used in biohazard f acilities and isolation rooms to pre vent migration of microbes from one room to another . Room pressure is dif f cult to maintain and control, and is also expensive. The effectiveness of ultraviolet germicidal irradiation (UVGI) to inhibit fungal growth and kill the spores depends on the air v elocity of the air f ow passing the UV GI, airf ow patterns, degree of

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maintenance, resistance of the microbes, and ambient humidity . A single-pass air f ow through a UVGI de vice may ha ve only limited ef fect. Ho wever, recirculation air f ow results in multiple exposures or chronic dosing.

24.4 FILTERS TO REMOVE CONTAMINANTS Physically, indoor airborne contaminants are in solid, liquid, or gaseous form. Solid particulate contaminants include paper and f brous particles, bacteria, fungus, spores, and viruses. Their size is between 0.003 and 100 m. Liquid contaminants include mists, water, paints, cleaning sprays, and printing inks that typically ha ve a size betweeen 1 and 50 m. Gaseous contaminants such as CO, CO2, and NO 2 and v olatile or ganic compounds are f ne particulates. Their size typically ranges from 0.003 to 0.006 m.

Remove Indoor Air Contaminants by Air Filters Basically, removing air contaminants by air f lters, as discussed in Sec. 5.14, has the following primary purposes: ●

●

●

To provide protection from respirable particulates, bioaerosols, and toxic and nuisance contaminants for occupants To protect the coils, ducts, and other air distrib ution f ow passages from accumulated dust, dirt, and microbial growth To reduce cleaning and maintenance expenses

Airborne particulate matter is typically in the size range of 0.01 to 100 m.” Dust is de f ned as “an air suspension of particles (aerosol) of any solid matter, usually with particle size less than 100 m.” However, particulates smaller than 10 m in diameter may penetrate into the upper and lo wer parts of the respiratory passage and lungs. The indoor concentration level of particulates less than 10 m must follo w National Ambient Air Quality Standards as listed in Sec. 4.10. In addition, size of the particulates or indoor contaminants is an essential f actor that af fects the ef f ciency of air f lters.

Filter Selection for IAQ As discussed in Section 15.14, ANSI/ASHRAE Standard 52.2-1999 de f nes 16 minimun ef f ciency reporting values (MERV) for coarse-, low-, medium-, and high-eff ciency air f lters based on particle size ef f ciency E1 in size range 0.3 to 1.0 m, E2 in size range 1.0 to 3.0 m, E3 in size range 3.0 to 1.0 m and arrestance ef f ciency. Filters should be selected primarily based on particle size eff ciency, the requirement of the conditioned space, pressure drop of the f lter, service life, and thelife-cycle cost analysis. Considering IA Q, Burroughs (1997) and Liu and Huza (1995) select the following air f lters to perform the specif ed duties: ●

●

SH__ ST__ LG__ DF

●

Employ medium-eff ciency air f lters MERV9 and higher with an E2 < 50 percent up to E2 > 95 percent in particle size 1 to 3 m to control respirable particulates between 2 to 8 m in size that may be breathed deeply into the lungs. Employ minimum air f ltration level of MER V7 with a particle size ef f ciency of 50 ≤ E3 < 70 percent for a size range of 3 to 10 m in an air system for the protection of the coil and air distribution system and to prevent the nutrition of microbial growth. Employ high-eff ciency air f lters MERV 13, 14, 15 with a lo w- or medium-ef f ciency pre f lter; employ ultrahigh-eff ciency air f lters, HEPA and ULPA f lters, with a medium-eff ciency pref lter to extend service life of these f lters.

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Gill (1996) recognized that the use of a single low-eff ciency f lter in an AHU or PU has given way to a f lter assembly with a pre f lter and high-ef f ciency f lter or at least medium-ef f ciency f lters. Kowalski and Bahnf eth (1998) recommended that HEPA f lters be used to remove microbes from the recirculating air. In a recirculating air system with an airf ow rate of 1 ach, the microbial concentration level drops to 40 percent of its initial level after 1 h of operation. If the airf ow rate is 4 ach, the microbial concentration drops to only few percent of the initial concentration level after 1 h of operation.

Service Life of Air Filters When the media used in an air f lter to remo ve air contaminants are clean, they are lo wer in ef f ciency. As the f lter is loaded with more dirt, its eff ciency gradually increases to a maximum. Then some particulates may ha ve migrated through the media and be unloading back into the airstream for low- and medium-ef f ciency f lters. At the same time, the pressure drop across the dirt-loaded air f lter increases to a limit, and the amount of air passing through the air f lter reduces to a point such that the f lter doesn’t pass through suff cient air for cooling or heating. This is often the end of the service life of the air f lter, and it should be replaced by a ne w one. It is dif f cult to choose an optimum changing point between the increase of eff ciency and the loss of airf ow. Ottney (1993) recommended the follo wing service life of air f lters as a reference for designers and operators.

Filter type

Typical change time

Max. pressure drop across air f lter, in. WC (Pa)

Flat panel 2-in. (50-mm) pleated 4-in. (100-mm) pleated 6- to 12-in. (150 to 300-mm) cartridge 21- to 36-in. bags HEPA

30 to 60 days 3 to 6 months 10 to 14 months 12 to 18 months* 12 to 24 months* 1 to 5 years*

0.5 (125) 0.9 (225) 0.9 (225) 1.5 (375) 1.5 (375) 2.0 (500)

*

With pref lters

The following are useful measures to approach an optimum service life: ●

●

●

As described in Sec. 23.9, a pressure sensor is used to measure the pressure drop pf l, in in. WC (Pa), across each air f lter. As soon as pf l exceeds a predetermined limit, an alarm signal appears on the display module asking for replacement or maintenance. F or a VAV system, pf l should be measured when the air f ow is approximately equal to the design air f ow (such as the air f ow during morning cool-down or purge mode operation). Sometimes the specif ed f nal pressure drop never reaches up because of (1) the blowout of the f lters, especially holes de veloped in bag f lters, (2) leakages from the f lter hardwares, (3) leakages from the suction side of the supply fan, and (4) possible pressure sensor defects. Plan a preventive maintenance schedule, and change the air f lters according to the planned schedule. Do not change the air f lter just because it looks dirty on the surf ace. For panel f lters, the f lter should be replaced when the air leaving side shows dirt.

Filter Installation In Ottney (1993), the following is emphasized: ●

Air f lters should be installed in gask eted channels. A fabric pile or polyprop ylene-f nned gasket material provides a better seal. Poor -f tting and missing f ller pieces are a primary f actor that affects leakage.

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●

●

●

●

Many dra w-through AHUs or PUs form a ne gative pressure re gion between the mix ed plenum (including air f lter) and the supply f an inlet. Any gaps in the casing, access panels, or doors downstream from the air f lter are pathways for unf ltered air to enter the air system. All joints of the casing and access doors should be caulk ed or tape-sealed. Check all pipe and electric conduit penetrations for leaks. Improper f t and loose seating that allo ws a substantial amount of air leakage to bypass the f lter are the most frequent problems observed. High-eff ciency microf ne f berglass media of the bag f lter have low tensile strength and are susceptible to punctures. When the media contact with sheet-metal scre ws or static pressure tips, it causes damage. If two adjacent pockets touch, abrasion results in a breach and the y should be replaced. The bottom row of the bag f lter may come in contact with condensate water in some f lter banks. W et f lters gro w microor ganisms. These bags are either mechanically supported or changed to rigid-style f lter to prevent wetness. Rigid extended surface f lters are pleated in box ed or cartridge style. They are usually damaged by rough handling. The seal of the f lter pack to the interior of its housing is essential. Panel-type, low-eff ciency f lters should be inspected to be sure that the y are not installed backward. The air entering side is more open and dry , and the air leaving side is more dense and oily . An inspection of particulate pathw ays or leaks should be conducted each year and the observ ations recorded in a logbook.

24.5 REMOVING GASEOUS CONTAMINANTS BY ADSORBERS AND CHEMISORBERS Indoor Gaseous Contaminants Among the indoor air contaminants as discussed in Sec. 4.10, combustion products, volatile organic compounds (VOCs), smoke, objectable odors, and radon are all gaseous contaminants. Radon is an odorless, colorless naturally occurring radiative gas. Only about 6 percent of U.S. homes ha ve annual average radon concentrations exceeding the acting level 147 Bq / m3 (4 pCi / L) set by the U.S. EPA. According to ASHRAE Handbook 1997, Fundamentals, for homeowners, a short-term charcoal canister is used as a screening technique to determine whether the long-term alpha track method is necessary. If long-term measurement (3 months to 1 year) f nds that the a verage indoor radon concentration exceeding the acting level of 147 Bq / m3 (4pCi / L), actions to reduce exposure to indoor radon concentrations must be taken. Exposures to indoor radon may be reduced by: ●

●

SH__ ST__ LG__ DF

Preventing radon entry into the building Removing or diluting radon and the radon progeny

An active subslab depressurization system is the most widely used ef fective method. This system uses a f an to dra w soil gas containing radon and its progen y from pipe work and e xhausts the soil gas outdoors. Because of the depressurization of the soil gas containing radon beneath the f oor slab, the indoor radon concentration may be signif cantly reduced. Tobacco smoke is the most common indoor gaseous contaminant. There are two types of smoke: mainstream and sidestream. Mainstream smok e goes directly to the respiratory tract of the smok er. Sidestream smok e is the part e xhaled by the smok er and smok e produced between puf fs. Both mainstream and second-hand smok e may cause cancer . In many public spaces in the United States smoking is prohibited, and in man y buildings, smoking is limited to special areas. Outdoor air requirements for ventilation in a smoking lounge are increased to 60 cfm / person (30 L / s person) in ASHRAE Standard 62-1999. Because gaseous indoor contaminants ha ve v ery small molecules, gaseous adsorbers and chemisorbers are often used

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To eliminate or to reduce indoor gaseous contaminants such as comb ustion products, VOCs, and objectable odors According to the ASHRAE Standard 62-1999, to reduce the concentration of the kno wn and specif able contaminants in recirculating air to the required concentration le vel based on indoor air quality procedures to achieve acceptable air quality

Activated Carbon Adsorbers Brown et al. (1994) reported that mean concentrations of VOCs in v arious of f ce b uildings ranged from 0.18 to 4.15 mg / m3 with a new building having a total VOC concentration as high as 39.3 mg / m3. For each individual compound, its mean concentration was about 0.05 mg / m3. Granular acti vated carbon (GA C) is one of the most widely used adsorbents for indoor gaseous contaminants. Adsorption (physical adsorption) is a surf ace phenomenona. In an adsorption process, the molecule of the gaseous contaminant ●

●

●

Must be transported from the carrier airstream across the boundary layer that surrounds the adsorber granule Must diffuse into the pore of the adsorbent and occupy the surface Must be bound to the surface

Adsorption is different from absorption in which the gaseous contaminant is dissolv ed in or reacts with the absorbing medium, which can be either a porous solid or a liquid.

GAC Performance As discussed in Sec. 15.16, removal eff ciency, adsorption capacity, and service life are important performance parameters for GA C. Toluene has been used as a surrogate for indoor VOCs for the testing of GA C performance. An increase in relati ve humidity generally reduces the adsorption of VOC on GAC. This may affect some compounds more than others, and it becomes more signif cant at low VOC concentrations. When the concentration of the contaminant downstream of the adsorber rises until its concentration is the same as that of the upstream, the penetration at this time is called breakthrough penetration. In addition to the adsorption capacity and the remo ving eff ciency discussed in Sec. 15.16, the performance parameter breakthrough time tb is de f ned as the elapsed time between the start of challenge and the time when the penetrating concentration reaches the speci f ed breakthrough fraction, 10 percent breakthrough time t10% or 50 percent breakthrough time t50%. Another performance parameter is called capacity at tb which is e xpressed as the ratio of the mass of contaminant collected at a particular breakthrough time to the mass of the GA C bed, in percentage. Van Osdell and Sparks (1995) reported the performance of GAC challenged with toluene at various concentrations at 77°F (25°C) and 50 percent relative humidity as follows:

Challenge concentration ppm Challenge concentration, mg/m3 Carbon bed mass, g t10%, h t50%, h Capacity at t10%, percent Capacity at t50%, percent

0.44 1.7 24.1 625 750 6.6 7.9

1.1 4.0 23.1 344 422 9.2 11.3

9.2 34 23.4 72 88 16.2 19.7

71.7 270 25.8 11.9 15.0 18.8 23.8

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Several test reports have estimated that the lifetime of GAC for indoor VOC control exceeds 1000 h depending on its concentration. The retenti vity is the maximum amount of contaminant that the GAC can retain in clean air. The retentivity of GAC for toluene at 2860 mg / m3 is 17 percent.

Simulating GAC Applications Van Osdell and Sparks (1995) did a study based on an IA Q simulated model utilizing a computer program RISKBETA, which is a thoroughly mixed room model incorporating source / sink behavior that can predict concentrations as a function of time. The simulated model is a cop y / storage room of 9.6 ft 13.1 ft (3 m 4 m) in a b uilding that has eight rooms. The air system serving this building has a supply v olume f ow rate of 1177 cfm (2000 m 3 /h) of 5 ach. It is assumed that 4.4 lb (2 kg) of toluene is spilled in the cop y / storage room. One-half of the spilled solv ent is cleaned up as a liquid; the other half v aporizes and is carried throughout the b uilding by the air system. There is no other contaminative source in this b uilding, and the outdoor air is assumed to be VOC-free. A GAC adsorber is installed in the air system to remove VOCs from the recirculating air. Because of the buildup of the concentration of VOC contaminants in the indoor air , they are adsorbed into the sinks (paints, carpets, upholstery) and loaded into high levels. As the concentrations of toluene of indoor air drops and decays to belo w the 0.1 mg / m3 level, the sinks become contaminative sources and reemit the toluene at low but signif cant rates. Figure 24.1 shows the predicted concentrations within a cop y room and the a verage concentration of the whole building with or without the GAC adsorbers based on the data and analysis of Van Osdell and Sparks (1995). The time between the beginning of the decay of the concentration of toluene in the copy room down to the 0.1 mg / m3 level without a GAC adsorber is less than 10 days. The difference between the concentration decay curv es with or without a GA C adsorber is comparati vely small for the f rst two days and then be gins to diverge because the GA C adsorber remo ves more contaminants than the sinks can reemit. Similar results can be found for the rest of the building. As the concentration of contaminant continues to drop, the GAC will reach its retenti vity limit for the incoming contaminant concentration at around 0.2 mg / m3 and will cease to collect. In

1000

VOC concentration, mg/m3

100 Copier’s room without adsorber

10

1 Rest of building average with adsorber

0.1

0.01

SH__ ST__ LG__ DF

0

40

Copier’s room with adsorber

80

120

160

Time, h FIGURE 24.1 VOC concentration versus time decay curve for a simulated copy room.

200

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Fig. 24.1, the sharp drop of the concentration decay curv e will be gin to parallel the curv e that is without GAC adsorbers. Because the building average concentration remains above 5 mg / m3 for about 15 h, using GAC adsorber alone does not mitigate the spill. Better cleaning, short-term increased ventilation, a separate e xhaust system for the cop y room, or other similar measures should be tak en to reach an acceptable level more quickly and to reduce the loading of the sources / sinks in the building. Chemisorption Chemisorption is a type of adsoption that occurs when the molecules of a v olatile contaminant are f rst physically adsorbed onto the adsorbent. As soon as the y are adsorbed, they react chemically with the chemical impre gnant added to the surf ace of the adsorbent. The chemical impre gnant makes the adsorbent more or less specif c for a contaminant or a group of contaminants. Chemisorption differs from the physical adsorption as follows: ●

●

●

●

Chemisorption is generally irreversible whereas adsorption is reversible. Chemisorption impro ves when temperature increases; adsorption impro decreases. Chemisorption does not produce heat, but requires heat. Water vapor often helps chemisorption but usually hinders adsorption.

ves when temperature

Most chemisorptive media are formed from a high porous support. Potassium permanganate-impregnated alumina (PIA) is the most widely used chemisorpti ve medium to remo ve gaseous contaminants. GAC does effectively remove most hydrocarbons, many aldehydes and or ganic acids, and nitrogen dioxides due to its high ratio of surf ace area to v olume. PIA is ef fective against sulfur oxides, formaldehyde, nitric oxide, hydrogen sul f de, and lo wer-molecular-weight aldehydes and organic acids. PIA is often used in conjunction with GA C to co ver a v ery broad spectrum of gaseous contaminant removal. Chemisorber Performance A chemisorber is an adsorber with chemically impre gnated media. Most chemisorbers consist of chemisorptive media from a high porous support such as acti vated alumina or acti vated carbon coated or impregnated with a chemical reactant. Three types of chemisorber cartridges are currently a vailable: V-bank of lar ge-mesh carbon trays, pleated dry composite media with f ne-mesh carbon, and pleated nonw oven carbon-coated fabric. Removal eff ciency, service life, and pressure drop are essential performance parameters. Removal Efficienc . (See Section 15.16.) Remo val eff ciency depends on chemical impre gnates, granular size, type of cartridge, and challenge concentration. Joffe (1996) reported that the remo val eff ciency of NO 2 measured at the be ginning of a 2000 cfm (940 L / s) and 160-ppb challenge for a pleated dry composite medium with f ne-mesh carbon chemisorber dropped from about 85 percent at the be ginning to about 60 percent after 10 h of operation. F or V-bank large-mesh carbon trays, the average removal eff ciency is only about 30 percent in a period of 10 h of operation. High removal eff ciency is important for toxic and corrosi ve gases and becomes critical in odor control. According to ASHRAE Handbook 1997 , Fundamentals, the human olf actory response S is nonlinear. The relationship between percei ved intensities and the contaminant concentration C can be expressed as S  k Cn where k  a multiplying factor n  exponential factor, varies from 0.2 to about 0.7

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This relationship indicates that the concentration of some odorants must reduce substantially to make a perceived odor level change. Upon e xposure to contaminated air , the acti ve surf ace is co vered by absorbed pollutant molecules, or the product of a chemical reaction. This results in a decrease of remo val eff ciency over time. Pressure Drop. There is al ways a compromise between higher performance and lo wer pressure drop across a chemisorber . At a f ace velocity of 500 fpm (2.5 m / s), the pressure drop is typically 0.4 in. WC (100 Pa).

24.6 MAINTENANCE TO GUARANTEE IAQ Inspection, Service, and Access Inspection, service, access, and maintenance to prevent system failure and component deterioration are essential factors to guarantee that air f lters, coils, ducts, dampers, and other system components are properly and ef f ciently operated. Keep all the surf aces clean that surround the air f ow passage and have contact with the outdoor air, recirculating air, and supply air, and without moisture except cooling coils, condensate pans, and evaporative coolers. Access is the critical f actor for inspection, service, and maintenance for an HV AC&R system. Many IAQ problems are the result of clogged coils or f lters, blocked condensate pans, or carryover of the condensate droplets. If there is no con venient access to inspect the coil bank and f lters, unless the AHU is partially dismantled, a small problem may become a disaster. Hinged access doors to the f lter section, fan sections, both sides of coil banks, condensate pans, and mixed plenum are essential. Standing water, mold and microbial growth, accumulated dust, and improperly operated sensors and instruments must be con veniently accessed, inspected directly and regularly, and promptly serviced. The more diff cult the access, the later the problem will be disco vered, and less chance that the system component causing poor IAQ will be promptly found and improved. Monitoring of Operating Conditions in Air Systems The operating conditions that should be monitored and inspected include the following: ●

●

●

●

●

●

●

Position of outdoor air dampers and minimum ventilation air controls Pressure drop across each f lter Cleanliness of condensate drain pans in AHUs, PUs, and fancoils Cleanliness of coils Microbial growth in standing-water systems such as open-water system in condensate pans, evaporative coolers, and cooling towers Minimum setting in VAV boxes Optimum operation of exhaust or relief air system

Coils and Ductwork

SH__ ST__ LG__ DF

Dust accumulated in coils and on the surf ace of the duct inner liner is best pre vented by installing a medium-eff ciency air f lter (MERV7 with 50 ≤ E3 < 70 percent or higher) and without an y air leakage bypassing the air f lter. Ottne y (1993) noted man y hospitals installed a pre f lter and a higheff ciency f nal f lter on the air leaving side of the AHU. The coils and ductwork between the pref lter

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and f nal f lter in air systems with leak-free system components remained relati vely free of visible particulate accumulation. For a wet coil with condensate, a f ace v elocity  550 fpm (2.75 m / s) should be selected for corrugated f ns to prevent condensate carryover. If a f berglass duct liner is used, a wet coil face velocity  500 fpm (2.5 m / s) should be selected to prevent microbial growth. As discussed in Sec. 15.11, the condensate pan and condensate trap should be properly designed and installed. Condensate in the pan should be drained completely . Dirt, scale, and debris must be cleaned regularly. For heating coils, if the coil leaks, the chemicals used in the hot w ater may be aerosolized and enter the air directly, causing IAQ problems.

24.7 SPACE PRESSURIZATION CONTROL As discussed in Sec. 4.13, the purpose of space pressurization control is to maintain a positive pressure in an occupied space, a noncontaminated space, or a clean space to prevent untreated air, contaminated air, air with objectable odors, or toxic air entering or in f ltrating this occupied space, and to maintain a ne gative pressure in contaminated or toxic area so that contaminated or toxic air (treated if required) e xhausts outdoors directly . Air always f ows from a place with a higher pressure to a place with a lo wer pressure. Space pressurization controls ha ve been successfully applied to many laboratories, clean spaces and rooms, and industrial applications.

24.8 LEGAL RESPONSIBILITY FOR IAQ CASES SBS or IAQ Cases According to Eisenstein (1992), the 1980s sa w litigation in volving signi f cant claims based on indoor air pollution, sick building syndrome (SBS), or IAQ. In the mid-1980s, there were homeowners and occupants of commercial b uildings f ling lawsuits claiming injuries from formaldehyde exposure. Cases f led by o wners of mobile homes where lar ge quantities of products containing formaldehyde are used ha ve drawn much attention. There were also cases f led claiming personal injuries resulting from radon contamination and pesticide pollutants.

Who Is Legally Responsible The most frequent cases of SBS or IA Q occurred in commercial of fice b uildings with sealed windows in which occupants rely on the HV AC&R system for their thermal comfort control and ventilation requirements. A typical case is often caused by a combination of f actors, including different toxic indoor contaminants from b uilding materials, inadequate v entilation air supply due to defective ventilation control, and improper operation and maintenance of sophisticated air systems. Eisenstein (1992) recognized that SBS cases ha ve the potential of spreading a web of liability to a lar ge number of professionals and industries. The prosecuting party , or plaintiff, is often an individual or group of indi viduals claiming personal injuries as a result of e xposure to toxic substances in off ce suites. Another type of the SBS plaintif f is the commercial tenant who alle ges that he or she cannot conduct b usiness in the leased space because of the poor IA Q, that employees of the corporate tenant are frequently ill due to the poor IA Q, or that inadequate v entilation air causes b usiness interruption and lost pro f t. Man y of the SBS cases are multimillion-dollar lawsuits.

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The list of defendants in SBS or IA Q cases v aries with the type of b uildings and may include some or all of the following: ●

●

●

●

●

●

●

●

Building owner / managers Real estate developers Architects HVAC&R consultants, mechanical engineers, control engineers IAQ consultants General contractors, HVAC&R contractors Building and HVAC&R products manufacturers Leasing agents

Among them, the target defendants in SBS or IAQ cases are building owners who have the responsibility to pro vide safe premises, the building managers who must ensure that the b uilding and HVAC&R system are maintained properly and the HVAC&R system is operated correctly, and all the entities that participate in the design, construction, and installation of the ventilation / HVAC&R system. The theories of liabilities used against defendants depend to some de gree on the la w in various states. The following causes of actions ha ve been successful in most states: contracts or breach of lease; professional malpractice or ne gligence; strict liability; fraud, mispresentation, and punitive damages.

HVAC&R Engineer If the poor IAQ is determined to be due to defects in the ventilation / HVAC&R system, the mechanical or HVAC&R engineer who designs the v entilation / air system will be vulnerable to claims of negligence. As discussed pre viously in this chapter , regarding HVAC&R, the primary tasks that must be done to ensure good IAQ are as follows: ●

●

●

●

●

●

SH__ ST__ LG__ DF

Provide adequate ventilation to dilute the indoor air contaminants. Prevent microbial growth. Use medium- and high-ef f ciency and HEPA f lters to remo ve particulates, bacteria, and viruses. Remove specif c gaseous contaminants by adsorbers and chemisorbers. Adopt pressurization control so that only f ltered, noncontaminated air of acceptable air quality is supplied to the occupied zone, and contaminated air is e xhausted directly from the contaminated area. Implement scheduled and proper HVAC&R maintenance for effective air system operation.

It is e xtremely important that the engineers ’ design follo w the IA Q-related federal and local codes and the ASHRAE IAQ-related standards, especially ASHRAE Standard 62-1999, Ventilation for Acceptable Indoor Air Quality. If an engineer f ails to design an air system in comformity with appropriate ASHRAE standards, he or she will be unable to defeat a claim of negligence. Although an engineer cannot be expected to foresee all the uses of the conditioned spaces during the life of the building, the changes in the occupancy rates and the alteration in the use of the space should usually be anticipated by the engineer , who must design a v entilation system to accommodate an adequate ventilation rate for high contamination load during remodeling of suites. Recently, the courts have ruled that a v entilation system in a b uilding is a product. Under strict liability, the designer , manufacturer, and installer of a defecti ve product are considered liable regardless of whether their conduct w as ne gligent. This theory emphasizes the product, not an y individual, so that it is often easier for a plaintiff to succeed for a strict liability.

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From the HVAC&R side, proper and careful consideration of the design, construction, and operation of an air system and the issues that may cause poor IA Q will greatly reduce the risk of SBSrelated lawsuits. If an IAQ-related lawsuit still arises, the chance of a successful defense will be signif cantly increased.

REFERENCES ANSI/ASHRAE Standard 52.2-1999, Method of Testing General Ventilation Air-Cleaning Devices for Removal Efficiency by article Size, ASHRAE Inc., Atlanta, GA, 1999. ASHRAE, ASHRAE Handbook 1997, Fundamentals, ASHRAE Inc., Atlanta, GA, 1997. ASHRAE, ASHRAE Handbook 1999, Applications, ASHRAE Inc., Atlanta, GA, 1999. ASHRAE Standard 62-1999, Ventilation for Acceptable Indoor Air Quality, ASHRAE Inc., Atlanta, GA, 1999. Brown, S., et al., Concentrations of Volatile Organic Compounds in Indoor Air — A Review, Indoor Air, Copenhagen, Denmark: Munksgaard International Publishers Ltd., vol. 4 no. 2, 1994, pp. 123 – 134. Burge, H. A., The Fungi: How They Grow and Their Effects on Human Health, HPAC, no. 7, 1997 pp. 69 – 74. Burroughs, H. E. B., Filtration: An Investment in IAQ, HPAC, no. 8, 1997, pp. 55 – 65. Burroughs, H. E. B., The Art and Science of Air Filtration Management in Health Care, HPAC, no. 10, 1998, pp. 79 – 86. Collett, C. W., Ross J. A., and Sterling, E. M., Quality Assurance Strategies for Investigating IAQ Problems, ASHRAE Journal, no. 6, 1994, pp. 42 – 50. Eisenstein, H., IAQ: Who Is Legally Responsible? Heating / Piping / Air Conditioning, no. 8, 1992, pp. 43 – 47. Gill, K. E., IAQ and Air Handling Unit Design, HPAC, no. 1, 1996, pp. 49 – 54. Godish, T., Sick Building: Definition Diagnosis, and Mitigation, Lewis Publishers, Boca Raton, FL, 1995. Hays, S. M., and Ganick, N., How to Attack IAQ Problems, Heating / Piping / Air Conditioning, no. 4, 1992, pp. 43 – 51. Industrial News, New Approach to IAQ Urged, ASHRAE Journal, no. 12, 1997, pp. 8– 10. Joffe, M. A., Chemical Filtration of Indor Air: An Application Primer, ASHRAE Journal, no. 2, 1996, pp. 42 – 49. Kowalski, W. J., and Bahnf eth, W., Airborne Respiratory Diseases and Mechanical Systems for Control of Microbes, HPAC, no. 7, 1998, pp. 34 – 48. Liu, R. T., and Huza, M. A., Filtraton and Indoor Air Quality: A Practical Approach, ASHRAE Journal, no. 2, 1995, pp. 18 – 23. Mendell, M. J., Non-specif c Symptoms in Off ce Workers: A Review and Summary of the Epidemiologic Literature, Indoor Air, vol. 3, 1993, pp. 227 – 236 Muller, C. O., and England, W. G., Achieving Your Indoor Air Quality Goals: Which Filtration System Works Best? ASHRAE Journal, no. 2, 1995, pp. 24 – 32. Ottney, T. C., Particle Management for HVAC Systems, ASHRAE Journal, no. 7, 1993, pp. 27 – 34. Persily, A., Ventilation Rates in Off ce Buildings, ASHRAE Journal, no. 7, 1989, pp. 52 – 54. Spicer, R. C., Microbial Growth Must Be Minimized, Maintenance Is a Must, Engineered Systems, no. 9, 1998, p. 34. Sundell, J., What We Know, and Don’t Know about Sick Building Syndrome, ASHRAE Journal, no. 6, 1996, pp. 51 – 57. Turner, W. A., Controlling Ventilation during Renovation, HPAC, no. 11, 1998, pp. 49 – 52. Van Osdell, D. W. and Sparks, L. E., Carbon Adsorption for Indoor Air Cleaning, ASHRAE Journal, no. 2, 1995, pp. 34 – 40.

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ENERGY MANAGEMENT AND GLOBAL WARMING 25.1 ENERGY EFFICIENCY AND UNIT ENERGY RATE REDUCTION 25.1 Energy Use and Energy Efficiency for HVAC&R Systems 25.1 Energy Efficiency during Design, Construction, Commissioning, and Operation 25.2 Reduction of Unit Energy Rates 25.2 25.2 GLOBAL WARMING 25.3 Kyoto Protocol 25.3 Total Equivalent Warming Impact 25.3 Mitigating Measures 25.4 25.3 ENERGY EFFICIENCY 25.5 Federal Mandates 25.5 Energy Use Intensities 25.5 Energy Audits 25.6 Procedure for Energy Retrofit 25.6 Performance Contracting 25.7 Green Buildings 25.8 Energy Star 25.10 25.4 ENERGY CONSERVATION MEASURES 25.10 25.5 CASE STUDY: ENERGY CONSERVATION MEASURES FOR AN OFFICE 25.12 25.6 RELATIONSHIP BETWEEN HVAC&R SYSTEM CHARACTERISTICS AND ENERGY USE 25.12 System Characteristics and Energy Use Intensities 25.12 Energy Use of Heating-Cooling Equipment and Fan 25.13 25.7 ELECTRICITY DEREGULATION 25.14 Electric Utilities prior to Deregulation 25.14

Electric Deregulation 25.14 California Approach 25.15 Real-Time Pricing 25.15 Case Study: Automated Control of RTP 25.16 25.8 SYSTEM SIMULATION 25.17 Energy Estimation and Energy Simulation 25.17 Performance Equations 25.17 Physical Modeling 25.18 Steady-State and Dynamic Simulation 25.18 Sequential and Simultaneous Simulation 25.19 25.9 ENERGY SIMULATION OF A CENTRIFUGAL CHILLER USING PHYSICAL MODELING 25.19 System Model 25.19 Operating Parameters Affecting Chiller Energy Performance 25.20 Simulation Methodology 25.20 Evaporator Model 25.20 Condenser Model 25.22 Cooling Tower Model 25.23 Centrifugal Compressor Model 25.23 25.10 ENERGY SIMULATION SOFTWARE DOE-2.1E 25.25 Energy Simulation Software 25.25 Loads 25.26 Systems 25.26 Energy Efficiency Measures 25.27 Plant 25.27 25.11 ASHRAE/IESNA STANDARD 90.1-1999 ENERGY COST BUDGET METHOD 25.28 REFERENCES 25.28

25.1 ENERGY EFFICIENCY AND UNIT ENERGY RATE REDUCTION Energy Use and Energy Efficiency for HVAC&R Systems Energy use or ener gy consumption indicates the amount of ener gy used or consumed. Ener gy effi ciency indicates ho w efficiently ene gy is used. An energy-efficient H AC&R system maintains a comfortable indoor environment with an acceptable indoor air quality (IAQ) and consumes optimum

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energy resources. According to the ASHRAE definition energy conservation indicates that energy is used efficientl . Energy management is the ef fort and measures tak en to ensure that ener gy is used efficientl , and the unit energy rates (electric and gas rates) are reasonably lo w for the sake of reduction of ener gy cost. Ener gy management of HV AC&R systems consists of tw o areas: energy effi ciency (the reduction of energy use or energy conservation) and the reduction of the unit energy rate. As discussed in Sec. 1.7, the estimate of annual U.S. ener gy use of the HV AC&R systems in 1992 w as approximately one sixth of the total national ener gy use. In addition, energy use for HVAC&R is closely related to the release of CO 2 to the outdoor atmosphere which causes the global warming effect. Energy efficien y is a challenge to every one of us in the HVAC&R industry now and for many years to come in the future. According to EIA Commercial Building Characteristics 1992, energy sources used for the heating of the 67.8 billion ft 2 of commercial buildings in the United States in 1992 were electricity, 22.8 percent; natural gas, 51.8 percent; fuel oil, 6.6 percent; and district heating, 7.3 percent. Ener gy sources used for cooling U.S. commercial b uildings in 1992 were electricity , 80.4 percent; natural gas, 2.8 percent; and district cooling, 3.0 percent. Electricity and natural gas are the tw o primary energy sources for HVAC&R systems.

Energy Efficiency during Design, Construction, Commissioning, and Operation Energy efficien y must be achieved in every stage of HVAC&R system construction and operation: design, construction, commissioning, and daily operation. A well-designed and effectively functioning ener gy management direct digital control (DDC) system is necessary for an ener gy-efficien HVAC&R system. ●

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Design. Various alternatives should be compared and analyzed in terms of either payback years or life-cycle cost to determine which is most energy-efficient and cost-e fective. Construction and installation. Roof and external walls should be well insulated. The efficien y of installed equipment should not be less than the minimum ef ficiencies specified ASHRAE / IESNA Standard 90.1-1999. The amount of air leakages from the ducts, and the duct and pipe insulation all affect the energy use. Commissioning. The capacity of equipment, the air and w ater balance, and the coordination between various components and control systems should be carefully measured, adjusted, and commissioned. A poorly commissioned HVAC&R system will ne ver function efficiently as specifie Operation. The energy use for chillers, compressors, fans, pumps, boilers, and furnaces should be monitored, periodically checked, investigated, reduced, and improved.

For HVAC&R, to reduce the emissions of CO 2 to the atmosphere by means of increasing energy efficien y in operation is the primary action to mitigate the global warming effect.

Reduction of Unit Energy Rates Facility o wners, facility managers, or the tenants as well as the designers and operators of the HVAC&R systems all are concerned about the unit ener gy rate Er, as it affects the energy cost and the operating cost even if there is no sa ving in HVAC&R system energy use. For electricity, energy cost Ce is the product of the price of unit ener gy rate Er and energy use, in kWh or therms. The lower Er, the smaller Ce. The reduction of unit energy rate Er is closely related to the following: ●

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The ratio of unit rate of kWh to therms which af fects the designer in selecting electric cooling or gas cooling Use of thermal storage to provide off-peak conditioning to reduce Er

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Energy cost is the dominant factor that affects the operating cost of an HVAC&R system.

25.2 GLOBAL WARMING Kyoto Protocol In December 1997, more than 150 countries met in K yoto, Japan, and agreed to call for 38 de veloped nations to reduce the emissions of the greenhouse gases (GHGs) to about 5.2 percent belo w 1990 levels by 2008 to 2012, in order to mitigate the potential risk of global w arming. The agreement is called the Kyoto Protocol and includes the following details: ●

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GHG emissions are to be reduced 8, 7, 6, and 6 percent belo w the 1990 le vel for the European Union, United States, Japan, and Canada, respectively. Six greenhouse gases are co vered by the K yoto Protocol: carbon dioxide (CO 2), methane (CH 4), nitrous oxide (N 2O), perfluorocarbons (PFCs) sulfur hexafluoride (S 6), and hydrofluorocarbon (HFCs). And CO2, CH4, and N 2O will use the 1990 emission as the base v alue, whereas PFCs, SF6, and HFCs can be compared with either the 1990 or 1995 emission as the base value. Currently, reduced emission controls are only applicable to de veloped nations. Ho wever, future negotiations will address the reduction of emissions and other issues for developing nations. Kyoto Protocol, using a basket approach, permits the limitation of emissions in an y combination as long as its CO 2 - equivalent target is met. The Kyoto Protocol also permits counting of benefit derived for “sinks” (forest, soil, and land used) for emissions trading. Details remain to be worked out. The Kyoto Protocol will be opened for signature from March 16, 1998, for one year and will enter into force after it has been ratified by at least 55 nations representing 55 percent of the total emissions from the developed nations.

U.S. Ambassador Stuart Eizenstat, who led the U.S. dele gation in Kyoto, said in the 9th Annual Energy Efficien y Forum, 1998, in Washington, “Kyoto is an insurance polic y against the potentially devastating impacts of global w arming. If we act no w, the premium on this polic y will be f ar more reasonable and less costly than if we delay and hope the problem goes a way. The evidence shows that it will not.” In the 1998 ASHRAE winter meeting, an international panel of global climate e xperts concluded that the threat of global warming can be reduced if the world acts now to implement energyconserving technologies in buildings and to reduce energy use. Experts also agreed that “ener gy efficien y” is technologically possible and economically viable.

Total Equivalent Warming Impact The assessments of the global w arming effect should be based on total equi valent warming impact (TEWI) concepts. In Baxter et al. (1998), total equivalent warming impact was defined as the com bined global warming effects corresponding with the CO 2 released due to the indirect ef fect of energy use over the lifetime of a system and the direct ef fect resulting from lifetime refrigerant emissions. TEWI can be calculated as TEWI  mrt(GWP)rt  mblow(GWP)blow  EannualL where mrt, mblow  mass of refrigerant and mass of blow agent released during its useful lifetime, lb (kg)

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(GWP)rt, (GWP)blow  equivalent global warming effect of CO2 equivalent emisssion per lb (kg) of refrigerant or blow agent released, lb (kg)   conversion factor to convert energy use to global warming effect Eannual  annual energy use L  equipment lifetime, years It is difficult to calculate an absolute alue of TEWI because of the uncertainties of annual refrigerant leakage and GWP v alues. According to Baxter et al. (1998), for a typical supermark et in the United States with lo w- and medium-temperature DX refrigeration systems, estimated emissions due to refrigerant leakage are between 10 and 15 percent of the char ge annually. If HFC-404A is used as the refrigerant, TEWI is about 8.6 million lb CO 2 equivalent. Of this, 57 percent is refrigerant emissions. F or centrifugal chillers, the refrigerant leakage is between 0.5 and 4 percent of the charge annually. F or a typical 1000-ton (3500-kW) centrifugal chiller in Atlanta, Georgia, where the refrigerant leakage is 2 percent of the char ge annually, if HCFC-123 is used as refrigerant, the TEWI is about 5.1 million lb of CO 2 equivalent; of this, about 1 percent is due to refrigerant emissions. If HFC-134a is used as refrigerant, TEWI is about 5.9 million lb, and about 2.5 percent is due to refrigerant emissions.

Mitigating Measures CO2 Release due to Energy Use. David Gardner, asssistant administrator for the En vironmental Protection Agency, who attended the International Climate Change Conference in June 1997, in Baltimore, reported that “85 percent of climate change problems result from comb ustion processes.” For a combustion process using solid carbon C as the fuel, its reaction process and the released heat of combustion can be evaluated as follows: Reaction C  O2  CO2 1 lb  2.67 lb  3.67 lb

Heat of combustion H (Qout), Btu / lb (kJ / kg)  14,100 Btu / lb ( 32,800 kJ / kg)

(25.2)

As expressed in Eq. (8.4), the combustion efficien y Ec, in percent, can be calculated as Ec 

100Q out Q fuel

where Qout, Qfuel  heat output and heat content rate fuel consumed, Btu / h (W). F or any combustion process and the successi ve energy transformation processes from heat to mechanical and then to electrical, and so on, in Eq. (8.4), if you require a specified Qout for equipment in an HV AC&R system, for the sake of reducing the CO 2 in Eq. (25.2), if the fuel C must be reduced as well as the Qfuel value in Eq. (8.4), there must be a corresponding higher comb ustion efficien y Ec. Not only higher Ec but all higher ener gy efficiencies reduce Qfuel as well as the CO 2 and other greenhouse gases in the combustion process. Refrigerant Emissions. HFCs are listed as one of the six greenhouse gases. Cox and Miro (1998) predict, by 2050, HFCs may account for 2 to 3 percent of GHGs. The measures to reduce the global warming effect because of the refrigerant emissions due to the leakage and pur ge are as follows: ●

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Refrigerant leakage in supermark et medium- and lo w-temperature systems should equal or not exceed the American Refrigeration Institute (ARI) tar get value of 6 percent of the char ge annually. For chillers, a refrigerant leakage of 1 to 3 percent of the annual charge is appropriate.

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25.5

Secondary-loop systems using brine as the cooling medium and distrib uted systems in which compressors are close to the display case have far less refrigerant leakage than DX systems. The use of greenhouse gases as blow agents should be minimized.

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The selection of en vironmentally-friendly refrigerant with lo w ozone depletion and global warming effect should be considered. From Table 9.1, compare the following refrigerants:

HFC-134a HFC-404A HFC-407A HCFC-123

Ozone depletion potential (ODP)

Halocarbon global warming potential (HGWP)

0 0 0 0.016

0.28 0.95 0.47 0.02

According to Baxter et al. (1998), the energy use of centrifugal chiller using HCFC-123a as refrigerant is only about 87 percent of that using HFC-134a as refrigerant. More research is required to determine a suitable alternati ve refrigerant for chillers when ODP , HGWP, and energy use are all considered at the same time.

25.3 ENERGY EFFICIENCY Federal Mandates According to Cox and Miro (1998), the U.S. go vernment is the single lar gest energy user. In late 1990s, the federal go vernment provided energy to approximately 500,000 b uildings with a f oor area of about 3.1 billion ft 2 (288 million m 2). Energy was needed in HVAC&R, lighting, and other building services. Approximately 77 percent of the f oor area was used for housing, off ce, storage, and other purposes; the remaining 23 percent belonged to hospital, school, prison, and other uses. Many federal buildings constructed before the energy crisis of 1973 are not energy eff cient. The president’s Executive Order 12759 in 1991 and the National Ener gy Policy Act of 1992 (EPAct) required federal buildings to reduce enery use by 20 percent from 1985 levels in the year 2000. In addition, Executive Order 12902 called for an ener gy eff ciency improvement of 30 percent o ver the 1985 level by the year 2005. The Federal Energy Management Program (FEMP) w as established in 1974 to pro vide guidance and assistance to impro ve energy efficiency in implementing ener gy management plans. It has focused on the ener gy bill and potential for impro vements. The Ener gy Polic y Act established the fiscal year (FY) 1995 goal of impro ving energy efficiencies in federal buildings by 10 percent from 1985 le vels on a Btu / gross ft 2 (W / m2) of floor area. This goal is e xceeded in FY 1994 with a total reduction in federal b uildings of 11.2 percent. In addition, energy costs were $3.8 billion in FY 1994, and that was $1.5 billion less than in FY 1985. The federal government accumulated an energy savings of $9.8 billion and reduced federal building petroleum-based fuel use by 45.4 percent.

Energy Use Intensities Energy use intensity (EUI) is the annual ener gy use per unit f oor area, in MBtu / ft2 yr (kWh / m2 yr). The EUI depends mainly on the locations of b uildings, building characteristics, operating characteristics, and the HVAC&R system characteristics. According to DOE / EIA (1992), the abridged EUIs for commercial buildings in 1989 were as follows:

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MBtu / ft2 yr Building Type

Electric Gas

Assembly Education Health care Mercantile / Serv. Off ce

27 27 75 44 66

39 47 113 46 32

Electric, kWh / ft2 yr

Total

NE

MW SO

WE

64 87 219 85 104

7.1 9.5 7.3 7.3 7.4 8.7 7.9 20.1 22.6 23.4 20.1 15.1 12.5 12.1 13.4 16.3 16.7 20.9 22.4

Natural gas, ft3 /ft2 yr NE

MW

57.5 28.2 61.0 74.0 156.4 62.9 62.1 20.8 55.8

SO

WE

25.9 33.6 82.0 29.4 23.6

45.8 66.4 69.1 31.1 26.8

NE indicates northeast, MW midwest, SO south, and WE west. Source: DOE / EIA (1992). Based on 4528 buildings. Reprinted by permission.

Energy Audits Energy audit refers to a month-by-month accounting, survey, and analysis of ener gy use in a b uilding. This energy use is check ed against a b udget or an ener gy estimate based on an hour -by-hour detailed system simulation in order to identify ener gy ef f ciency opportunities. According to ASHRAE Handbook 1999, HVAC Applications, following an energy use estimate, an energy analysis can proceed on three levels: ●

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Walk-through assessment analyzes ener gy bills and a brief surv ey of the b uilding for the sak e of identifying low-cost / or no-cost energy eff ciency measures. Energy survey and analysis include a more detailed building survey and energy analysis with a breakdown of energy use within the building. Actual energy use can be compared with EUI of corresponding building types and locations. This level identifies and provides the savings and cost analysis of all practical ener gy ef ficiency measures that meet the o wner’s economic criteria. Detailed analysis of capital intensive modif cations focuses on potential capital-intensive modif cations and involves more detailed f eld data collection, system simulation, and engineering analysis.

In the detailed analysis le vel, an energy-efficient and cost-ef fective energy use estimate can be calculated according to the ASHRAE / IESNA Standard 90.1-1999 Ener gy Cost Budget (ECB) method. Refer to Sec. 25.11 for more information. The actual month-by-month ener gy bill can be compared against the ECB and energy cost breakdowns on the building envelope, load calculations, and HVAC&R system and components. Ener gy ef ficiency opportunities can then be determined. Procedure for Energy Retrofit Energy retrof t is a project whose purpose is ener gy eff ciency, i.e., to convert an existing system to an energy-eff cient system. Energy retrof t can often proceed as follows:

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1. Organize a team. One individual should be chosen to be responsible for the team. 2. Draft a plan. 3. Make a field survey by walking through the building; talking with the operating manager , engineers, operators, and electric utility representati ve; tak e f ield measurements and tests, if necessary. 4. Proceed with an energy audit. 5. Justify the energy eff ciency opportunities after cost analysis. 6. Implement the plan, starting with no-cost and low-cost opportunities.

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7. After energy retrof t is complete, operate and control the HVAC&R system according to the requirements of the proposed ener gy ef f ciency opportunities. Check the ener gy costs against those from prior to the energy retrof t.

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Performance Contracting According to Mahoney and Weiss (1997), a performance contract legally guarantees energy and operating savings based on the performance of the contractor o ver a specified period of 5 to 25 years. Performance contracting also allows a facility to complete a major upgrade in energy efficiency funded by the sa ving of ener gy costs. The contractor must pay for sa vings that f ail to materialize. Energy Service Companies. During performance contracting, most often an ener gy service company (ESCO) initiates an HVAC&R system energy eff ciency upgrade proposal to a facility owner tailored to the f acility’s needs. A performance contract includes design, installation, f nancing, project management, maintenance, and monitoring. Most of the components of a performance contract are the responsibility of the ESCO, and therefore, an experienced, technically competent person who has a good relationship with a f nancial party is the key factor for sucessful performance contracting. There are three kinds of ESCOs: ●

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Engineering ESCOs are companies that al ways specialized to pro vide engineering services, including ener gy audits and system design. More often, engineering ESCOs also include project management, f nancing, and performance guarantees. Equipment ESCOs are manuf acturers of ener gy users or controlling equipment. They establish their own energy eff ciency service divisions to expand their marketing. Many utility ESCOs are owned by utilities which provide not only energy eff ciency services, but also electric and natural gas po wer. There are also ESCOs aligned between utilities and consolidated HVAC&R contractors.

General Steps. Mahoney and Weiss (1997) noted that a performance contract w orks best when ESCOs and local contractor are well coordinated. The ESCO is reponsible for engineering, project management, guaranteed savings, and long-term project funding. The local contractor tak es care of the installation, maintenance, and other services. All involved parties, such as f acility owner, ESCO, local contractor , and equipment manuf acturers, work well when the follo wing general steps are taken: ●

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Conduct an energy use estimate. This involves assessing current conditions, collecting data to estimate potential savings, and determining factors affecting energy costs. Field surveys; interviews; review of building and HVAC&R system drawings, energy use of electric, gas, or oil for a number of years; occupanc y schedules; and required indoor en vironmental parameters are important for an energy estimate. Perform a detailed engineering analysis. Computer modeling and simulation can be used to evaluate the current ener gy use as the baseline. Compare v arious alternatives, calculate and determine each possible energy-saving measure. Design the ener gy-saving measures for retro f t with reasonable payback time. F acility o wner should approve an implementation plan prepared by the ESCO. Arrange project f nancing. After the approval of the implementation plan, the ESCO f nds suitable third-party f nancing for the payment of installation and other costs. As the transaction costs are roughly the same for small, medium, and large loans and are e xpensive, a large or combination project often helps.

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Install approved equipment. The facility owner selects the equipment type and brand and decides whether a local contractor is needed or to use in-house staf f for a small or medium-sized project. Commission the system and train operators. The ESCO v erif es the operation characteristics of the new or modif ed system components and the HVAC&R system. The facility’s staff learn how to operate new equipment, and the new or upgraded HVAC&R system. Monitor energy savings. The operation and control of the b uilding’s indoor environment and the energy ef f ciency-upgraded HVAC&R system remain with the f acility’s staf f. The ESCO also monitors the building’s indoor environment and the performance of the upgraded HV AC&R system remotely. The facility manager is noti f ed if an abnormal condition occurs. On man y occasions, the ESCO ’s monitoring service to ensure that ener gy sa vings are achie ved should be included in the performance contract. Guarantee energy savings. If energy saving does not meet the projections, the ESCO pays the difference. The ESCO monitors and reviews energy use throughout the life of the contract. The contract also allows adjustments of the facility’s baseline energy use when there is a change in the facility’s size, occupancy, or equipment. Performance contracting provides the energy eff ciency upgraded cost for the HV AC&R facility owners. Because the National Energy Policy Act of 1992 and Executive Orders 12759 and 12902 as well as utility companies of fer fewer rebates, more and more federal b uilding managers and facility owners in the pri vate sector will rely on performance contracting as a viable w ay to upgrade the ener gy ef f ciency of HV AC&R systems. According to Mahone y and Weiss (1997), “Today, performance contracting projects in the United States ha ve generated $2.3 billion in work and equipment . . . . The market si growing at a rate of 15 percent and has reached $750 million a year .” There are more than 100 companies pro viding performance contracts. One of them, an engineering ESCO, has entered into more than 1000 agreements related to ener gy eff ciency since 1974.

Case Study: Performance Contracting. According to the Heating , Piping, and Air Conditioning journal (January 1998, pp. 12 – 25), as the legislation permitted public school corporations to advertise for self-funding ener gy ef f ciency proposals from quali f ed performance contractors in 1995, the Evansville Vanderburg County School Corporation (EVCSC) in Ev ansville, Indiana, accepted a performance contract proposal from the Ener gy System Group (ESG), a utility ESCO, for energy eff ciency upgrading of 35 school b uildings. The ESG performance contract with EVCSC totaled more than $35 million. The performance contract includes fully automatic summer / winter changeover two-pipe system in 19 b uildings, pulse combustion gas boilers and DDC controls in 19 schools, and new lighting systems, windows, and doors. Based on the actual energy use records: ●

●

The new boilers lowered the natural gas consumption 45 to 55 percent in the f rst year of use. Based on the original proposal, the energy saving program will sa ve $369,000 per year . After 15 months of operation, it will actually save about half a million dollars annually.

Green Buildings

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Basics. A green building implies a building, including HVAC&R systems, that is energy-eff cient and environmentally friendly. It provides a healthy and comfortable indoor en vironment for its occupants, is friendly to the outdoor en vironment by releasing minimum pollutants into the outdoor atmosphere, and at the same time is considered sustainable, energy-eff cient, and well maintained (refer to Sec. 1.8). Sustainable design considers resources consumed in a w ay that balances the current needs with those of future generations. According to Cole and Larsson (1998), Green Building Challenge ‘98 (GBC ‘98) is a tw o-year international collaborative process of de veloping and demonstrating an impro ved method to assess

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the enviromental performance of b uildings. This process, initiated by Canada, has representatives from 13 other countries.

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GBA Tool. The second-generation assessment method de veloped during the GBC ‘98 process is called the green b uilding assessment (GB A) tool. GB A tool includes the follo wing characteristics: ●

It covers the following: Resource consumption: energy, land, water, and materials Enviromental pollutants: airborne emissions, solid waste, liquid waste, and others Quality of indoor en vironment: IAQ, thermal quality (such as indoor temperature and humidity), lighting quality, noise, and controllability of systems Longevity: sustainability, adaptability, and maintenance of performance Process: design, construction, commissioning, and operations planning Climate factors: locations, outdoor climate, sun, daylight, wind impacts

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●

It has a core set of detailed criteria and subcriteria that can be modi f ed to suit national, regional, and building type variations. Assessments are based on applicable re gulations or industrial norms in various regions. The structure can be used at various levels of detail. The tool re f ects a consistenc y in terminology and a scoring system. All criteria and subcriteria are assessed on a  2 to 5 scale in which 0 is typical practice, 3 is best current practice, and 5 is a demanding, attainable goal. It comprises a weighting system that can be modif ed to emphasize regional priorities.

Case Studies of Green Buildings. The follo wing is a description of the essentials of tw o green buildings: Gottfried et al. (1997) introduced San Diego’s Ridgehaven Building (RB), San Diego, California, and Morrison (1998) reported on the Ener gy Resource Center (ERC), in Downey, California: RB Building type Year constructed Floor area, ft2 HVAC&R system Energy eff ciency, kWh / ft2  yr IAQ OA, (outdoor air) cfm / person Filter eff ciency, dust spot Daylighting

Off ce 1981 73,000 WSHP (water-source heat pump) 9 20

ERC Off ce-laboratory 1950s, retrof t 1990s 45,000 Packaged system 22 percent less than code 40 percent Maximized

Additional energy-eff ciency and IAQ measures are as follows. For the Ridgehaven Building, ●

●

●

●

●

Most material removed from the building had to be reused. Exterior foil, metallic f lm-coated foam, and e xhaust systems are used to minimize glass VOC, and particulates entering the conditioned space. VAV systems and terminals are provided in variable-occupancy areas. Variable-speed drive for outdoor air fan and condenser water pumps are used. High-eff ciency motors are utilized.

f ber,

For the ERC: ●

●

Direct-f red 30-ton double-effect absorption chiller-heaters Two dual-wheel desiccant dehumidif cation units

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●

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●

●

AHU with evaporative coolers Variable-speed drive to save fan energy use Nontoxic products such as interior paints containing no petroleum deri vatives, VOCs, and nontoxic f oor sealers, padding, tile adhesives, and recycled carpeting material Linings in the ducts and plenums to prevent the growth of mold and bacteria Use of CO2 -based demand-controlled ventilation

Energy Star The U.S. Department of Ener gy (DOE) and the U.S. En vironmental Protection Agency (EPA) are collaborating in rating to promote an energy-eff ciency program for the recognization of EPA’s Ener gy Star Benchmarking Tool assesses a f acility’s comparati ve ener gy performance against standardized indoor criteria and compares with similar -use buildings in United States according to Hoggard (2000). Each b uilding is assigned a benchmark score from 0 to 100. All buildings scoring 75 or better qualify for an Ener gy Star label for b uildings. Currently, the Energy Star label is only available for off ce buildings, and will be available for schools, retail stores, and other buildings later on. Equipment of the top 25 percent of products such as computers, printers, copiers, and HVAC&R equipment concerning energy eff ciency with the Energy Star label. ●

●

Buildings with the Ener gy Star label are ener gy-eff cient b uildings, whereas green b uildings are both energy-eff cient and environmentally friendly.

25.4 ENERGY CONSERVATION MEASURES In principle, adopting an inte grated design approach by considering the b uilding as a whole, minimizing heating and cooling loads, selecting high-eff ciency equipment sized as closely to the design load as possible, emphasizing commissioning, and optimizing operation promote ener gy eff ciency (energy conserv ation) in b uildings. The follo wing are possible ener gy conserv ation measures (ECMs) for the HVAC&R system in buildings: ●

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●

●

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Turn off electric lights, off ce appliances, and other equipment when they are not needed. Shut down AHUs, PUs, fan coils, VAV boxes, chillers, fans, and pumps when the space they serve is not occupied or when the y are not needed, except during w arm-up, cool-down, or the pur ging period prior to the morning occupied period. The time chosen to start or to stop the AHUs, PUs, chillers, and exhaust fans daily should be optimum. The set point of space temperature, relative humidity , cleanliness criteria, and indoor -outdoor pressure differential should be optimum. Using different space temperature set points for summer and winter seasons for comfort air conditioning is often energy-eff cient. The dischar ge air temperature from the AHU or PU and the temperature of w ater lea ving the chiller or boiler should be reset for part-load operation based on the space temperature, system load, or outdoor temperature. Reduce air leakages from ducts, dampers, equipment, and HVAC&R system components. Use weather stripping to seal windo ws and re volving exterior doors to reduce in f ltration. Ducts and pipes should be well insulated. Carefully design the layout of ducts and pipes to minimize their length, the number of duct and pipe f ttings, as well as their pressure losses.

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25.11

Use more energy-eff cient cooling methods such as free cooling, evaporative cooling, or groundwater cooling instead of refrigeration. Replace refrigeration with an air economizer , water economizer, evaporative cooler , and e ven desiccant dehumidi f cation if doing so pro vides the same cooling results, is more ener gy-eff cient and is cost-ef fective. An evaporative condenser is often more energy-eff cient than a water-cooled or air-cooled condenser in many U.S. locations. Use heat recovery systems, waste heat from gas-cooling engines, or heat pumps to provide winter space heating when they are applicable and cost-effective. Use variable-air-volume systems instead of constant-v olume systems, and variable-f ow building loop water systems instead of a constant- f ow one if the air f ow or the w ater f ow should be reduced during part-load operations. Lar ge v ariable-speed aerofoil centrifugal f ans are ener gyeff cient and often cost-effective in VAV systems. Use energy-eff cient chillers, boilers, AHUs, PUs, and motors. Todesco (1996) recommended that chillers have an ener gy use inde x of 0.50 to 0.55 kW / ton, condensing boilers of 90 percent ef f ciency, and large supply and return fans of 65 percent total eff ciency and higher.

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ASHRAE / IESNA Standard mandates that minimum nominal full-load eff ciency, in percent, for electric motors shall comply with the requirements of Energy Policy Act of 1992 as follows:

●

●

●

Open motors, rpm

Enclosed motors, rpm

Motor horsepower, synchronous speed, rpm

3600

1800

1200

3600

1800

1200

1 1.5 2 5 10 20 50 100 200

82.5 84.0 85.5 88.5 90.2 92.4 93.0 94.5

82.5 84.0 84.0 87.5 89.5 91.0 93.0 94.1 95.0

80.0 84.0 85.5 87.5 90.2 91.0 93.0 94.1 94.5

75.5 82.5 84.0 87.5 89.5 90.2 92.4 93.6 95.0

82.5 84.0 84.0 87.5 89.5 91.0 93.0 94.5 95.0

80.0 85.5 86.5 87.5 89.5 90.2 93.0 94.1 95.0

Adopt a direct digital control (DDC) ener gy management and control system for lar ge and medium-size HVAC&R new and retrof t projects. Adopt double-pane windo ws with lo w-emission coatings. Todesco (1996) recommended a U value of 0.25 Btu / h ft2 °F (1.4 W / m2 °C) for exterior windows, also lower U values for external wall and roofs. Reduce the heat gain in the summer and heat loss in the winter in locations where the outdoor temperature is high in the summer and low in the winter. Use daylighting and controls for the perimeter zone. Also use energy-eff cient f uorescent lamp and electronic ballasts with a goal to achieve connected lighting loads of 0.75 W / ft2 (8 W / m2) or better.

According to EIA’s Commercial Building Characteristics (1994), in commercial b uildings, the ratio of the f oor area that had adopted the captioned ener gy conservation feature to the total conditioned f oor area, in percentage, in 1992 in the United States was as follows: Off-hours heating reduction Off-hours cooling reduction Energy management and control system (EMCS) Energy audit Variable-air-volume (VAV) systems Economizer cycle HVAC&R maintenance

68.1 percent 63.0 percent 21.1 percent 21.8 percent 20.5 percent 27.0 percent 72.4 percent

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25.5 CASE STUDY: ENERGY CONSERVATION MEASURES FOR AN OFFICE Parker et al. (1997) introduced an ener gy-eff cient off ce building design for the Florida Solar Energy Center (FSEC) in the hot and humid climate of Cocoa, Florida. This off ce building consists of off ces, a visitors’ center, and laboratories and has a f oor area of 41,000 ft 2 (3809 m2). Laboratories were not included in this ener gy analysis. Eighteen ener gy conserv ation measures (ECMs) were considered during the ener gy analysis. Some of the ECMs are not ef fective for the hot, humid climate in Florida; e.g., the air economizer c ycle sho wed little sa vings, and higher than code-mandated levels of w all and roof insulation resulted in little adv antage. Only the follo wing ten ECMs provided cost-effective energy savings. ●

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●

A high-eff ciency lighting system of 0.9 W / ft2 (9.7 W/ m2) with T-8 f uorescent lamps in a ref ective troffer and with electronic ballasts saves energy use. Windows comprised of glazing units with high visible transmittance of 0.56 for daylighting and a low shading coef ff cient of 0.33 reduce unw anted solar heat gain. Also a lo w U value of 0.31 Btu / h  ft2  °F (1.76 W / m2  °C) reduces conductive heat gain. Daylighting perimeter illumination coupled with solar control system on the south f acade enables effective use of the perimeter of f ce illumination. Dimming electronic ballasts are controlled by photometric sensors to adjust ballast output combined with the a vailable daylight to maintain a constant desktop illumination level. Two high-eff ciency screw chillers provide an IPLV of 0.65 kW / ton (COP  5.41) and of 0.60 kW / ton (COP  5.86) at 50 percent part load. Energy star personal computers, printers, and copiers are used to save energy. A ref ective white single-ply roof membrane is chosen instead of a gray or black one, to reduce the solar irradiated heat gains through the roof. Use of occupancy sensors shuts off the VAV terminals when a room is not occupied. Using a DDC EMCS permits an increase in the cooling set point from 75 to 76°F (23.9 to 24.4°C) because of a f ner control tolerance as well as provides optimal start and stop capability. Using variable-speed fans and variable-speed building pumps reduces the fan speed or pump speed during part load to save fan and pump energy use. Analysis showed that they are cost-effective. CO2 -based demand ventilation control is adopted for intermittently occupied spaces and zones.

In Florida, the a verage ener gy use of 160 state-o wned of f ce b uildings in 1991 w as 67 MBtu / h  ft2  yr (211 kWh / m2  yr). With these ten ECMs, the predicted ener gy use for the of f ce building of FSEC will drop to 27 MBtu / h  ft2  yr (85 kWh / m2  yr).

25.6 RELATIONSHIP BETWEEN HVAC&R SYSTEM CHARACTERISTICS AND ENERGY USE System Characteristics and Energy Use Intensities

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As discussed in Sec. 25.3, for an HVAC&R (air conditioning) system, if the location of the building and the construction of the b uilding it serv es as well as its operating parameters are nearly the same, the energy use of this HVAC&R system varies as its system characteristics (types and conf gurations) change. According to the DOE / EIA 1998 nonresidential b uildings ener gy consumption survey of commercial b uildings’ consumption and e xpenditures in 1995, the energy use intensity (EUI) of various HVAC&R systems is listed in Table 25.1.

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TABLE 25.1 Energy Use of Various HVAC&R Systems System characteristics

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EUI, kBtu / ft  yr (kWh / ft  yr) 2

Heat pumps (including air-to-air and WSHPs) Packaged systems (including VAV systems) Evaporative systems Individual systems Single-zone, constant-volume packaged systems Central systems (including VAV central systems)

2

28.1 37.1 38.0 43.2 44.0 44.7

(8.2) (10.9) (11.1) (12.7) (12.89) (13.1)

In Table 25.1, WSHP represents w ater-source heat pump. Heat pump systems that include the air-to-air heat pumps and WSHP systems consumed the least ener gy because ener gy use in space heating was the lowest, as shown in the breakdo wn of heating, cooling, ventilation, lighting, off ce equipment, etc. below. Also, central systems consumed the most ener gy because the sensible cooling load of lighting, off ce equipment, etc. was about 50 percent higher in central systems than in packaged systems. Energy consumption, kBtu / ft2  yr Heat pumps Packaged systems Individual systems Central systems

Heating

Cooling

Ventilation

Lighting

Off ce equipment

Others

16.1 26.0 36.0 29.0

9.0 7.9 5.5 10.0

2.8 3.2 1.7 5.7

23.4 24.6 16.5 34.0

7.6 6.7 3.9 10.5

6.6 6.4 6.6 11.2

The energy consumption of cooling and ventilation in individual systems was the lowest. This indicated that the fan energy use was lowest with a lower fan total pressure and a poorer outdoor v entilation air supply than in other systems.

Energy Use of Heating-Cooling Equipment and Fan As the heating energy used in an HVAC&R system mainly depends on the heating system characteristics, the efficiency of the direct-f ired gas furnace and of the gas boiler is around 90 percent for the condensing furnace and boiler and 75 to 80 percent for con ventional furnaces and boilers. The energy use of cooling equipment mainly depends on the characteristics of the refrigeration system. For centrifugal chillers, it is assumed that ener gy-use for the centrifugal compressor is 0.55 kW / ton (6.39 COP), and the energy use of water pumps and cooling tower fans is 0.2 kW / ton. For packaged systems, the energy use of reciprocating or scroll compressors including the air cooled condenser fans is from 1.1 to 1.2 kW / ton. The energy use of v entilation depends mainly on the v entilation rate and the f an characteristics. From Eq. (15.6), fan energy use P, in hp (kW), can be calculated as Pf 

pf V˙ 6356f m d

(25.3)

where pf  system total pressure, in. WC (Pa) V˙  volume f ow rate, cfm (L / s) f md  combined fan, motor, and drive eff ciency According to ASHRAE / IES Standard 90.1-1989, typically pf and fmd have the following values:

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TX Constant-volume single-zone packaged system Multizone packaged system VAV reheat packaged system VAV reheat central system Dual-duct VAV central system Fan-coil system Water-source heat pump system

pf, in WC (Pa)

f md

2.5 (625) 4.0 (1000) 4.0 (1000) 5.0 (1250) 6.0 (1250) 0.5 (125) 0.5 (125)

0.40 0.45 0.45 0.55 0.55 0.25 0.25

For both f an-coil and w ater-source heat pump systems, for each ton of refrigeration capacity there is about a corresponding 320 cfm (151 L / s) of v olume f ow for the small f an in the f an-coil unit and 80 cfm (38 L / s) of volume f ow in the dedicated outdoor ventilation system. The dedicated outdoor ventilation system has often a total pressure loss of 5 in. WC (1250 Pa). For VAV systems, the system total pressure often has a 1 in. WC (250 P a) f xed pressure loss. The variable part will drop to an a verage of its 65 percent design system total pressure loss during part-load operation.

25.7 ELECTRICITY DEREGULATION Electric Utilities prior to Deregulation An electric utility is v ertically inte grated from three main components: generation, transmission, and local distribution network, as shown in Fig. 12.11 b. There are usually four kinds of generation plants: high-eff ciency coal and nuclear plants and ineff cient diesel and gas-turbine plants. Prior to the electricity dere gulation, large f acilities purchased electricity often based on the time-of-use (TOU) electric rate schedule that di vides a 24-h working day into two or three periods, on-peak (such as from 9 a.m. to 10 p.m. weekdays) and of f-peak periods (10 p.m. to 9 a.m. weekdays plus week ends and holidays) or on-peak, partial-peak, and off-peak periods. Higher rates are charged for on-peak hours than for of f-peak hours. In addition, there is a demand char ge and a monthly service charge based on the largest on-peak electricity power demand. In the 1990s prior to the electricity deregulation, unit electric rates ranged between $0.02 and $0.16 per kWh with an a verage around $0.07 per kWh in the United States. By using a time-of-use electric rate schedule, electric utilities tend to shift the electricity from on-peak hours to off-peak hours, and they shift the daytime peak operation of ineff cient diesel and gas-turbine plants to nighttime base-load high-ef f ciency coal and nuclear plants to reduce operating costs.

Electric Deregulation

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Electric utilites were originally re gulated to pre vent duplicated in vestments in electricity generation, transmission, and distribution netw orks to serv e the same community . Electricity dere gulation encourages open mark ets and competition and is generally considered as a means to reduce the cost of electricity and services. According to Warwick (1997), in the United States, deregulation began 20 years ago with telecommunications, gas, banking, airlines, and trucking. For electric deregulation, the Public Utility Re gulatory Policies Act (PURPA) of 1978 required retail utilities to b uy po wer from small generation plants de veloped by independent producers. This formed a competitive market in power generation and power wholesalers. In 1995, the Federal Energy Regulatory Commission (FERC) required utilities with b ulk power transmission lines to pro vide open access to wholesale power buyers and sellers. As of 1997, two of the three major operations, generation and transmission, are open to competition as the result of the federal reforms. The ne xt

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25.15

phase of electric deregulation requires direct access to retail mark ets based on the condition that a competitive power market is provided.

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California Approach In late 1990s, several states follo wed California ’s lead in ha ving le gislation in place by setting a specif c date for electric dere gulation, such as Arizona in 1999, California in 1998, Massachusetts in 1998, New Hampshire in 1998, Pennsylvania in 1999, and Rhode Island in 1997. In 1999, according to Warwick (1998), Waintroob (1998), and Gottfried (1997), the following holds: ●

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●

States in which electric utilities are dere gulating typically adopt one of the follo wing tw o approaches. The f rst approach is to provide electric consumers with an immediate rate cut but to retain re gulatory control o ver power rates for the incumbent utility . Power-supplying competitors have to offer power below the new, discounted rates. In most states, the new rate is accompanied by a rate freeze or cap lasting for several years. The second approach is to separate the power cost from the old utility rate and to allow electric users to buy from alternate competitors. Power supply is the major electric utility service for competition under electric utility dere gulation. California established a po wer exchange pool which acts as a po wer auction house and an independent (transmission) system operator , or ISO, which manages all po wer transfers and settles all power sales accounts when utilities continue to o wn and maintain their transmission lines. The experience of direct access in Ne w Hampshire using bilateral trading allo ws individuals to negotiate for the best deal the y can get. Those who do well have lower-cost good-quality electricity. Bilateral trades are contracts with recei ve and delivery terms. The buyer has to arrange transmission, if the contract is for power only. To implement competition without involving lawsuits by incumbent electric utilities and other jurisdictions that rely on the tax dollars generated from po wer plant v aluations, electric utilities were of fered means to reco ver their stranded costs within a certain period. In California, the state’s degeneration bill AB 1980, enacted in September 1996, calls for a nonbypassable stranded cost surcharge of approximately 2.5 to 6 cents per kWh payable o ver 4-year period until 2002 by rate payers. In 2002, California consumers will realize the bene f ts of dere gulated rate. Stranded cost is the cost difference between a plant cost carried on its books and the real market value.

Real-Time Pricing Recently, the use of the computer and data handling softw are, the improvements in telecommunications, as well as the demand for better ener gy management follo wing electric dere gulation ha ve made it possible for electric utilities to of fer to customers an electric rate structure that more precisely represents the actual cost of providing the electric service. In addition, this electric rate structure also depends on the climate and demand; v aries hourly, 24 h daily and 8760 h annually; and is known as real-time pricing (RTP). It began in 1985 in California. Many states allowed electric utilities to of fer RTP electric rate structure to commercial and industrial customers in the late 1990s. Other rate structures will also be a vailable in a dere gulation environment as the po wer pro viders maneuver to offer the most competitive plans possible. According to Bynum (1998), the components of a typical bill, in percentage, in the late 1990s are as follows: Cost of generation Transmission, local distribution, service, prof t, etc.

40 percent 60 percent

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25.16

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Summer peak

1.4

Unit electric rate, $/kwh

1.2 1.0 0.8 0.6 0.4

Winter peak

0.2 0

4

0

8

12 Hours

16

20

24

FIGURE 25.1 Real-time pricing (RTP) unit electric rate daily range at summer and winter peaks.

Figure 25.1 shows the daily range of RTP electric rates at summer and winter peaks. Because of the use of the RTP electric rate structure instead of the TOU rate structure follo wing electric deregulation, Norford et al. (1998) noted the following: ●

●

●

The need for such crude price signals as the demand charge is diminishing. Commercial and industrial facility operators have the opportunity to rationally e valuate the tradeoffs and the possible reduction or shifts in b uilding services, such as lights, heating, cooling, and process loads. The general strategy of shifting loads from high-priced to low-priced hours is common to both TOU and RTP approaches; however, RTP has a much greater range of price variation than TOU. Knowledge-based system computer programs are used for R TP cost optimization control. R TP control strate gies can be summarized thus: use of thermal storage systems; utilization of fuel switching of on-site auxiliary or emer gency generation or cogeneration, as well as gas cooling such as absorption chillers or gas-engine-driven chillers; general load shedding of lights and f ans; and batch scheduling of production processes.

Case Study: Automated Control of RTP

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According to Gabel et al. (1998), the f rst automated RTP control was installed in the 51-story Marriott-Marquis Hotel with 1.8 million ft 2 (167, 200 m 2) in Ne w York City. The mechanical systems include three 900-ton (3170-kW) centrifugal chillers and man y constant-volume and VAV systems. Heating is pro vided by purchased steam. The 1900 guest rooms were e xcluded from R TP control for primary room comfort. About 70 air-handling, ventilation, and lighting systems for the remaining 55 percent of f oor space were placed under RTP control.

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The control strate gy for each system or equipment has criteria that must be met before its load can be shed, reduced, or rescheduled. Here is an e xample of control logic. If the R TP price is greater than $0.10 per kWh, if the AHU AC-2 is in an occupied period, and if the space temperature AC-2 serves is not o ver 80°F (27°C), then command AHU AC-2 off. If the abo ve control parameters are not met, AC-2 continues to operate. The electric utility pro vided day-ahead notice of upcoming hourly R TP rates do wnloaded via dial-up phone line. The RTP control system responds to the rate schedule hour by hour, f rst by analyzing its preprogrammed control strate gies and then by e xecuting the most appropriate e vents in sequence. The control system also ree valuates the control logic for each controllable load each hour, making adjustments if necessary. About 70 air system loads were con f gured for RTP control. During the three most recent years with TOU electric rates, the electric energy use averaged approximately 38 million kWh per year, with a peak electric load of 6 MW. The electric bill averaged $3 million per year. With automated control response to RTP rates: ●

●

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The electric bill saved $1 million in 4 years. Average energy use dropped more than 2 million kWh each year. Peak electric load reduced 1500 kW during high-price periods.

25.8 SYSTEM SIMULATION Energy Estimation and Energy Simulation Annual, seasonal, or short-term estimation of ener gy use in an air conditioning or HV AC&R system is an important tool to assess ener gy performance and cost in both design and daily operations. The annual energy use of v arious types of air , water, refrigeration, and heating systems can be predicted and analyzed. Based on the results of ener gy estimation, the most ener gy-eff cient system can be selected. For the daily operation of air conditioning systems in e xisting buildings, detailed simulation reveals the system ’s negative and positi ve effects on ener gy use and performance by comparing the estimated evaluation against the actual measured results. The energy use of man y alternative improvements can be predicted and assessed. These estimates provide great potential for reducing the energy use and operating cost of existing air conditioning systems. The detailed simulation method uses a comprehensi ve computer program through a po werful personal computer (PC) to simulate the thermal beha vior of the b uilding, and the performance of the air, water, refrigeration and heating plants at various outdoor weather conditions, to estimate the energy use of an air conditioning system. Because of the po werful capability of the computer program, detailed simulation methods can include most of the important f actors that affect system energy use. The most widely used program is the year -round, hour-by-hour energy simulation software. Short-term simulation is sometimes used in dynamic modeling to e valuate the in f uence of the thermal storage effect of system and components on their performance. Energy simulation is the representation of ener gy use of an actual system or component by a model of analogous characteristics, and it is used to predict the performance and operating parameters of a system or component. Mathematical equations are used to describe the operating characteristics of the w orking substance and w ork and energy transfer of the system or component being simulated. Energy use simulation can be represented by either performance equations or physical modeling. Performance Equations Generally, a system component ’s performance with re gard to one or tw o independent operating variables is often a vailable from the manuf acturer’s catalog. A polynomial expression is used to f t

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CHAPTER TWENTY-FIVE

the catalog data or actual performance into a re gression equation, to mathematically relate the dependent variable z to independent variable x in steady-state simulation as follows: z  a0  a1x  a2x2  a3x3

(25.4)

To relate a dependent variable with two independent variables x and y, the polynomial expression is z  b0  b1x  b2x2  b3y  b4y2  b5xy  b6x2y  b7xy2  b8x2y2

(25.5)

where a0, a1, . . . , a3, b0, b1, . . . , b8  coeff cients. Computer programs are a vailable to solve the coeff cients according to the data from the manufacturer or actual performance data.

Physical Modeling Modeling. Setting up a component or a system model is the physical modeling. Modeling includes the following: ●

●

f rst step of ener gy simulation by

Description of system or component con f guration whether it is an air , water, refrigeration, or heating system, or whether it consists of many components or only contains a single device. Description of the operating characteristics of the system or component, and the interaction between system components, whether it can be simpli f ed to a steady-state model or a dynamic model. A simplif cation of the physical model that results in an error of only fe w percent of the f nal result is recommended, in order to simplify the calculation and analysis.

Developing Mathematical Equations. Mathematical equations are de veloped to describe the operating characteristics of the working substance, and work and energy transfer. Solving for Outputs. Computer programs are used to solv e equations simultaneously, in sequence, or by iteration. The required operating parameters, the outputs, can thus be obtained. Sometimes performance equations are used to link the required outputs with one or two operating parameters. Verification The results of predicted performance during ener gy simulation can be v erif ed against actual measured readings of similar models and operating conditions. According to Scientific Computin (1997), using performance equations to simulate the ener gy use of an HV AC&R system or component is simple, and lumps the capacity, energy use, or other required operating parameters in one performance equation. The disadvantage is that the performance data must be provided to create the performance equations. Physical modeling allo ws more freedom in con f guring the characteristics of the system or equipment. Steady-State and Dynamic Simulation According to its operating characteristics, energy simulation can be classi f ed as steady-state or dynamic simulation. Steady-State Simulation. In steady-state simulation, the relationship between v arious operating parameters within a certain time increment is described by mathematical equations independent of time, such as z  F(x, y, . . . )

SH__ ST__ LG__ DF

(25.6)

where x, y, z  operating variables that are not a function of time. Steady-state simulation al ways simulates relatively long-term system or component characteristics, such as annual or seasonal energy simulation. A time increment of 1 h is typically used for analysis. Within the time increment, the operating parameters are independent of time. Ho wever, the magnitude of the same operating parameter may be different in successive hours.

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25.19

In steady-state simulation, the heat capacity of the w orking substance is often f ar greater than the heat being absorbed and released from the equipment, pipes, ducts, and surroundings when the temperature of the w orking substance fluctuates. Therefore, heat absorbed by the equipment, pipes, and ducts and the heat transfer to or from the surroundings are often ignored. An hour-byhour, year-round energy simulation computer program is an e xample of steady-state simulation.

__RH TX

Dynamic Simulation. In dynamic simulation, the relationship between various operating parameters is described by a mathematical equation that is a function of time z  F(x, y, t, . . .)

(25.7)

where t  time variable. Dynamic simulation is mainly used to simulate short-term characteristics. Its purpose is to analyze the interaction between the control actions and the dynamic response, and the step change of an operating parameter on system performance. A time increment of minutes or even seconds is required to analyze the rapid response of the system during simulation. In dynamic simulation, the heat being absorbed and released from the equipment, instrument, pipe, and ducts is often taken into consideration. The variation of zone temperature and supply v olume f ow caused by a step change of the zone lighting load is an example of dynamic simulation.

Sequential and Simultaneous Simulation When a system comprises man y components linked together by a w orking substance and forms an open circuit, the calculation can be started at a component whose output is the input of the ne xt linked component, and can progress in sequence to the component where the f nal result can be obtained. Such an approach is called sequential simulation. In sequential simulation, the output of a successive component usually does not affect the output of the former component already calculated. If the output of a component does af fect the output of the former component already calculated, a computing loop is used until the calculated output is equal to the assumed value. The fuel input rate to a direct- f red gas heater is an e xample of sequential simulation. Calculation starts from the heating requirement at the w arm air heater . The next calculation is mass f ow and the temperature drop of the comb ustion gas, and f nally the rate of fuel gas supplied to the burner, which is the required result, is then calculated. When the working substance that links the components together forms a recirculating closed loop and many operating parameters are interrelated, all unknown operating parameters should be solv ed simultaneously. A simultaneous simulation is required. The power input to a compressor in a reciprocating compression refrigeration cycle is an example of closed-loop simultaneous simulation.

25.9 ENERGY SIMULATION OF A CENTRIFUGAL CHILLER USING PHYSICAL MODELING The following illustrates the detailed energy simulation processes of a centrifugal chiller for the sake of calculating its po wer input and thus its accumulated annual ener gy use using physical modeling.

System Model The centrifugal chiller has a tw o-stage compressor using HCFC-123 as refrigerant. Its refrigeration cycle is the same as sho wn in Fig. 9.7. The centrifugal compressor is dri ven by a hermetic motor of variable-speed drive. After compression, the refrigerant is dischar ged to a shell-and-tube water-cooled condenser. A f ash cooler or economizer is installed between the condenser and the shell-and-tube f ooded evaporator. Multiple orif ce plates are used as the throttling de vice between

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condensing pressure pcon and intermediate pressure pint, and between pint and evaporating pressure pev. Water from the cooling tower is used as the condenser water. A proportional-integral (PI) DDC panel is used to maintain a nearly constant temperature of chilled w ater as it leaves the evaporator. During part-load operation, the DDC unit modulates the speed of the compressor to reduce both the mass f ow rate of refrigerant m˙r entering the compressor and its corresponding system head, to create a new mass and energy balance. The energy simulation of this centrifugal chiller is considered at steady state within a time increment of 1 h. Sequential simulation, iterations, and perfomance equations are used to calculate the final results. Man y simplifications are adopted to reduce the calculations during simulation.

Operating Parameters Affecting Chiller Energy Performance Operating parameters that affect the energy performance of this centrifugal chiller are as follows: ●

●

Refrigeration load ratio Rload is def ned as the ratio of the operating refrigeration load Qev to the design full-load Qev,d. Ratio Rload affects the rate of heat transfer at the e vaporator and condenser. It also affects the eff ciency of the two-stage compressor. Temperature of condenser water entering the condenser Tce is a function of outdoor wet-bulb temperature To and the performance of the cooling to wer. Temperature Tce is closely related to the condensing temperature Tcon in the condenser, and therefore the pressure lift p  pcon  pev, the temperature lift T  Tcon  Tev, and the power input to the compressor Pin.

Simulation Methodology From Eq. (13.5), power input to the compressor motor Pin,m, in hp, can be calculated as Pin,m 

m˙rWisen 2545cpmecmot

(25.8)

where m˙r  mass f ow rate of refrigerant, lb / h cpmecmot  compression eff ciency, mechanical eff ciency, and motor eff ciency of compressor Work input Win or Wisen, in Btu / lb, is directly proportional to the enthalpy difference of the refrigerant between condensing and e vaporating pressure. To f nd Win, it is necessary f rst to determine the condensing pressure pcon, the condensing temperature Tcon, the e vaporating pressure pev, and the evaporating temperature Tev. So the following information is required: ●

●

●

●

An evaporator model is required to determine Tev at various operating load ratios Rload. A condenser model is required to determine Tcon at various Rload and outdoor wet-bulb temperatures To . Calculation of isentropic work input Wisen to the compressor is required based on the two-stage refrigeration cycle at various Rload and To values. The actual po wer input to the compressor motor based on the compressor model and the annual energy use is f nally calculated.

Evaporator Model

SH__ ST__ LG__ DF

From Eq. (13.7), the rate of heat transfer at the tube surf ace of the e vaporator or the refrigeration load Qev, in Btu / h (W), can be calculated as Qev  AevUo,ev Tev

(25.9)

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25.21

From Eq. (13.8), the overall heat transfer coef f cient based on the outer surf ace area of the copper tubes Uo,ev in Btu / h ft2  °F (W / m2  °C), is given as Uo,ev 

1 1 / (f h o)  AoR f / Ai  Ao / (Aih i)

__RH TX

(25.10)

The boiling coeff cient for HCFC-123 is slightly lo wer than that for refrigerant CFC-11. From Eq. (13.9), for a shell-and-tube f ooded-type e vaporator, the boiling coef f cient of HCFC-123 can be evaluated as Q ev

n

A 

h o  Cb

ev

According to the e xperimental results in Webb and Pais (1991) and Jung and Radermacher (1991), for HCFC-123 in copper tubes with inte grated f ns of 26 f ns / in. (0.98 mm), constant Cb can be taken as 2.5 and e xponential index n is approximately 0.7. F or enhanced surfaces, according to the results of Webb and Pais (1991), ho is 35 percent higher. The copper integrated-f n tubes currently used for evaporators usually have a f n spacing of 19 to 35 f ns / in. (1.3 to 0.73 mm), typically 26 f ns / in. (0.98 mm). The outside diameter of copper tubes varies from 58 to 34 in. (15.9 to 19.1 mm). The ratio of outer surface area to inner surface area Ao /Ai is often between 3 and 4. For a closed-circuit chilled-water system for evaporators with conventional water treatments, the fouling factor Rf can be taken as 0.00025 h  ft2  °F / Btu (0.000044 m2  °C / W). From Eq. (10.9), water-side heat transfer coeff cient hi, in Btu / h  ft2  °F (W / m2 °C), can be calculated as Nu D 

h iDh 0.4  0.023 Re 0.8 D Pr k

(25.11)

From Eq. (13.11), the log-mean temperature dif ference between refrigerant and chilled w ater Tev can be calculated as Tev 

Tee  Tev  (Tel  Tev) ln [(Tee  Tev) / (Tel  Tev)]

(25.12)

The mass flow rate of chilled water m˙w,ev , in lb / min (kg / min), flowing through the copper tubes in the e vaporator usually remains approximately constant during operation. Temperature Tel is often set and reset according to the requirement of the air system and the ener gy efficiency of the system. From Eq. (13.13) Tee  Tel B Q ev / (AevUo,ev) The evaporating temperature can then be determined as Tev 

e B Tel  Tee eB  1

(25.13)

At design load, Qev is equal to the design refrigeration load Qev,d, in Btu / h (kW). In part-load operation, Qev  RloadQev,d and the evaporating temperature at part load Tev,p can be similarly calculated.

(25.14)

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For a shell-and-tube water-cooled condenser, from Eqs. (10.16a) and (13.7), the rate of heat transfer at the condenser Qrej, in Btu / h (kW), or the total heat rejection, can be calculated as Q rej  Q ev 

2545Pcom mot

 AconUo,con Tcon

(25.15)

From Eq. (25.8) po wer input Pcom is known only when condensing temperature Tcon, work input Win, and mass f ow rate of the refrigerant m˙r all have been calculated. Therefore, from Eq. (10.17) Qrej  FrejQev; assume a heat rejection f actor Frej f rst. Then Tcon, Win, and Pcom can be calculated. If the calculated Pcom and Qrej do not equal the assumed v alues, try another Frej until the assumed and calculated values are equal to each other. For comfort air conditioning, Frej usually varies from 1.20 to 1.35. As in the e vaporator model, from Eq. (13.8), the overall heat-transfer coef f cient based on the outer surface area of the condenser Uo,con, in Btu / h  ft2 °F (W / m2  °C), is Uo,con 

1 1 / (fh con)  AconR f / Ai  R g  Acon / (Aih i)

(25.16)

From Eq. (13.11), the log-mean temperature dif ference between the condensing refrigerant and the condenser water Tcon, in °F (°C), is Tcon 

Tcl  Tcon  (Tce  Tcon) ln[(Tcl  Tcon) / (Tce  Tcon)]

(25.17)

From Eq. (13.10), the condensing coeff cient hcon, in Btu / h  ft2  °F (W / m2  °C), is

 Q 1A 

h con  Ccon

1/3

rej con

In a f lmwise condensation shell-and-tube w ater-cooled condenser having 58-in or 34-in.- (15.9- or 19.1-mm-) diameter copper tubes with inte grated f ns and using HCFC-123 as refrigerant, constant Ccon can be taken as 10,500. In a condenser using a well-maintained cooling to wer water with proper water treatment, a fouling factor Rf  0.00025 h  ft2 °F / Btu (0.000044 m 2  °C / W) is recommended. In industrial areas, if a brush cleaning system is installed, Rf  0.0002 h  ft2  °F / Btu (0.000035 m2  °C / W). The operating pressure of HCFC-123, like CFC-11, is lower than atmospheric pressure, so air and other noncondensable gases leak into the e vaporator. The compressor transports them to a higher level and then they accumulate in the condenser. Noncondensable gases reduce the condensing area and raise the condensing pressure. Their effect is similar to that of a gas-side resistance Rg, in h  ft2  °F / Btu (m2  °C / W), at the condenser as follows: Rg  0.00778  0.0173Rload  0.0114Rload2

(25.18)

As in the evaporator model, C is calculated as follows: Tcl  Tce C Q rej / (AconUo,con) Condensing temperature is therefore calculated as SH__ ST__ LG__ DF

Tcon 

e C Tcl  Tce eC  1

(25.19)

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Cooling Tower Model The approach of condenser w ater leaving the cooling to wer Tce  To is mainly in f uenced by the load ratio Rload of the condenser, the conf guration of the cooling to wer, the outdoor wet-bulb temperature To , and the number of transfer units (NTU) of the f ll. In a counter f ow induced-draft cooling to wer using PVC packing f ll, with a condenser w ater f ow rate of 3 gpm per ton ref (0.054 L / s  kW) of heat rejection and a w ater-air ratio of 1.2, if the outdoor wet-b ulb temperature 65 °F To 78°F (18.3 °C To 25.6°C), the temperature of condenser water entering the condenser Tce, in °F (°C), can be roughly estimated by follo wing performance equations:

Tce  To  Tap,d Kload Kwet Kload  0.1  0.9Rload

(25.20)

Kwet  4.8  0.0475To



where Tap,d  approach of cooling tower at design condition, °F (°C) Kload  load factor Kwet  factor considering drop of outdoor wet-bulb temperature

Centrifugal Compressor Model From the refrigeration c ycle shown in Fig. 9.7 b and Eq. (9.33), isentropic work input Wisen  Win, in Btu / lb (kJ / kg), for a two-stage compressor can be calculated as Win  (1  x)(h2  h1)  h4  h3

(25.21)

From Eq. (9.35), the mass f ow rate of the refrigerant m˙r at the condenser, in lb / h (kg / h), is given as m˙r 

Q ev (h 1  h 9)(1  x)

(25.22)

From Eq. (9.30), the fraction of refrigerant evaporated in the f ash cooler x can be evaluated as x

h 5  h 8 h7  h8

(25.23)

On a pressure-enthalpy p-h diagram, the specif c enthalpies of saturated liquid and v apor are functions of temperature and pressure only. In centrifugal chillers used in comfort air conditioning systems, the range between evaporating temperature Tev and condensing temperature Tcon is usually 20 to 120°F (11 to 66°C), and a simple polynomial representation can be used to calculate the specif c enthalpy from the known Tev and Tcon with acceptable accuracy. At the f ash cooler, because the saturated temperature and pressure are interrelated, for simplicity, the intermediate saturated temperature between condensing and e vaporating temperature can be estimated as Tint 

Tev  Tcon 2

(25.24)

According to the thermodynamic properties of HCFC-123 and its p-h diagram, the polynomial expression to calculate the speci f c enthalpy of the saturated v apor of refrigerant HCFC-123 hv, between 20°F (11°C) and 70°F (21°C), in Btu / lb (kJ / kg), is hv  89.7  0.145(Tev  20)

(25.25)

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The polynomial expression used to calculate the speci f c enthalpy of the saturated liquid of HCFC123 hl, between 60°F (15.6°C) and 110°F (43.3°C), in Btu / lb (kJ / kg), is hl  20.2  0.24(Tl  60)

(25.26)

Specif c enthalpy differences h4  h3 and (1  x)(h2  h1) to determine the isentropic w ork can be calculated more simply and with acceptable accuracy according to the corresponding saturated temperature of the gaseous refrigerant at condensing, interstage, and evaporating pressures along the constant-entropy lines of similar prof les in the superheated region. The polynomial expression used to calculate the enthalp y difference of the gaseous refrigerant HCFC-123 between 60 and 120 °F (15.6 and 48.9°C) along the constant entropy line hs, in Btu / lb (kJ / kg), is hs  0.135(Tint  Tev)  0.135(Tcon  Tint)

(25.27)

where Tint  temperature of saturated gaseous refrigerant HCFC-123 at intermediate pressure, °F. In Eq. (25.8) cp is the compression eff ciency of the centrifugal compressor, which is def ned as the ratio of isentropic work to actual work delivered to the gaseous refrigerant during compression. Also, mechanical efficiency mec is defined as the ratio of w ork delivered to the gaseous refrigerant to the w ork input to the compressor shaft. The difference in these tw o w ork inputs is mainly due to the loss in the bearings, and gear train, and during transportation of refrigerant in the centrifugal chiller. For centrifugal compressors operated at a certain speed, mec is considered a fixed value. Mechanical eff ciencies mec for centrifugal chillers manufactured after 1973 can be taken as Without gear train With gear train

0.87 0.85

Motor efficiency mot is a function of motor size and the load ratio of the chiller . Normally, for motor size greater than 125 hp (93 kW), ASHRAE Standard 90.1-1999 recommends an ef ficiency level equal to or greater than 93.0 percent as listed in Sec. 25.4. If the motor is to be operated more than 750 h annually , a high-efficiency motor is most cost-ef fective. The efficiency of a 200-hp (150-kW) motor should be 94.5 percent or greater . If a hermetic motor is used, another 2 to 4 percent of po wer input is required to pro vide the refrigeration capacity to cool the hermetic motor. Compression efficiency cp is often the most influential parameter during annual energy simulation of a centrifugal compressor. Compression efficiency cp is a function of the volume flow rate of refrigerant V˙ rf and the system head or pressure lift pt. Volume flow V˙ rf is closely related to load ratio Rload, and system head is closely related to temperature lift T  Tcon  Tev. In most centrifugal chillers, the chilled w ater lea ving the chiller Tel is usually set at a constant value when the outdoor wet-b ulb temperature is high. As the outdoor wet-b ulb temperature To drops, the condenser w ater entering the condenser Tce and, therefore, the condensing temperature Tcon fall accordingly. It is assumed that when To drops below 70°F (21°C), the reset of Tel offsets the fall of To . Figure 25.2 sho ws the compressor map of a v ariable-speed centrifugal compressor (see Fig. 13.10). In Fig. 25.2, point A is the operating point of the centrifugal chiller at design load conditions (Rload  1) and at a condenser entering w ater temperature Tce  85°F (29.4°C). If Tce remains at 85°F and Rload drops to 0, then the operating point is at point B. At point B, temperature lift T is calculated as follows: (Tcon  Tev)B  (Tcon  Tev)A  (Tel  Tev)A  (Tcon  Tcl)A  (Tcl  Tce)A

SH__ ST__ LG__ DF

(25.28)

In a typical centrifugal chiller, if (Tel  Tev)A  (Tcon  Tcl)A  8°F, (Tcl  Tce)A  10°F, and a typical (Tcon  Tev)A  67°F (similar to scheme A in Sec. 13.6), the operating curves, or required system head, of the compressor at other Tce values can then be plotted on the compression map, as shown in Fig. 25.2.

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FIGURE 25.2 Compressor map of a centrifugal chiller at v arious load ratios Rload and Tce.

The compression eff ciency cp can therefore be calculated by a polynomial re gression (performance equation) based on two variables, Rload and Tce, as follows:

cp  C1  C2Rload  C3R2load  C4Tce  C5T 2ce  C6RloadTce  C7R2loadTce  C8RloadT 2ce  C9R2loadT 2ce

(25.29)

where C1, C2, . . . ,C9  coeff cients. From the manuf acturer’s compressor map and the added operating curves at v arious outdoor wet-b ulb temperatures To , the coeff cients can be determined by a computer program. In a tw o-stage centrifugal compressor, the mean value of compression eff ciency of two individual stages is assumed to equal the o verall compression eff ciency cp for simplif cation.

25.10 ENERGY SIMULATION SOFTWARE DOE-2.1E Energy Simulation Software For energy audit, energy eff ciency improvements, and the selection of ener gy-eff cient HVAC&R systems, Lawrence Berk eley Laboratory and Hirsch Associates, sponsored by the Department of Energy (DOE) collaboratively developed energy simulation computer program DOE-2.1E. It is the most widely used because it includes v arious types of systems, equipment, and energy-eff ciency measures and is friendly to the user . DOE2.1E is an hour -by-hour energy simulation software. The other tw o widely used ener gy simulation softw are programs are TRACE 600, developed by The Trane Company, and HAP E 20-II, developed by Carrier Corporation. Building loads analysis and systems thermodynamics (BLAST) is often used in energy research works.

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TX

According to Scientif c Computing (1996), another DOE-sponsored energy simulation software DOE-2.1E 100 or DOE-2.2 has similar system and equipment capacities to DOE-2.1E. The following table shows their differences:

Equipment time step Sequential-iterative Zone volume f ow automatically calculated Heat recovery Heat recovery variations during cooling

DOE-2.1E

DOE-2.2

1h Sequential Supply Sensible

15 min Sequential-iterative Outside Sensible and latent As a function of To and To

Loads DOE-2.1E has the follo wing primary simulation programs: loads, systems, energy eff ciency measures, and plant. The LOADS programs calculate the sensible and latent components of the hourly heating or cooling loads for each designated conditioned space (zone) in a b uilding under the following conditions: ●

●

●

●

●

●

The software uses 8760 h of actual weather data annually. Indoor temperature is assumed to be kept at a constant value within the hour. The LOADS program is reponsi ve to weather and solar conditions; to schedules of people, lighting, and equipment; to inf ltration; to the heat storage effect of massive walls and roofs; and to the effect of building shade due to solar radiation. The LOADS program considers both e xternal and internal loads. External loads are due to heat transfer through w alls, heat transfer through windo ws, inf ltration through w alls and windo ws, and solar heat gain through windows. Internal loads are due to heat gains from people, lights, and equipment inside the conditioned space. The LO ADS program performs calculations in a hierarchical order: building (system), space (zone), wall, window, or door .The program calculates the e xternal loads for all windo ws and doors on a wall, then for all walls in a space. For the space, LOADS calculates the internal loads, combined with e xternal loads, and gives the total space loads. When all space (zones) loads are summed at a consistent time, this gives the heating or cooling load for the hour of the b uilding (system). During heating and cooling load calculations, the typical f oor of a b uilding is usually divided into man y zones, such as f ve or nine zones; each often has e xternal walls and windo ws facing one, two, or even three orientations except the interior zone. The LOADS program performs load calculations by using weighting factors (transfer function coeff cients). Custom weighting f actors for v arious room con f gurations are also used as an option instead of a preset group of weighting f actors. Detailed calculations of each e xternal or internal load using weighting factors (transfer function method) can be found in Sec. 6.6.

Systems The energy use simulation of the following air and water systems is included: ●

●

SH__ ST__ LG__ DF

●

●

Single-zone constant-volume central systems Multizone constant-volume reheat central systems Multizone variable-air-volume (VAV) reheat central systems Multizone fan-powered VAV central systems

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●

●

●

●

●

●

●

●

●

●

●

25.27

Multizone dual-duct VAV central systems Multizone VAV reheat packaged systems Multizone fan-powered VAV packaged systems Desiccant dehumidif cation packaged systems Evaporative cooling systems Two-pipe fan-coil system Four-pipe fan-coil systems Packaged terminal air conditioner Water-source heat pump systems Ground-source heat pump systems Gas-engine-driven heat pump systems Thermal storage systems

__RH TX

Energy Efficiency Measures Based on Scienti f c Computing (1996), the energy use is calculated and analyzed with the follo wing energy eff ciency measures and controls: ●

●

●

●

Control of the air leaving coil temper ature. Constant, coldest / warmest, reset schedule based on outdoor temperature, and user-def ned reset temperature schedule. Air economizer control. No economizer, air economizer actuated by a f xed temperature, air economizer actuated by comparing outdoor air enthalp y with return air , air economizer actuated by comparing outdoor air temperature with a f xed temperature, and supply air contains f xed amount of outdoor ventilation air. Type of air-to-air heat recovery. Heat exchanger, rotary wheel, and run-around coils. Fan operating schedule and control. Fan is operating according to user-def ned schedule; fan control includes constant-v olume, variable-speed f an inlet v anes, discharge damper , cycles on and off, high and low speeds, as a function of part-load ratio.

Plant Energy simulation of the following heating and cooling plant equipment can be performed either by performance equations or by physical modeling: ●

●

●

●

●

Boilers include multistage resistance-type electric boilers, gas / oil-f red hot w ater boilers, and gas / oil-f red steam boilers. Boilers can be operated only when load is passed from the system, or in standby operation. Boilers can also be operated at a modi f ed ener gy input / heat output as a function of the part-load ratio. Single-effect or double-effect direct-f red absorption chiller is used in which the cooling capacity can be modif ed as a function of refrigeration load required, or as a function of chilled water temperature entering the chiller. Open and hermetic centrifugal chillers modify the capacity and electric po wer input as a function of chilled water temperature entering the chiller, or their part-load ratio. Centrifugal chillers are used with double-b undle condensers in which condenser w ater is used as the hot water for winter heating in perimeter zone. Open and hermetic reciprocating chillers are used with the type of condenser speci f ed, including air condenser or evaporatively-cooled condenser.

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CHAPTER TWENTY-FIVE ●

●

Reciprocating chiller driven by a diesel engine has its COP modif ed as a function of the part-load ratio when the engine is at minimum speed, and COP is modi f ed as a function of chilled w ater temperature entering the chiller when the engine is operated abo ve minimum speed at part load. Cooling towers use the follo wing controls of the condenser w ater temperature leaving the tower: f xed set point; or wet-bulb temperature reset as outdoor wet-bulb temperature drops; or control of the number of cells operating, either actual number of cells needed or all cells. The control of the temperature of condenser water leaving the tower comprises the following devices and setups: Fluid bypass, a three-way valve to bypass condenser water around the cooling tower Cycling, cycling on and off of the tower fans Two-speed fans, that cause the tower fans to cycle between off, low, and high speeds Variable-speed fans to vary the speed of the tower fans

25.11 ASHRAE/IESNA STANDARD 90.1-1999 ENERGY COST BUDGET METHOD The ener gy cost b udget method is an alternati ve to the prescripti ve pro visions in Standard 90.11999 (which is similar to the b uilding envelope trade-off option to the prescripti ve building envelope option in Sec. 3.13). Using the energy cost budget method compliance will be achieved if ●

●

●

All mandatory provisions are met The design ener gy cost does not e xceed the ener gy cost b udget when e valuated in accordance with specif ed provisions The energy eff ciency level of components specif ed in the building design meet or exceed the eff ciency level used to calculate the design energy cost Compliance shall be documented and submitted to the authority ha ving jurisdiction as follo ws:

1. The energy budget for the b udget building design and the design ener gy cost for the proposed design 2. A list of energy-related features that are included in the design 3. The input and output reports from the simulation program including a breakdo wn of energy use of components 4. An explanation of any error messages noted in the simulation program output The simulation program shall be a computer -based program to analyze the ener gy consumption in buildings (such as, DOE-2, BLAST). The simulation shall be performed using hour values. The simulation program shall include the calculation methodologies of the b uilding components being modeled. For details, including calculation of design ener gy cost, calculation of the ener gy cost b udget, and HVAC&R systems, refer to Standard 90.1-1999.

REFERENCES

SH__ ST__ LG__ DF

Akbari, H., Konopacki, S. J., Lister, L. D., and Debaillie, L. P., Energy End-Use Characterization at Fort Hoot, Texas, ASHRAE Transactions, 1996, Part II, pp. 724 – 733. ASHRAE, ASHSRAE Handbook 1999, HVAC Applications, ASHRAE Inc., Atlanta, GA, 1999. ASHRAE / IES, ASHRAE / IES Standard 90.1-1989, Energy Eff cient Design of New Buildings Except New Low-Rise Residential Buildings, ASHRAE Inc., Atlanta, GA, 1989.

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ASHRAE / IESNA Standard 90.1-1999 Energy Standard for Buildings Except Low-Rise Residential Buildings, ASHRAE Inc., Atlanta, GA, 1999. Baxter, V., Fischer, S., and Sand, J. R., Global Warming Implications of Replacing Ozone-Depleting Refrigerants, ASHRAE Journal, no. 9, 1998, pp. 23 – 30. Better Late Than Never: Energy Eff ciency Is a Federal Priority, Air Conditioning, Heating & Refrigeration News, Aug. 18, 1997, p. 16. Bobenhausen, W., and Lahiri, D., HVAC Design for Green Buildings, HPAC, no. 2, 1999, pp. 43 – 50. Bynum, R., The Current Situation, Engineered Systems, no. 12, 1998, pp. 32 – 36. Clark, D. R., Hurley, C. W., and Hill, C. R., Dynamic Models for HVAC System Components, ASHRAE Transactions, 1985, Part I, pp. 737 – 751. Cole, R. J., and Larsson, N., Green Building Challenge ‘98, ASHRAE Journal, no. 5, 1998, pp. 20 – 23. Cox, J. E., and Miro, C. R., Reducing the Federal Energy Bill, ASHRAE Journal, no. 12, 1998, p. 24. Cox, J. E. and Miro, C. R., Montreal and Kyoto Protocols’ Relevance to HCFs, ASHRAE Journal, no. 1, 1999, p. 16. DOE / EIA Nonresidential Building Energy Consumption Survey: Commercial Building Consumption and Expenditures 1989, DOE / EIA 0318 (89), 1992. Dolan, W. H., Gas Cooling for the Commercial Sector — Present and Future Perspective, ASHRAE Transactions, 1989, Part I, pp. 968 – 971. Gabel, S. D., Carmichael, L., and Shavit, G., Automated Control in Response to Real-Time Pricing of Electricity, ASHRAE Journal, no. 11, 1998, pp. 26 – 29. Ginsberg, M., and Parker, S., Meeting Federal Mandates, Engineered Systems, no. 9, 1996, pp. 52 – 60. Gottfried, D. A., Implications of U.S. Electricity Deregulation, HPAC, no. 5, 1997, pp. 55 – 58. Gottfried, D. A., Schoichet, E. A., and Hart, M., Green Building Environmental Control: A Case Study, HPAC, no. 2, 1997, pp. 71 – 78. Hansen, S. J., Performance Contracting: Fantasy or Nightmare? HPAC, no. 11, 1998, pp. 71 – 76. Hicks, T. W., and Clough, D. W., Building Performance with the ENERGY STAR Label, HPAC, no. 10, 1997, pp. 49 – 54. Hoggard, J., Shoot for the Star, Engineered Systems, no. 1, 2000, p. 38. Jung, D. S., and Radermacher, R., Prediction of Heat Transfer Coeff cients of Various Refrigerants during Evaporation, ASHRAE Transactions, 1991, Part II, pp. 48 – 53. Lawrence Berkeley Laboratory, DOE-2 Users Guide, Version 2.1, DOE National Technical Information Service, Springf eld, VA, 1980. Mahoney, J. W., and Weiss, D. W., Performance Contracting: A Guaranteed Solution, HPAC, no. 3, 1997, pp. 61 – 65. Miller, D. E., A Simulation to Study HVAC Process Dynamics, ASHRAE Transactions, 1982, Part II, pp. 809 – 825. Morrison, D., Green Design Focus: Energy Eff ciency and Environmental Responsibility, HPAC, no. 2, 1998, pp. 66 – 76. Norford, L. K., Englander, S. L., and Wiseley, B. J., Demonstration Knowledge Base to Aid Building Operators in Responding to Real-Time-Pricing Electricity Rates, ASHRAE Transactions, 1998, Part I A, pp. 91 – 103. O’Neal, D., and Kondepudi, S. N., Demonstrating HVAC System Performance through System Simulation, ASHRAE Transactions, 1986, Part II B, pp. 116 – 129. Park, C., Bushby, S. T., and Kelly, G. E., Simulation of a Large Off ce Building System Using the HVACSIM  Program, ASHRAE Transactions, 1989, Part I, pp. 642 – 651. Parker, D. S., Fairey, P. W., and McIlvaine, J. E. R., Energy Eff cient Off ce Building Design for Florida’s Hot and Humid Climate, ASHRAE Journal, no. 4, 1997, pp. 49 – 57. Quiriconi, M. A., Dean, M. L., and Litven, N. A., Failing Grade for Ailing Equipment, Engineered Systems, no. 12, 1997, pp. 72 – 78. Schiess, K., RTP  TES  ? Engineered Systems, no. 10, 1998, pp. 102 – 110. School Saves Nearly $500,000 Annually on Utility Bills, HPAC, no. 1, 1998, pp. 12 – 19. Scientif c Computing, HVAC System Design: Energy Simulation Software Review, Engineered Systems, no. 2, 1996, pp. 42 – 64.

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Scientif c Computing, Up for Review (Again), Engineered Systems, no. 1, 1998, pp. 76 – 84. Silver, S. C., Jones, J. W., Peterson, J. L., and Hunn, B. D., CBS / ICE: A Computer Program for Simulation of Ice Storage Systems, ASHRAE Transactions, 1989, Part I, pp. 1206 – 1213. Todesco, G., Super-Eff cient Buildings: How Low Can You Go? ASHRAE Journal, no. 12, 1996, pp. 35 – 40. Waintroob, D., Retail Electric Competition, Engineered Systems, no. 5, 1998, p. 48. Wang, S. K., Air Conditioning, vol. 4, Hong Kong Polytechnic, Hong Kong, 1987. Warwick, W. M., Into the Looking Glasses: Utility Industry Restructuring and You, HPAC, no. 5, 1997, pp. 47 – 53. Warwick, W. M., Top 10 Lessons from Competitive Power Purchases, HPAC, no. 8, 1998, pp. 77 – 82. Webb, R. L., and Pais, C., Pool Boiling Data for Five Refrigerants on Three Tube Geometries, ASHRAE Transactions, 1991, Part I, pp. 72 – 78. Wong, S. P. W., and Wang, S. K., System Simulation of the Performance of a Centrifugal Chiller Using a Shelland-Tube Type Water-Cooled Condenser and R-11 as Refrigerant, ASHRAE Transactions, 1989, Part I, pp. 445 – 454. Zaidi, J. H., and Howell, R. H., Energy Use and Heat Recovery in Water-Loop Heat Pump, Variable-AirVolume, Four-Pipe Fan-Coil, and Reheat HVAC Systems: Part 1 and Part 2, ASHRAE Transactions, 1993, Part II, pp. 13 – 39.

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AIR CONDITIONING SYSTEMS: SYSTEM CLASSIFICATION, SELECTION, AND INDIVIDUAL SYSTEMS 26.1 CLASSIFICATION OF AIR CONDITIONING SYSTEMS 26.1 Basic Approach 26.1 Air Conditioning Systems Classification 26.2 Air, Heating, and Cooling Systems Designation 26.2 26.2 AIR CONDITIONING SYSTEM, SUBSYSTEM, AND MAIN COMPONENTS SELECTION 26.3 Selection Levels 26.3 Requirements Fulfilled during Selection 26.4 Applications and Building Occupancies 26.4 System Capacity 26.5 Indoor Air Quality 26.5 Zone Thermal Control and Sound Problems 26.6 Energy Efficiency 26.7 Fire Safety and Smoke Control 26.7 Space Limitations 26.8 Maintenance 26.8 Initial Costs 26.8

26.3 INDIVIDUAL AIR CONDITIONING SYSTEMS 26.8 Basics 26.8 Advantages and Disadvantages 26.9 26.4 ROOM AIR CONDITIONING SYSTEMS 26.9 Equipment Used in Room Air Conditioning Systems 26.9 Configuration and Cooling Mode Operation 26.11 Energy Performance and Energy Use Intensities 26.11 Controls 26.12 Features 26.12 System Characteristics 26.12 26.5 PACKAGED TERMINAL AIR CONDITIONING SYSTEMS 26.13 Equipment Used in Packaged Terminal Air Conditioning Systems 26.13 Heating and Cooling Mode Operation 26.13 Minimum Efficiency Requirements 26.14 System Characteristics 26.15 Applications 26.15 REFERENCES 26.15

26.1 CLASSIFICATION OF AIR CONDITIONING SYSTEMS Basic Approach Because of the v ariation in b uilding occupancies as well as the outdoor weather in v arious locations, and because of dif ferent operating requirements from f acility developers, owners, and users, air conditioning or HV AC&R systems are usually designed, installed, and operated in dif ferent types and configurations with different system characteristics to meet these requirements.

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The purpose of classifying air conditioning systems is to distinguish one type from another and to provide a background for the selection of an optimum air conditioning system based on requirements. As each air condition system consists of air , water, heating, and refrigeration systems, the classifica tion of air conditioning systems and is often mixed with the classification of air systems and ater and refrigeration systems. If a designer cannot properly classify an air conditioning system and distinguish it from others, it will be difficult for him or her to select an appropriate system for the client During the classification of air conditioning systems the following points should be considered: ●

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The classification of air conditioning systems or HVAC&R systems, should include the primary aspects of air systems as well as heating and cooling systems if possible, as the air system directly affects the control of the indoor thermal environment and indoor air quality. The system and the primary equipment used should be compatible with each other . For example, the primary equipment in a unitary packaged system is the packaged unit. System classification should mainly be based on practical applications. or example, as the indoor air quality becomes one of the primary criteria used to select an air conditioning system, it must be considered whether an “all-w ater system” without outdoor v entilation air supply can exist. System classification should be simple and each of the air conditioning systems should be clearly different from the others.

Air Conditioning Systems Classification As discused in Chap. 1, air conditioning systems can be classified currently into eight cat gories according to their configurations and operating characteristics ●

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Individual room air conditioning systems or simply individual systems Evaporative cooling air conditioning systems Desiccant-based air conditioning systems or simply desiccant systems Thermal storage air conditioning systems or simply thermal storage systems Clean room air conditioning systems or simply clean room systems Space conditioning air conditioning systems or simply space systems Unitary packaged air conditioning systems or simply packaged systems Central hydronic air conditioning systems or simply central systems

First, as discussed in Sec. 1.5, the individual, space, packaged, and central systems together had more than 98 percent of floor area both in commercial uildings in 1992 and in air conditioned homes in 1991 in the United States. Clean room and desiccant-based air conditioning systems are often processed air conditioned systems. Except where deep-well w ater or an air economizer is available, evaporative cooling systems are the cheapest cooling systems and are widely used in arid southwestern areas of the United States. Thermal storage systems ha ve a quite dif ferent water system and operating characteristics from a central system. Second, there are many specific air condi tioning systems used in man y industrial applications such as air conditioning used for te xtile mills with air washers, and also new air conditioning technology will be invented and developed. As soon as their importance and their distinct system characteristics are recognized by the HV AC&R industry, a new category of air conditioning system should be added. Air, Heating, and Cooling Systems Designation SH__ ST__ LG__ DF

The title of air conditioning systems listed in later sections of this chapter and succeeding chapters, such as VAV reheat rooftop packaged systems, also designates the air, heating, and cooling systems:

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26.3

A packaged system always has a cooling system that uses the DX coil to cool air directly . A desiccant-based system usually uses the DX coil as supplementary cooling. An indi vidual system also uses a small, self-contained, factory-assembled refrigeration system that has a DX coil to cool air. A central system has a cooling system that uses chilled w ater as a cooling medium to cool air indirectly. A thermal storage system is al ways a central system that uses chilled w ater or brine as the cooling medium. As discussed in Sec. 9.18, for DX coils in indi vidual systems, rotary compressors are the most widely used. F or DX coils in packaged systems, reciprocating and scroll compressors are most widely used. Scroll compressors are gradually replacing the reciprocating compressors in packaged systems because of their higher ener gy efficien y. For chillers in central and space systems, centrifugal and screw compressors are widely used in lar ge chillers whose cooling capacities are greater than 75 tons (264 kW); and scre w, scroll, and reciprocating compressors are widely used in medium-size and small chillers. Screw and scroll types are also gradually replacing the reciprocating compressors in small and medium-size chillers. As discussed in Sec. 8.1, hot water, heat pump, and direct-fired arm air furnace heating systems were installed in commercial and residential b uildings with a floor area of both about 75 percen in the early 1990s. In packaged systems, direct-fired arm air furnaces and heat pumps are often used. For individual systems, either heat pumps or electric resistance heaters are often used. In central and clean space systems, hot water heating systems are most widely used. As discussed in Sec. 20.9, air systems can be classified into three cat gories:

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Constant-volume (CV) outdoor recirculating air mixing systems Variable-air-volume (VAV) outdoor recirculating air mixing systems Dedicated ventilation and space recirculating systems Only air systems of the space air conditioning systems are dedicated v entilation and space circulating systems. Individual, packaged systems in residentials and clean room air conditioning systems are mostly CV air systems. All the other air conditioning systems may be either CV or VAV systems and will be so designated.

26.2 AIR CONDITIONING SYSTEM, SUBSYSTEM, AND MAIN COMPONENTS SELECTION Selection Levels During the design of a ne w or retrofit system air conditioning system, subsystems, and main components selection is performed mainly on three levels: ●

●

Air conditioning system le vel. This level deals with such aspects as whether a VAV packaged air conditioning system or a fan-coil air conditioning system should be selected. Air system, water system, central plant cooling and heating system, and control system level. For a VAV packaged system, it must be determined whether the system is to be a single-zone VAV air mixing system, a multizone VAV reheat air mixing system, or a f an-powered VAV air mixing system. In a packaged system, its cooling system is always a DX system with either reciprocating or scroll compressors. Its heating system is either a direct-fired arm air heating system or, in locations where the unit electric rate is low, a heat pump system or an electric heating system. For a space air conditioning system, its air system is al ways a dedicated v entilation and space recirculating system. F or the w ater system, the choice is between a plant-b uilding loop with variable fl w in the b uilding loop and a plant-through-b uilding loop. Either a hot w ater heating system or an electric heating system can be selected for a f an-coil system. A f an-coil system

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requiring large tonnage of refrigeration load often uses centrifugal or scre w chillers, and a smalltonnage fan-coil system usually uses screw and scroll chillers. Except for room air conditioning systems and packaged systems of small size, usually an energy management and control system (EMCS) with direct digital control (DDC) is cost-ef fective. For a specific project the control zones, control points, and selection of generic and specific con trol functions should be optimum. Main system components le vel. For air systems this in volves the selection of main components, such as: A forward-curved centrifugal, an airfoil centrifugal, and a vane-axial supply fan. A supply-relief fan combination or a supply and return fan combination. A CO2-based demand control ventilation, or a mixed plenum pressure ventilation control. An air economizer or a water economizer, a thermal wheel, run-around coils, or a heat pipe air-toair heat recovery system. An adjustable frequenc y variable-speed drive, inlet vanes, or an inlet cone duct-static pressure control. A mixing fl w, stratified mixing f w, or displacement fl w pattern and supply outlets and return inlets layouts.

For water systems, the main system components level involve selection of A plant-building loop or a plant-through-building loop In a plant-building loop, whether balance valves are necessary In a plant-through-building loop, whether a bypass throttling fl w, distribution pumping, or variable fl w An open expansion tank or a diaphragm expansion tank For refrigeration systems, main system components level involves the following: For heat rejection systems, whether an air -cooled, water-cooled, or e vaporatively cooled condenser is selected For a centrifugal chiller , whether a tw o-stage or a three-stage impeller , and whether a v ariablespeed drive or an inlet vane’s capacity modulation is selected For reciprocating or scroll refrigeration systems, whether a v ariable-speed, two-speed, cylinder unloader, or on / off capacity control is selected

Requirements Fulfilled during Selection To properly, effectively, and energy-efficiently design an air conditioning (H AC&R) system, the designer, facility owner, or developer should collaborate to select the system according to the following requirements:

Applications and Building Occupancies

SH__ ST__ LG__ DF

When air conditioning systems are used for dif ferent applications or b uilding occupancies, they need different design criteria, operating hours, and different system characteristics. F or example, a constant-volume central system is always used for a class 10 clean room to f abricate semiconductor wafer. When the clean room is in operation, adequate clean air must be pro vided to maintain unidirectional fl w to prevent the contamination of semiconductor w afers by submicrometer-size particulates. A case study of a clean-space air conditioning system for a class 10 clean room is discussed in Chap. 30.

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A four-pipe fan-coil system is the most widely used air conditioning system for guest rooms in luxury hotels. This is because a four -pipe fan-coil system can pro vide individual temperature and fan speed controls as well as a positi ve supply of adequate outdoor v entilation air. The fan coil can be turned of f con veniently when the room is not occupied. It is isolated acoustically from adjacent rooms. The most anno ying space maintenance tasks, such as changing the f ilters and periodic inspection and maintenance of the f an coil, can be done when the guest room is unoccupied. Specific design criteria are a ways related to applications. Specific design criteria usually dictat the type of air conditioning system that should be selected. F or instance, a high-precision constanttemperature room maintained at a temperature of 70  0.2°F (21  0.1°C) within a limited w orking space needs a constant-v olume central system with electric terminal reheat. Electric reheat can be modulated more precisely than an y other type of heating system. A central system can maintain a more uniform discharge air temperature at the AHU than DX coil packaged units. Usually, such a high-precision constant-temperature room is a clean space. A constant-v olume system is al ways preferable to a VAV system for such rooms.

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System Capacity System capacity is closely related to the selection of an air conditioning system. F or a single-story small retail shop, a constant-volume packaged system is often chosen. If the conditioned space is a large indoor stadium with a seating capacity of 70,000 spectators, a single-zone VAV central system with man y AHUs using peripheral nozzles at dif ferent le vels to supply conditioned air is often selected. Indoor Air Quality Indoor air quality (IAQ) depends mainly on the minimum v entilation control; the removal of bacteria, particulates, irritating vapors, and toxic gases; preventing dampness; and proper maintenance of the HVAC&R system. Among various minimum ventilation controls, the dedicated ventilation system with CO 2-based demand-controlled v entilation guarantees the minimum outdoor v entilation rate specified by ASHRAE Standard 62-1999 and is energy-efficient. The dedicated ventilation system, the mixed plenum pressure control, and the outdoor air injection fan are good minimum ventilation control systems with satisfactory performance. All guarantee the specified constant- olume fl w of minimum v entilation rate at the outdoor air intak e in the AHU or PU. The supply and return f an tracking systems, and the direct measurement of outdoor air intak e in the AHU and PU are considered as poor minimum v entilation control systems. As the volume fl w rate of the outdoor v entilation air is usually less than one-fourth that of the total supply air , it is difficult to measure it correctly by measuring the olume fl w rate of supply and return air . This may result in an insuf ficient outdoor entilation intak e, especially during part-load operation in VAV systems. Air economizer cycle and purging operation prior to the occupied period on weekday mornings both improve IAQ because of the introduction of all outdoor air to the conditioned space. As discussed in Secs. 15.16 and 24.5, to remove particulates, bacteria, viruses, irritating vapors, and toxic gases from the conditioned space, air filtration and gaseous adsorption and chemisorptio in the AHUs or PU can be rated into the following four grades: ●

●

Low-efficien y air filters are used to rem ve dusts of size 3 to 10 mm, such as, spores, pollens, and textile fibers. anel filters made of wire mesh and synthetic and glass fibers are often used the filtration medium Medium-efficien y filters are used to rem ve dusts of siz e 1 to 3 mm, such as, bacteria, automotive-emissions, and fine dust for the protection of coils and air distri ution systems and to pre vent

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the nutrition of microbial growth. Pleated filters with synthetic and glass fibers as the filt s media, are also used. High-efficien y filters are used to rem ve dusts of size 0.3 to 1 mm, such as, bacteria, cooking oil fumes, and tobacco smoke. Bag or cartridge filters with synthetic and glass fibers of submicrom ter diameter are used as filter media. In ultrahigh-efficien y filtration HEPA has a DOP ef ficien y of 99.97 percent, and ULPA filter have a DOP efficien y of 99.999 percent. HEPA and ULPA filters are used to rem ve viruses, carbon dust, combustion smoke, and radon progeny. They are widely used in clean rooms and clean space. To remo ve objectionable odors, irritating v apors, and toxic gases of molecular size between 0.003 and 0.006 m, activated carbon adsorbers and chemisorbers are often used. When high-efficien y filters HEPA and ULPA filters activated carbon adsorbers, and chemisorbers are used, they must be protected by medium-efficien y filters to xtend their service life.

Zone Thermal Control and Sound Problems Zone thermal control comprises the temperature and humidity control of the control zone, which affects directly the thermal comfort of the occupants in the conditioned space. In addition to the heating and cooling capacities needed to meet the space load requirements, the scope of control and control quality can be assessed in the following three areas: Scope. Today zone control is the basis of thermal indi vidual control. Usually, a room or a conditioned space with a floor area often between 100 and 1000 f 2 (10 and 100 m 2) is considered as a controlled zone. For each control zone, there is often a sensor, a controller, and a corresponding terminal device (a VAV box, a fan coil, or a water-source heat pump), to control the space temperature and space humidity according to the requirements. For a large open office additions of cubical partitions, and use of supply outlets that can adjust the volume fl w or the fl w direction to regulate are means to control the space temperature and humidity that cover a smaller control zone. Recently, a desktop task conditioning system has been de veloped from an underfloor up ard fl w space air diffusion system which can provide zone control for an individual worker in the conditioned space, as discussed in Sec. 18.9. Volume Flow Control and Supply T emperature Control. For cooling, the zone control quality depends on the volume fl w control of cold supply air of a VAV box to meet the space sensible load reduction at part-load operation or the reduction of the chilled w ater fl w to a cooling coil. F or heating, the zone control quality depends on the capacity control of a reheating system or a perimeter heating system. Modulating the volume fl w of cold supply air during part-load operation only slightly reduces the space relative humidity. However, reducing the chilled w ater fl w to a cooling coil may signifi cantly increase the space relative humidity especially when the fan coil is oversized.

SH__ ST__ LG__ DF

Control Modes. More often, the zone control mode is a DDC proportional-inte gral (PI) modulation control without of fset. It can be also a tw o- or three-fan speed control (small f an in fan coil or in water-source heat pump) plus on / off control of the w ater fl w. A DDC with PI or proportionalintegral-derivative (PID) control al ways has a better control quality because it has no of fset except due to instrument inaccuracy. Sound problems often are the result of improperly designed systems and poorly selected f ans, pump, and compressors. Noise always is very annoying. Some air conditioning systems such as individual systems and space conditioning systems ha ve inherent sound problems because the room air conditioners and w ater-source heat pumps are often installed directly in the conditioned space. Many fan-coil units are hung in the ceiling plenum directly abo ve the conditioned space where the recirculating air inlet provides a short and convenient sound transmission path. The estimated sound levels in the occupied zone for various air conditioned systems are as follows:

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Condition

Air conditioning system

Equipment

Estimate

Poor Acceptable Good Excellent

Room air conditioning system Four-pipe fan-coil system VAV reheat central system Single-zone VAV central system

Room air conditioner Fan-coil unit (400 cfm, 189 L / s) AHU, VAV-box AHU

45 to 50 dBA 35 to 40 dBA 25 to 35 NC 15 to 20 NC

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Energy Efficiency As discussed in Sec. 25.6, each type of air conditioning system has its o wn system characteristics and ener gy use intensity (EUI), expressed in kBtu / h  ft2  °F (kWh / m2  °C), as sho wn in Table 25.1. Boilers, furnaces, compressors, fans, and pumps are ener gy users. To properly select an air conditioning system, the designer should kno w the ener gy efficien y of main ener gy users as well as the estimated EUI of the selected air conditioning system. Clearly distinguish the type and configuration of ene gy users in an air conditioning system: Heating equipment Direct-fired arm air furnace Hot water boilers, or condensing hot water boilers Refrigeration compressors: reciprocating, scroll, screw, or centrifugal Water pumps: centrifugal hot water pumps, centrifugal chilled water pumps, or centrifugal condenser water pumps Fans: forward-curved / aerofoil centrifugal supply fans, centrifugal / plug / axial return fans, or axial relief f ans, air-cooled axial condenser f ans, evaporatively-cooled centrifugal condenser f ans, or propeller cooling tower fans, and exhaust fans Motors: high-efficien y motors Estimate the combined efficien y of various energy users and the EUI of the selected system. Consider energy conservation measures (ECMs), such as VAV systems, air or w ater economizers, heat recovery devices, variable-speed drives, and occupancy sensors, for the selected air conditioning system to improve its energy efficien y.

Fire Safety and Smoke Control In case of a b uilding fire in highrise uildings, as discussed in Sec. 22.12, the water fl w indicator of the automatic sprinkler system actuates smok e control systems and stairwell pressurization systems, and the b uilding automation system (B AS) coordinates v arious air systems to pro vide a smokefree escape route through the emergency exits and stairwells to the outdoors, as follows: ●

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Stairwell pressurized systems are actuated. Smoke is e xhausted to outdoors from the f ire floor (smok e zone) through the smok e e xhaust system. Supply outdoor air to floor or floors a ve and below the fire floor or adjacent to the sm e zone to pressurize them to prevent smoke contamination.

For air systems with air economizers, the exhausted volume fl w from the fire floor and the olume fl w to pressurize floors or control zones adjacent to the smo e zone are approximately equal to the supply volume fl w rate of the air economizer cycle. For air systems with dedicated outdoor ventilation systems, or with a local outdoor ventilation air intake (such as air systems in the fan-coil, water-source heat pump systems, or room air conditioning

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systems), a smoke exhaust system of about 6 ach (air changes per hour) should be installed; and the dedicated outdoor ventilation system can be used to pressurize the adjacent control zones according to local and federal fire code requirements

Space Limitations Space limitations specified by the architect or acility owner also influence the selection of the ai conditioning system. For instance, for a high-rise building of more than 30 stories, if rooftop space is not available for the penthouse of AHUs and other mechanical equipment, or if there is no space left for supply and return duct shafts, a floo -by-floor AHU central system may be the practical choice. Because of the a vailable ceiling space, low-velocity duct systems are often used in highceiling industrial buildings to save air-transporting energy.

Maintenance A central system with AHUs, a few water-cooled centrifugal chillers, and cooling towers needs less maintenance w ork than a packaged system with man y rooftop air -cooled reciprocating packaged units. A VAV reheat central system needs less maintenace w ork in the f an and plant rooms than a fan-coil system, which often requires much maintenance w ork in the ceiling space directly abo ve the conditioned space.

Initial Costs Initial cost and operating costs (mainly energy cost) are always primary factors that affect the selection of an air conditioning system, especialy for a de veloper who sells b uildings with installed air conditioning systems. The initial cost of the air conditioning (HV AC&R) system in a b uilding, expressed in $ / ft2 ($ / m2), depends on the b uilding occupancies, system configurations size of the building, and capabilities of specific systems. According to Gladstone and Humphre ys (1995), the initial costs of an HVAC&R system for school buildings are as follows: System type

$ / ft2

Single-zone, constant-volume packaged systems Multizone reheat constant-volume central system Multizone dual-duct central systems

5.25 6.60 8.25

Generally, the more complex an air conditioning system becomes and the more features it has, the higher will be the initial cost. Among all these, the Class 1 clean room central system is the most expensive.

26.3 INDIVIDUAL AIR CONDITIONING SYSTEMS Basics

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As discussed in Sec. 1.3, an individual room air conditioning system or simply an individual system uses a self-contained, factory-made packaged air conditioner to serve an individual room. It is ready to use after electric cable and necessary w ater drainage are connected. Indi vidual systems al ways use a DX coil to cool the air directly . Individual systems can be subdi vided into the tw o following air conditioning systems:

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Room air conditioning (RAC) systems Packaged terminal air conditioning (PTAC) systems

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Advantages and Disadvantages Individual air conditioning systems have the following advantages: ●

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There are no supply, return, or exhaust ducts. Individual air conditioning systems are the most compact, fl xible, and lower in initial cost than others, except portable air conditioning units. Building space is saved for mechanical rooms and duct shafts. It is easier to match the requirements of an individual control zone. They are quick to install.

Individual systems have the following disadvantages: ●

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●

●

Temperature control is usually on / off, resulting in space temperature swing. Air filters are limited to coarse or l w-efficien y filters Local outdoor ventilation air intake is often affected by wind speed and wind direction. Noise level is not suitable for critical applications. More regular maintenance of coils and filters is required than for packaged and central systems

Individual room and unitary packaged air conditioning systems are both self-contained, factorymade, packaged systems using DX coils for cooling. Ho wever, individual systems are single-zone units, whereas a packaged system can be either a single-zone or multizone unit. Indi vidual systems have no supply and return ducts. Also, individual systems pro vide far smaller heating-cooling capacities, poorer outdoor v entilation intake, lower-efficien y filters simpler control system, and far less ECMs than packaged systems. Room air conditioners are often windo w-mounted. Most of the packaged units are rooftop units. As discussed in Sec. 1.5, individual air conditioning systems were found in about 26 percent of 66 million air conditioned homes and packaged systems in 74 percent in 1991 in the United States. Also, individual systems were found in 22 percent of commercial b uildings and packaged systems in 48 percent in 1992 in the United States.

26.4 ROOM AIR CONDITIONING SYSTEMS Equipment Used in Room Air Conditioning Systems Room air conditioning systems are the most widely used indi vidual air conditioning system in the United States. Window-mounted or through-the-w all types of room air conditioners; room heat pumps (RHPs) to pro vide both heating in winter and cooling in summer; and split air conditioners (SACs) with an outdoor condensing unit and an indoor air handler , for the sake of greater fl xibility in location of the air handler as well as isolation of compressor noise outdoors are three kinds of equipment used in room air conditioning systems. A room heat pump has a similar configuration to room air conditioner e xcept a four -way reversing valve is added to change from cooling mode to heating mode, and vice versa. According to Mahoney (1996), annual shipments of window-mounted

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and through-the-wall room air conditioners increased from 2.8 million units in 1991 to 4.3 million in 1995 in the United States. The cooling capacity of currently a vailable room air conditioners ranges between 5000 and 34,000 Btu / h (1.5 and 10 kW). Refrigerant HCFC-22 and alternati ve refrigerants HFC-407C and HFC-410A are the refrigerants used now and after 2020. Electric resistance heating is often used in room heat pumps to supplement the winter heating when the outside weather is cold. According to Rosenquist (1999), the average service life of a room air conditioner is 12.5 years.

Condensing coil

Outdoor fan

Outdoor ventilation air intake

Cooling air

Indoor fan

Rotary compressor

Coarse air filter

Cooling air

Motor

Return air

DX coil

Supply air

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( a)

(b)

Return air FIGURE 26.1 A window-mounted room air conditioner: (a) Sectional view; (b) front view.

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Configuration and Cooling Mode Operation Figure 26.1 shows a window-mounted room air conditioner. The cabinet is di vided into indoor and outdoor compartments which are separated by insulated w all to reduce the heat transfer . The DX coil and indoor fan are in the indoor compartment. The outdoor compartment contains the compressors, condensers, outdoor f an, capillary tube, and f an motor. The f an motor often has a doubleended shaft which drives both fans. Return air from the conditioned space f ows through a coarse air f lter, is cooled and dehumidif ed in a DX coil, and then enters the inlet of the indoor f an. In a room air conditioner , the indoor fan is a forw ard-curved centrifugal f an. The conditioned air is pressurized in the impeller and forced through the air passage that leads to the supply grille. The conditioned air is then supplied to the conditioned space to offset the space cooling load. Outdoor air is e xtracted by the propeller f an (outdoor f an) and forced through the condensing coils, in which hot gaseous refrigerant is condensed to liquid refrigerant. During condensation, condensing heat is released to the outside through the cooling air . A portion of outdoor v entilation air is extracted by the indoor f an and mixed with the return air . The opening of the outdoor v entilation air intake is adjustable. According to ASHRAE Handbook 1996, HVAC Systems and Equipment, RAC is rated under the following standard conditions: Evaporating temperature Compressor suction temperature Condensing temperature Liquid temperature Ambient temperature

45°F (7.2°C) 55°F (12.8°C) 130°F (54.4°C) 115°F (46.1°C) 95°F (35°C)

The schematic diagram and refrigeration cycle of a room heat pump system are sho wn in Fig. 12.1. The heating and cooling modes operation are similar to those in a rooftop heat pump packaged system, discussed in Sec. 12.2 and shown in Fig. 12.3. Energy Performance and Energy Use Intensities According to Rosenquist (1999), the room air conditioners b uilt in the early 1980s often had an EER of 7.3 and greater . In 1999, the maximum rotary compressor ef f ciencies range from 10.7 to 11.1 EER. Many RAC manufacturers are using slit-type aluminum f ns and grooved or rif ed copper refrigerant tubing in the heat e xchange coils. Incorporating a lar ge surface area to capacity ratio, it is able to yield rotary compressor eff ciencies of 11.1 to 11.3 EER (3.25 to 3.31 W / W). Permanent split capacitor (PSC) motors are the predominant f an motor used in RACs. PSC motors have an eff ciency between 55 and 70 percent. Some RA C fan motors use shaded pole motors with an eff ciency of 30 to 40 percent. ASHRAE / IESNA Standard 90.1-1999 specif es the minimum eff ciency requirement for room air conditioners with louvered sides and room air conditioner heat pumps with louv ered sides as follows: Equipment type Room air conditioner

Room air conditioner heat pump with louvered sides

Capacity Qrc Btu / h

Minimum eff ciency

Eff ciency as of 10 / 29 / 2001

Qrc  6000 6000  Qrc  8000 8000  Qrc  14,000 14,000  Qrc  20,000 Qrc  20,000 Qrc  20,000 Qrc  20,000

8.0 EER 8.5 EER 9.0 EER 8.8 EER 8.2 EER 8.5 EER 8.5 EER

9.7 EER 9.7 EER 9.8 EER 9.7 EER 8.5 EER 9.0 EER 8.5 EER

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For other types of room air conditioners and heat pumps, refer to Standard 90.1-1999. According to DOE / EIA Commercial Buildings Consumption and Expenditures 1995, individual systems using mainly room air conditioners in commercial b uildings in 1995 in the United States had an annual ener gy consumption of 43.2 kBtu / ft2  yr (12.7 kWh / ft2  yr or 137 2 kWh / m  yr), which is only less than the EUI for central systems and single-zone constant-v olume packaged systems, as shown in Table 25.1. Higher energy use in RAC systems is mainly due to the lower eff ciency of the compressor and fan and the motor and drive combined eff ciencies.

Controls RACs usually use a tw o-stage thermostat to separate cooling mode and heating mode operations. One of the RA C controls selects the operation mode. Another control adjusts the set point and c ycles the compressor on and of f to maintain the required set point. The third one changes the f an speed if required. There is another alternati ve. The thermostat reduces the indoor f an speed when the space temperature approaches the set point by using a tw o- or three-stage speed operation. If the space temperature drops further, it cycles the compressor of f. If the space temperature drops still further , the indoor fan is f nally cycled off. For a room heat pump system, a two-stage thermostat often selects the heating or cooling mode operation manually. Then it cycles the compressor and e vaporator fan separately or simultaneously to maintain a space temperature set point. Usually , there is a 2 to 5 °F (1.1 to 2.8 °C) difference between heating and cooling set points for room air conditioning systems. If room air conditioners use a tw o-speed (high-low) or three-speed (high-medium-lo w) indoor fan and if the f an speed is re gulated only a few times a day manually , then the supply v olume f ow of this RA C is often steady within an hour , and its air system is constant-v olume in nature. If the thermostat in an RA C reduces the f an speed when the space temperature approaches its set point automatically, the air system is actually a VAV system.

Features Many RACs have the following features: ●

●

●

●

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Most RACs are designed to e xtract outdoor v entilation air and mix it with the recirculating air . Many RACs also can exhaust a portion of the return air through the condenser fan. Many RAC condenser f ans have a slinger -ring which splashs the condensate that has been collected in the pan under the DX coil where it f ows onto the condenser . The water drops hit the outside surface of the air-cooled condenser, evaporate, and increase the air-side heat-transfer coeff cient of the air -cooled condenser. However, in coastal areas, the corrosive effect of w ater with dissolved salts must be considered. The supply v elocity from most RA Cs ranges from 300 to 1000 fpm (1.5 to 5 m / s). RACs have adjustable vanes mounted on the supply outlet to alter the direction of the air jet as well as its throw, to prevent drafts. Some RACs incorporate subcoolers to subcool the hot liquid refrigerant dischar ge from the condenser. Subcoolers are often added after the condenser outlet and submer ged in the condensate before the capillary tube inlet. Subcoolers impro ve the refrigeration ef fect and thus the COP of the RAC system.

System Characteristics System characteristics def ne system capabilities, functions, and the differences between a specif c system and other systems. System characteristics of room air conditioning systems are listed in Table 26.1.

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TABLE 26.1 System Characterisics for RAC, PTAC, and Their Heat Pumps RAC, RHP, SAC Zone thermal and sound control Control zone Control methods Control modes Heating-cooling mode changeover Sound control Indoor air quality (IAQ) Minimum ventilation air control Filters Air systems Types Indoor fan Indoor fan (IF) total pressure Combined IF-motor drive eff ciency Volume f ow control Outdoor fan Cooling systems Refrigeration compressor Refrigerants Evaporator Condenser Refrigerant f ow control Heating systems Type Energy use, cooling

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PTAC, PTHP

Single Electric, two-stage thermostat or DDC, HI-LO, or HI-MED-LO fan speed On-off Manual 45 to 50 dBA

Single Electric, two-stage thermostat or DDC On-off Manual, automatic 45 to 50 dBA

Constant-volume f ow, affected Constant-v by wind direction and speed Coarse or low-eff ciency f lters

olume f ow, affected by wind direction and speed Coarse or low-eff ciency f lters

Constant-volume or VAV air mixing Forward-curved centrifugal 0.6 in. WC 25% HI-LO or HI-MED-LO fan speed Propeller

Constant-volume air mixing Forward-curved centrifugal 0.6 in. WC 25% Single-speed or HI-LO speed Propeller

Rotary HCFC-22, HFC-407C, HFC-410A DX coil Air-cooled Capillary tube, RHP with four-way reversing valve

Rotary and reciprocating HCFC-22, HFC-407C, HFC-410A DX coil Air-cooled Capillary tube, PTHP with four-way reversing valve

Heat pump, or electric heating 8.0 to 10.0 EER

Hot water, electric heating, or heat pump 8.5 to 10.0 EER

26.5 PACKAGED TERMINAL AIR CONDITIONING SYSTEMS Equipment Used in Packaged Terminal Air Conditioning Systems Packaged terminal air conditioners (PT ACs) and packaged terminal heat pumps (PTHPs) are tw o main types of equipment used in packaged terminal air conditioning systems. P ackaged terminal heat pumps add a four -way reversing valve and pro vide winter heating through the release of condensing heat from the outdoor coil and supplementary electric heater. Figure 26.2 shows a packaged terminal air conditioner . The main chassis contains an air conditioner to provide heating and cooling. A cabinet sleeve is used to slide the chassis through the w all. A front panel with a return air grille pro vides a neat appearance. There is also an e xterior louver to show a more attracti ve appearance on the outer b uilding facade. Heating can be pro vided by heat pumps, electric heating, gas heater, or hot w ater heating system, especially an e xisting hot w ater system. According to ASHRAE Handbook 1996, all-electric PTACs are dominating the current market. In the mid-1990s, PTACs with electric resistance had a share of 49 percent of packaged terminal air conditioning systems, and PTHPs a share of 45 percent. PT ACs are a vailable with cooling capacity of 6000 to 18,000 Btu / h (1.8 to 5.3 kW). Heating and Cooling Mode Operation In a typical PTHP , the return air enters from the return grille on the front panel. It then f ows through a coarse f lter and an indoor coil (DX coil) in which it is cooled and dehumidi f ed during

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Exterior louver

Cabinet sleeve Chassis

Supply grille

Filter

DXcoil

Return air grille

FIGURE 26.2 A packaged terminal air conditioner.

cooling mode operation. The conditioned air is then mix ed with the outdoor v entilation air e xtracted by the indoor fan and discharged through the adjustable supply grilles upward. A thermostat senses the zone temperature and controls the refrigerant supplied to the DX coil in on / off mode to maintain a preset zone temperature. During heating mode operation, the indoor coil acts as a condenser. The return air is then heated by the indoor coil to maintain the required zone temperature. As in an RA C, spray the condensate on the outdoor coil (condensing coil) where it e vaporates and enhances the release of condensing heat. In coastal areas, if corrosive saltwater is found in the condensate sump, consider installing a condensate draining system.

Minimum Efficiency Requirements ASHRAE / IESNA Standard 90.1-1999 speci f es that the minimum ef f ciency requirements for ne w installed packaged terminal air conditioners and packaged terminal heat pumps in cooling mode operation can be calculated as: EER  10.0

0.16Q rc 1000

(26.1)

Eff ciency as of October 29, 2001 for packaged terminal air conditioners: SH__ ST__ LG__ DF

EER  12.5

0.213Q rc 1000

(26.2)

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Eff ciency as of October 29, 2001 for packaged terminal heat pumps: EER  12.3

0.213Q rc 1000

For new installed packaged terminal heat pumps in heating mode operation, ciency requirements can be calculated as: COP  2.9

0.026Q rc 1000

COP  3.2

0.026Q rc 1000

__RH

(26.3) their minimum ef f (26.4)

Eff ciency as of October 29, 2001:

where Qrc  cooling capacity , Btu / h. F or f actory-labeled replacement of packaged terminal air conditioners and heat pumps, only their required minimum EER or COP will be slighly lo wer as of October 29, 2001. Refer to ASHRAE/IESNA Standard 90.1-1999 for details.

System Characteristics System characteristics of PT ACs and PTHPs are listed in Table 26.1. Compare the system characteristics of RACs with PTACs, also RHPs with PTHPs; they are quite similar, except for the following differences: ●

●

●

RACs and RHPs are often windo w-mounted, whereas PTACs and PTHPs are mounted through the wall. PTACs and PTHPs ha ve a better appearance on the b uilding f acade outdoors as well as on the front panel indoors. Some PTACs and PTHPs of fer more control functions such as automatic changeo ver from cooling mode operation to heating mode operation, and vice v ersa, evaporator freeze-up protection, limited operation when zone temperature limits are e xceeded, thermostat set points adjustments during the unoccupied period, and fault detection and diagnostics.

Applications According to Air Conditioning, Heating & Refrigeration News on April 15, 1996, PTAC shipments in 1994 in the United States comprised 75,000 units and PTHPs 100,000 units. PT ACs and PTHPs are widely used in commercial b uildings. They are mostly adopted in relati vely small zones in the perimeter zones of buildings such as hotels, motels, apartments, nursing homes, and off ces.

REFERENCES ASHRAE, ASHRAE Handbook 1996, HVAC Systems and Equipment, ASHRAE Inc., Atlanta, GA, 1996. ASHRAE, ASHRAE Handbook 1999, HVAC Applications, Atlanta, GA, 1999. ASHRAE, ASHRAE / IES Standard 90.1-1999 Energy Standard for Buildings Except New Low-Rise Residential Buildings, Atlanta, GA, 1999. DOE / EIA, 1998 Nonresidential Building Energy Consumption Survey: Commercial Buildings Consumption and Expenditure 1995, DOE / EIA-0318 (95). Gladstone, J., and Humphreys, K. K., Mechanical Estimating Guidebook, McGraw-Hill, New York, 1995.

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Harold, R. G., Sound and Vibrations Consideration in Rooftop Installations, ASHRAE Transactions, 1991, Part I, pp. 445 – 453. Lindford, R. G., and Taylor, S. T., HVAC Systems: Central vs. Floor-by-Floor, Heating / Piping / Air Conditioning, July 1989, pp. 43 – 57. Mahoney, T. A., 5 Million Unitary Shipments May Become Industry Yardstick, Air Conditioning, Heating and Refrigeration News, April 15, 1996. McGreal, M. P., Engineered Smoke Control Systems, HPAC, no. 4, 1997, pp. 69 – 72. Rosenquist, G., Window-Type Room Air Conditioner, ASHRAE Journal, no. 1, 1999, pp. 31 – 36. Wang, S. K., Air Conditioning, vol. 4, Hong Kong Polytechnic, Hong Kong, 1986.

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AIR CONDITIONING SYSTEMS: EVAPORATIVE COOLING SYSTEMS AND EVAPORATIVE COOLERS 27.1 EVAPORATIVE COOLING AND EVAPORATIVE COOLING SYSTEMS 27.1 27.2 DIRECT EVAPORATIVE COOLING AND DIRECT EVAPORATIVE COOLERS 27.2 Direct Evaporative Cooling Process 27.2 Saturation Efficiency 27.2 Direct Evaporative Coolers 27.4 Operating Characteristics 27.6 27.3 INDIRECT EVAPORATIVE COOLING AND INDIRECT EVAPORATIVE COOLERS 27.6 Indirect Evaporative Cooling Process 27.6 Indirect Evaporative Coolers 27.7 Heat-Transfer Process 27.7 Effectiveness 27.10 Operating Characteristics 27.11 Part-Load Operation and Control 27.12 27.4 INDIRECT-DIRECT TWO-STAGE EVAPORATIVE COOLING SYSTEMS 27.13 Indirect-Direct Two-Stage Evaporative Cooler 27.13 Operating Characteristics Using Outdoor Air as Cooled and Wet Air 27.15

Operating Characteristics Using Return Air as Wet Air and Outdoor-Return Air Mixture as Cooled Air 27.15 Energy Efficiency Ratio and Energy Use Intensities 27.16 Case Study: A Two-Stage Evaporative Cooling System in Nevada’s College 27.16 System Characteristics 27.18 27.5 ADD-ON EVAPORATIVE COOLERS 27.18 Add-on Indirect-Direct Evaporative Cooler to a DX Packaged System 27.18 Tower Coil and Rotary Wheel Combination 27.20 Tower, Plate-and-Frame Heat Exchanger, and Coil Combination 27.22 Plate-and-Frame Heat Exchanger 27.23 27.6 DESIGN CONSIDERATIONS 27.24 Scope of Applications 27.24 Beware of Dampness, Sump Maintenance, and Water Leakage 27.24 Selection of Summer Outdoor Design Conditions 27.24 REFERENCES 27.26

27.1 EVAPORATIVE COOLING AND EVAPORATIVE COOLING SYSTEMS Evaporative cooling is an air conditioning process that uses the e vaporation of liquid w ater to cool an airstream directly or indirectly so that the final dry- ulb or dry- and wet-bulb temperatures of the airstream being cooled are lower than those before undergoing the evaporative process. An evaporative cooling air conditioning system, or simply evaporative cooling system, is an air conditioning system in which more than 50 percent of the total cooling pro vided annually is evaporatively cooled. It consists of mainly evaporative coolers, fans, pumps, filters heat recovery devices, heat exchangers, a mixing box, dampers, controls, and other components. An evaporative cooler is a piece of equipment in which the e vaporative cooling process proceeds. A central or a packaged air 27.1

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conditioning system may be comprised of an add-on e vaporative cooler which pro vides 50 percent or less of the total cooling annually. There are three types of e vaporative cooling processes: (1) direct e vaporative cooling, (2) indirect evaporative cooling, and (3) indirect-direct evaporative cooling, as shown in Fig. 27.1.

27.2 DIRECT EVAPORATIVE COOLING AND DIRECT EVAPORATIVE COOLERS Direct Evaporative Cooling Process In a direct e vaporative cooling process, the airstream to be cooled comes directly into contact with the wetted medium or w ater spray. The direct e vaporative cooler in which the direct e vaporative cooling process proceeds is sho wn in Fig. 27.1 a. Air enters the direct e vaporative cooler at point 1 and lea ves at point 2. The release of the latent heat of e vaporation from the directly cooled airstream lowers the airstream temperature; the airstream’ s humidity ratio increases because of the added water vapor.

Saturation Efficiency Saturation efficien y is an inde x used to assess the performance of a direct e vaporative cooler. As discussed in Sec. 15.21, saturation efficien y sat is defined in Eq. (15.54) a sat 

Tae  Tal Tae  T* ae

(27.1)

where Tae, Tal  temperature of air entering and leaving direct evaporative cooler, °F (°C) T* ae  thermodynamic wet-bulb temperature of entering air, °F (°C) For a direct evaporative cooler, if Tae and sat are known, Tal can be determined from a psychrometric chart. The value of sat depends on the following factors: 1. Face velocity va of the air fl wing through the direct e vaporative cooler, in fpm [m /(60 s)]. For a specific cooler ( ater dipping or w ater spraying) with fi ed face area Aa, in ft2 (m2), and a given water fl w rate V˙w, in gpm (L / s), a higher va yields the following results: ●

●

Higher volume fl w rate V˙a of the cooled air, in cfm (m3 /s) Greater evaporative cooling effect Qev,c, in Btu / h (W), which can be calculated as follows: Qev,c  60va Aaacpa(Tae  Tal)

(27.2)

where a  air density, lb / ft3 (kg / m3) cpa  specific heat of moist ai , Btu / lb °F (J / kg °C) ●

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Lower saturation efficien y sat, mainly due to a smaller water fl w rate for each ft3 (m3) of cooled air.

For most direct e vaporative coolers used for comfort air conditioning, the f ace v elocity should usually be no greater than 600 fpm (3.0 m / s) in order to pre vent the carryo ver of w ater droplets. Otherwise, a water eliminator should be installed, which significantly increases the ai -side pressure drop.

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FIGURE 27.1 Types of evaporative systems: (a) direct evaporative cooling; (b) indirect evaporative cooling; (c) indirect-direct evaporative cooling.

27.3

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2. Water-air ratio m˙w /m˙a. This is the ratio of the mass f ow rate of spraying w ater to the mass f ow rate of cooled air, both in lb / min (kg / min). A greater m˙ w /m˙a indicates a comparatively greater contact area between air and water and, therefore, a higher sat. 3. Conf guration of the wetted surface. Wetted media that provide a greater contact surface and a longer contact time between the air and water yield higher value of sat. When a direct evaporative cooler is used to supply cooled air to maintain a space temperature of 80°F (26.7°C) during the summer , it is important to realize that a higher sat always means a point nearer to the saturation curv e, point 2. For a predetermined space temperature Tr  80°F (26.7°C), it means a higher space relative humidity r, as shown in Fig. 27.1a. In a direct evaporative cooling process, recirculating water is usually used in order to save water, which is often more economical. The temperature of the recirculating w ater always approaches the wet-bulb temperature of the cooled air. Because the air is sprayed or in contact with dipped w ater, direct evaporative cooling provides a certain degree of air cleaning. Ho wever, if the cooled air contains a great deal of dirt or particulate matter, an additional f lter should be used to prevent clogging of the wetted medium or nozzles. Other parameters should be considered to assess the performance of a direct e vaporative cooler including ●

●

Use of freshw ater or mak eup w ater, usually e xpressed in gal / h per 1000 cfm (L / h per m 3) of cooled air Air-side pressure drop, in. WC (Pa)

Direct Evaporative Coolers A stand-alone self-contained direct e vaporative cooler that can pro vide cooled air for a conditioned space independently consists of mainly the follo wing: a wetted medium, a fan (which is usually a centrifugal fan to provide the required total pressure loss and a lo wer noise level), dampers, a control system, and a sump at the bottom. F or water-spraying systems, a circulating pump and piping connection are needed to distribute water evenly. To drip water on the medium from the top (e xcept for rotary e vaporative coolers), air f lters, dampers, and an outer casing are necessary . Provisions should be made to bleed off the water in order to prevent mineral buildup. Direct e vaporative coolers can be cate gorized according to the characteristics of the wetted medium as follows:

SH__ ST__ LG__ DF

1. Air washers. An air w asher, or water-spraying chamber, is itself a direct e vaporative cooler. The characteristics and performance of air washers are discussed in detail in Sec. 15.21. 2. Evaporative pads . These media are generally made of 2-in.- (50-mm-) thick aspen w ood f bers with necessary chemical treatment and additi ves to increase wettability and to pre vent the growth of microorganisms, as shown in Fig. 27.2a. Evaporative pads are mounted in removable galvanized steel or plastic frames. Because e vaporative pads require comparati vely lower face velocities, in a self-contained direct evaporative cooler integrated with a centrifugal fan, three sides of the fan cabinet are often mounted with evaporative pads to increase the surface area. 3. Rigid media. These are sheets of rigid and corrugated material made from plastic, impregnated cellulose, or fiberglass, as shown in Fig. 27.2b. Air and water typically flow in a crossflow arrangement so that horizontal channels for airflo w and v ertical channels for w ater flow meet between two corrugated sheets. The depth of the rigid medium is typically 12 in. (300 mm) in the direction of airflo w but may v ary from 8 to 16 in. (200 to 400 mm). Rigid media need no supporting frame. They have lower air pressure drops and can easily be cleaned by w ater flushing. 4. Rotary wheel. A wetted medium in the shape of a rotary wheel is made of corrosion-resistant materials such as plastic, impregnated cellulose, f berglass, or copper alloy, as shown in Fig. 27.2 c.

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FIGURE 27.2 Wetted media for direct e vaporative coolers: (a) e vaporative pad; ( b) rigid media; ( c) rotary wheel.

The depth of the wheel along the airstream direction is from 6 to 10 in. (150 to 250 mm). The rotary wheel is often driven by a motor and gearbox and rotates slowly at a speed of 1 to 2 r / min. The bottom of the wheel is submer ged in a w ater tank. Air f ows through v arious channels of the medium in the direction along the depth of the rotary wheel.

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TABLE 27.1 Operating Characteristics of Direct Evaporative Coolers with Various Wetted Media Type of medium Air washer Evaporative pad Rigid media Rotary wheel

Saturation eff ciency sat 0.80 – 0.90 0.80 0.75 – 0.95

Face velocity, fpm

Air-side pressure drop, in. WC

Water-air ratio m˙w / m˙a

400 – 800 100 – 300 200 – 400 100 – 600

0.2 – 0.5 0.1 0.05 – 0.1 0.5

0.1 – 0.4

Water usage, gal / h 1000 cfm 1.3

Remarks Pad thickness of 2 in. Thickness of 8 – 12 in.

Operating Characteristics Table 27.1 lists the operating characteristics of direct e vaporative coolers with v arious types of wetted media. The e vaporative pad is the traditional type of wetted medium widely used in residential and small commercial b uildings. It has a lo w initial cost and is easy to operate and maintain. Rigid media such as impre gnated cellulose need no support structure, do not emit debris, and have a service life as long as that of aspen pads. They withstand a comparatively higher face velocity, provide a lo wer air pressure drop, and have a slightly greater saturation ef f ciency than aspen pads. The rotary wheel has a more complicated structure. Ho wever, it has no w ater-recirculating system. It is easier to connect in series with other refrigeration coolers or desiccant dryers as an add-on direct e vaporative cooler in an AHU or PU for ener gy-efficient and cost-ef fective operation. An air w asher is a lar ge-capacity, bulky, and expensive direct e vaporative cooler. It is usually used for both humidi f cation and evaporative cooling in industrial applications. The saturation eff ciency values for direct evaporative coolers usually range from 0.75 to 0.95. Among these direct e vaporative coolers, only the cooled air dischar ged from the air w asher is often at an o versaturation state. It carries numerous minute liquid w ater droplets. Ov ersaturation causes wetted surf aces and dampness along the e vaporatively cooled air distrib ution passage. Wetted debris are the nutrients for molds and other microor ganisms. To prevent serious IAQ problems, it is recommended that evaporative pad, rigid media, and rotary wheel be selected as the evaporative wetted media in comfort air conditioning. If an air w asher must be used in process air conditioning, refer to Sec. 15.23 for necessary remedies.

27.3 INDIRECT EVAPORATIVE COOLING AND INDIRECT EVAPORATIVE COOLERS Indirect Evaporative Cooling Process

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In an indirect e vaporative cooling process, the primary airstream to be cooled is separated from a wetted surface by a f at plate or a tube w all, as shown in Fig. 27.1 b, and is called cooled air. The cooled air does not directly contact the e vaporating liquid. A secondary airstream f ows o ver the wetted surface so that the liquid w ater will e vaporate and e xtract heat from the primary airstream through the f at plate or tube w all. This wet secondary airstream is kno wn as wet air. The purpose of the secondary airstream is to cool the wetted surf ace, evaporatively approaching the wet-b ulb temperature, and to absorb the evaporated water vapor. In an indirect e vaporative process, the cooled airstream ’s humidity ratio remains constant because the air to be cooled does not contact the e vaporating liquid. This process is represented by horizontal line 1-2 on the psychrometric chart shown in Fig. 27.1b.

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Indirect Evaporative Coolers Figure 27.3 sho ws a typical indirect e vaporative cooler made by an Australian manufacturer. The main components of this cooler are a plate heat e xchanger, a water spray and recirculating system, an outdoor air intake with f lters, a supply fan and exhaust fan connected by the same vertical shaft, and a f berglass or stainless-steel casing to prevent corrosion. The core part of this indirect e vaporative cooler is the plate heat e xchanger. The plates are made from dimpled, thin polyvinyl chloride plastic. These plates are spaced from 0.08 to 0.12 in. (2 to 3 mm) apart and form alternate horizontal and v ertical passages (i.e., the air to be cooled f ows horizontally, and the air that is sprayed f ows v ertically). Because the plates are only about 0.01 in. (0.25 mm) thick, the thermal resistance of each plastic plate is v ery small, although the thermal conductivity of the plastic is low. Hot, dry outdoor air at point o enters the intake and filters and is extracted by the supply fan. It then enters the back of the e xchanger and is forced through the horizontal passages, in which it releases its heat through the plastic plates to the adjacent wetted surf aces of the v ertical passages. The cooled air at point s flows out the front to the conditioned space, as sho wn in Fig. 27.3a. Water sprays o ver the v ertical passages at the top of the plate heat e xchanger, and forms both wetted surf aces and w ater droplets. Ev aporation from these wetted surf aces and droplets absorbs heat released from the air f owing horizontally through the plastic plates. Excess w ater drops to a sump, which recirculates it to spraying nozzles by means of a pump. Mak eup w ater is supplied from the city w ater supply to account for the e vaporation and carryover. Water is periodically bled off to prevent the buildup of solid matter. Return air from the conditioned space at point r is drawn through the vertical passages between the plastic plates. It absorbs the evaporated water vapor, and its humidity ratio increases. The higher the velocity of the wet air, the greater the wet surface heat-transfer coeff cient hwet, the larger the enthalpy difference hs,w between the saturated air f lm at the wetted surf ace and the wet airstream, and the higher the pressure drop of the wet airstream. Wet air is then forced through the exhaust fan and discharged to the outdoor atmospheric at point ex. Other types of indirect evaporative coolers may use absorbent-lined vertical passages to drip water from the top through distrib uting troughs instead of using w ater sprays. Propeller f ans may be used instead of centrifugal f ans for wet air e xhaust. Dampers may be used to e xtract outdoor air from outdoors or return air from the conditioned space as the wet air depends on which of them has a lower entering wet-bulb temperature.

Heat-Transfer Process There are three f uid streams in a plate heat exchanger: cooled air, wet air, and water f lms along the vertical passages. Because the temperature of the saturated air f lm abo ve the wetted surf ace is nearly equal to the wet-b ulb temperature of the wet air f owing over the surf ace, the heat from the airstream to be cooled on the other side of the plastic plate is transferred to the wetted surf ace to evaporate liquid w ater. The heat transfer process in an indirect e vaporative cooler tak es place mainly between the cooled and wet airstreams. On the cooled air side, the amount of water vapor that permeates the plastic plate is very small and can be ignored; therefore, the entering air is sensibly cooled from point o to point s at a constant humidity ratio, i.e., along a horizontal line os on the psychrometric chart sho wn in Fig. 27.3d. According to Do wdy and Karabash (1987), the heat-transfer coef ficient hair on the air side, in Btu / h  ft2  °F (W / m2  °C), is a function hair  f (Reo0.8 Pr0.33) and can therefore be calculated as h air  0.023

ka

 D  Re h

0.3 0.8 D Pr

(27.3)

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FIGURE 27.3 A typical indirect e vaporative cooler: (a) schematic diagram; ( b) airstream f owing through the passages; (c) heat transfer through plastic plates; (d ) process on psychrometric chart.

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27.9

__RH

90 %

80

60 %

ex

40 %

70

50 %

80 %

0.020

0.012

60 rp1

s

50

60

r

o

rp2

sp

50

0.016

0.08

op

70

90

80

0.04 100

(d) FIGURE 27.3 (Continued)

where ka  thermal conductivity of air, Btu / h ft  °F (W / m2 °C) Dh  hydraulic diameter of cooled air passage, ft (m) In Eq. (27.3) the hydraulic diameter Dh is given as Dh 

4Aca Pca

(27.4)

where Aca  area of cooled air passage, ft2 (m2) Pca  perimeter of cooled air passage, ft (m) On the wet air side, water is sprayed onto the wet airstream; ho wever, because of the heat transfer from the cooled air through the plastic plate to the wet air , the saturation process is no longer adiabatic. Return air from the conditioned space (which becomes the wet airstream) is humidi f ed from point r to ex, as shown in Fig. 27.3d. According to Wu and Yellot (1987), the relative humidity of the air e xhausted from the indirect e vaporative cooler is about 95 percent, and the change in the dry-bulb temperature is rather small. Consequently, there is an increase in wet-bulb temperature and air enthalpy due to the increase in latent heat. In the plate heat e xchanger, the cooled air and the wet air are in a cross f ow arrangement. The temperature of the saturated air f lm on the wet air side depends on the wet-b ulb temperature of the local wet airstream, and the wet-bulb temperature of the wet air gradually increases during the humidifying process. The increase in the wet-b ulb temperature of the wet air can be determined from its enthalpy increase hwet, in Btu / lb (J / kg), and can be calculated as h wet  h ex  h r 

V˙cacacpa(To  Ts ) V˙wetwet

(27.5)

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where hex, hr  enthalpy of wet air at points ex and r, Btu / lb (J / kg) To, Ts  temperature of cooled air at points o and s, °F (°C) V˙ca,V˙wet  volume f ow rate of cooled air and wet air, cfm (m3 /s) ca, wet  cooled air and wet air density, lb / ft3 (kg / m3) cpa  specif c heat of moist air, Btu / lb °F (J / kg°C) The average temperature of the saturated f lm on the wet air side Ts,a, in °F (°C), is approximately equal to the a verage w ater temperature in the w ater sump Tw,s, in °F ( °C). According to actual observation, Tw,s is about 3°F (1.7°C) higher than the wet-b ulb temperature of the return air for this indirect evaporative cooler. As with Eq. (10.21) in Sec. 10.8, the surface heat-transfer coeff cient on the wet air side hwet can be calculated as h wet 

 cm  h pa

(27.6)

dry

In Eq. (27.6), hdry indicates the sensible heat-transfer coef f cient from the wetted surf ace when it is dry; it can be calculated as in Eq. (27.3): k wet

 D Re

h dry  0.023

0.8

h,w

Pr 0.4

(27.7)

where kwet  thermal conductivity of wet air, Btu / h  ft °F (W / m °C) Dh,w  hydraulic diameter of wet air passage, ft (m) Lwet  length of wet air passage, ft (m) Effectiveness As the heat f ow in an indirect evaporative cooler is often in a crossf ow pattern, the performance of an indirect evaporative cooler is mainly determined by its ef fectiveness. Usually, the cooled air has a smaller heat capacity rate than the wet air . Therefore, the indirect evaporative cooler effectiveness in is def ned as follows: Tca,e  Tca,l

in 

Tca,e  Twet,e

(27.8)

where Tca,e, Tca,l  temperature of air to be cooled entering and leaving indirect evaporative cooler, °F (°C) Twet,e  wet-bulb temperature of wet air entering indirect evaporative cooler, °F (°C) Peterson and Hunn (1992) and Peterson (1993), based on e xperiments in the performance of indirect e vaporative coolers, recommended that the ef fectiveness of an indirect e vaporative cooler in either outdoor air or mixing air applications be represented by the following relationships: in 

1 1 Cmin / Cmax

(27.9)

Cca Cmin  Cmax Cwet

(27.10)

and from Eqs. (15.36) and (15.37) C SH__ ST__ LG__ DF

Cca  60 V˙cacacpa Cwet  60 V˙wetwetcsat

(27.11)

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where Cca, Cwet  heat capacity rate of cooled air and wet air, Btu / h°F (W / °C) V˙ca, V˙wet  volume f ow rate of cooled air and wet air, cfm [m3 /(60 s)] ca, wet  density of cooled air and wet air, lb / ft3 (kg / m3) cpa  specif c heat of moist air, Btu / lb  °F (J / kg °C) In Peterson (1993), when cooled and wet airstreams are all outdoor air , the theoretically calculated effectiveness in,c and the measured in based on tests are as follows: Test

Cmin / Cmax

in,c

in

1 2 3 4 5 6 7 8

0.295 0.322 0.363 0.411 0.467 0.543 0.680 0.868

0.77 0.76 0.73 0.71 0.68 0.65 0.60 0.54

0.75 0.76 0.73 0.71 0.68 0.66 0.60 0.55

Wang (1996) conducted tests on the surface wettability effect on heat exchanger plates. The base test used 0.063-in.- (1.6-mm-) thick aluminum plates with v arious coatings 0.004 in. (0.1 mm) thick. The cooled air was set at a volume f ow rate of 26.5 cfm (0.75 m 3 /s), and the ratio of wet air to cooled air w as set at 1.09. Also, the inlet temperature of spraying w ater was controlled at 77 °F (25°C). For a typical test (ANC plates), when the water-air ratio by mass increased from 0.067 (10 mL / min) to 0.27 (40 mL / min), the mean value of effectiveness in of six types of coated aluminum plates increased from 0.38 to 0.63 as follows: Water spraying rate, lb / min (mL / min)

Water-air ratio

Average in

0.011 0.022 0.044 0.055

0.38 0.57 0.63 0.63

0.022 (10) 0.044 (20) 0.088 (40) 0.11 (50)

As the water-air ratio exceeds 0.044, any further increase has an insignif cant effect on effectiveness in. In Eq. (27.11), csat indicates the saturation speci f c heat per de gree of wet-b ulb temperature of the wet air at constant pressure, in Btu / lb °F (J / kg °C); as discussed in Sec. 10.8, m  csat 

dh s

 dT 

(27.12)

where hs  enthalpy dif ference along the saturation curv e, Btu / lb (J / kg). In Peterson ’s (1993) tests, csat varied between 0.825 and 0.899 Btu / lb °F (3.46 and 3.77 kJ / kg °C). The total cooling capacity of the indirect e vaporative cooler Qc, in Btu / h (W), can then be calculated as Q c  60V˙cacacpain(Tca,e  Twet,e)

(27.13)

Operating Characteristics Stand-alone, self-contained indirect e vaporative coolers are made in sizes to handle v olume f ow rates of 1060, 2600, and 3200 cfm (0.53, 1.3, and 1.6 m 3 /s). The size of the indirect e vaporative 3 cooler sho wn in Fig. 27.3 is 2600 cfm (1.3 m /s). Based on observ ation, its maximum po wer

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CHAPTER TWENTY-SEVEN

consumption on hot summer days is 1.68 hp (1250 W), and the centrifugal f an’s total pressure is from 1 to 1.3 in. WC (250 to 325 Pa). In indirect e vaporative coolers, either outside air or return air from the conditioned space can be used as the air to be cooled or the wet air. It depends on which can provide better indirect evaporative cooling results and indoor temperature and relati ve humidity . In Phoenix, Arizona, the cooling design wet-b ulb temperature corresponding to 1 percent annual cumulati ve frequency of occurrence is 75 °F (23.9 °C). F or a summer space temperature of 80 °F (26.7 °C) and a relati ve humidity of 50 percent, the wet-bulb temperature of return air is only 66.5 °F (19.2°C). However, when the return air is used as the cooled air (primary airstream), a speci f ed amount of outdoor ventilation air must be mix ed with return air , and also there is al ways a space latent load; therefore, the space relati ve humidty of the space air wr using the return-outdoor mixture is al ways signif cantly greater than that when only the outdoor air is used. Refer to the discussion in the ne xt section for more details. According to Peterson and Hunn (1992), a seasonal ener gy ef f ciency ratio (SEER) of 17.7 Btu / Wh (5.2 COP) is expected from an indirect evaporative cooler in Dallas, Texas. This is 70 percent higher than an air conditioner with refrigeration. The operating characteristics of indirect e vaporative coolers are af fected by the f ow rates and the pressure drops on the cooled air and wet air sides. F or a specif c cooler, the greater the v olume f ow, the greater the heat-transfer coef f cients, the higher the air v elocity f owing through the passages in the plate heat exchanger, and the higher the pressure drop. The air velocity of the cooled air f owing through the passages is usually from 400 to 1000 fpm (2 to 5 m / s). It is important to limit the air v elocity of the wet air in order to pre vent carryover of water droplets. The indirect evaporative cooler effectiveness in usually ranges from 0.60 at an air side pressure drop of about 0.2 in. WC (50 P a) to an ef fectiveness of 0.80 at a pressure drop of about 1 in. WC (250 Pa). Usually, the cooled air -side pressure drop of indirect e vaporative coolers ranges from 0.2 to 1.5 in. WC (50 to 375 P a) depending on the air v elocities in the heat e xchanger and in the distrib uting duct. The wet air -side pressure drop v aries from 0.5 to 1 in. WC (125 to 250 P a). The volume f ow ratio of cooled air to wet air changes from 0.85 to 1.67. This ratio is the k ey parameter that af fects the heat capacity ratio C and, therefore, the indirect cooler ’s effectiveness, as shown in Eq. (27.9). The altitude of the unit ’s location also has a signi f cant effect on its air density and, therefore, its performance.

Part-Load Operation and Control For a constant-airf ow unit, if the conditions of the outdoor air and the sensible heat ratio of the space conditioning line both remain constant, when there is a reduction in the space cooling load, the space conditioning line sr, as shown in Fig. 27.3 d, tends to e xtend a shorter distance from point s. The space temperature Trp1 drops, and space relati ve humidity rp1 is slightly higher at part-load operation, as shown by point rp1 on the psychrometric chart. If the space cooling load remains the same and the condition of outdoor air changes from point o to point op, with lower outdoor dry-bulb and wet-bulb temperatures, then the supply temperature Tsp, space air temperature Trp2, and space relative humidity rp2 will all be lower at part load. For a small, stand-alone, self-contained indirect evaporative cooler, the fans automatically cycle on and off by means of a control system according to the space temperature at part-load operation. In large coolers, a multispeed f an motor can often be modulated at part-load operation when the space temperature drops belo w a predetermined limit. Serial combination of tw o indirect e vaporative coolers has the same ef fect of increasing the w ater-air ratio as well as contact area and is usually not cost-effective. SH__ ST__ LG__ DF

Example 27.1. An indirect evaporative cooler is used in a retail store in Den ver, Colorado. The 1 percent summer design wet-b ulb temperature for Den ver is 63 °F (17.2 °C), and the coincident dry-bulb temperature is 80°F (26.7°C). If the summer indoor design temperature for this retail store

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is 77°F (25°C) and the supply temperature dif ferential is 10 °F (5.6 °C), calculate the effectiveness required for this indirect evaporative cooler. Solution If outdoor air is used as the cooled air, from Eq. (27.8), in 

Tca,e  Tca,l Tca,e  Twet,e



80  (77  10) 80  63

 0.76

A cooling effectiveness of 0.76 is required for this indirect evaporative cooler.

27.4 INDIRECT-DIRECT TWO-STAGE EVAPORATIVE COOLING SYSTEMS Indirect-Direct Two-Stage Evaporative Cooler When the cooled air lea ves an indirect e vaporative cooler during a hot summer , its dry-bulb temperature may be still abo ve 70°F (21.1°C) with a relati ve humidity between 60 and 80 percent. It is benef cial to add a direct e vaporative cooler so that the temperature of the cooled air can be reduced further with an increase in relative humidity. A higher relative humidity of cooled supply air is often acceptable when the supply air absorbs the space sensible load and maintains a desirable space relati ve humidity during hot summer . Compared with a single-stage indirect e vaporative cooler, an indirect-direct tw o-stage evaporative cooler signi f cantly lowers the space temperature it serves. In an indirect-direct two-stage evaporative cooling system (TSECS), a direct evaporative cooler is always connected in series after an indirect e vaporative cooler to form an indirect-direct e vaporative cooler . If an indirect e vaporative cooler is connected in series after a direct e vaporative cooler, such a direct-indirect evaporative cooler is actually a single-stage direct e vaporative cooler with a considerably higher dischar ge cooled air temperature than that of an indirect-direct cooler. An indirect-direct tw o-stage cooler can pro vide cooled air with a dry-b ulb temperature of 67 °F (19.4°C) and a relati ve humidity of about 95 percent when the entering outdoor air is at a dry-b ulb temperature of 93 °F (33.9°C) and a wet-b ulb of 70 °F (21.1°C). The temperature of this cooled air may increase to an indoor temperature of 78° F (25.5°C) with an indoor relative humidity of 66 percent after absorbing the space cooling load during a hot summer . Indirect and direct e vaporative cooling processes connected in series are illustrated in Figs. 27.1 c and 27.4a along with a psychrometric chart. For an indirect-direct tw o-stage cooler at part-load operation, one of the stages, either direct or indirect, can be turned of f based on the indoor and outdoor conditions. In an indirect-direct tw ostage cooler, plate heat exchangers are widely used in the indirect cooler. On the wet air side, either a water spray or a wetted surf ace treated with water-absorbent material is used. In the direct cooler , a rigid medium with impregnated cellulose is often used. Compared with central systems using a chilled w ater cooling coil or a DX packaged system using a DX coil, an indirect-direct two-stage evaporative cooling system with or without a DX coil refrigeration has the following advantages: ●

●

●

Evaporative cooling replaces all the refrigeration power required and maintains an indoor thermal environment within the ASHRAE summer comfort zone when the outdoor air condition is f avorable. More outdoor ventilation air can be extracted so that IAQ will be improved. Better indoor relati ve humidity control is possible in winter heating when the indoor relati ve humidity is very low.

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80

80

%

50 %

60

90

%

%

w, lb/lb

70

%

40

78/66% r

s

0.016

0.012

60

o

ca, l 50

0.020

s

r

93/70

75/45%

20%

0.008

40

40

50

60

70

(a)

80

90

80

%

70

% 90

60

100

%

50

r

ca, l

100

90

%

80

70 ss s

0.004

o 93/60

ca, l

%

w, lb/lb

%

0.016

40 ca, e

60

o

50

30%

0.008

50

14.

13.

5

13.0

0

20%

0.012

60

70

80

90

(b)

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FIGURE 27.4 Operating characteristics of an indirect-direct two-stage evaporative cooling system: (a) using outdoor air as cooled air and wet air; ( b) using 30 percent outdoor air and 70 percent recirculating air mixture as cooled air and return air as wet air.

27.14

0.004 100

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__RH

The disadvantages include more maintenace for e vaporative coolers because of w ater sprays and wetted media.

Operating Characteristics Using Outdoor Air as Cooled and Wet Air According to the analysis in ASHRAE Handbook 1995, HVAC Applications, for an indirect-direct evaporative cooler that uses all outdoor air for its cooled and wet air , if the ef fectiveness of its indirect evporative cooler is taken as 60 percent, then the saturation efficiency of its direct evaporative cooler is 90 percent. Also, its cooled air dischar ged from the indirect-direct e vaporative cooler can be controlled to maintain a space air condition within ASHRAE summer comfort zones after absorbing the space cooling load at a temperature of 78 °F (25.6°C) or 75°F (23.9°C), with varying space relative humidity as the condition of outdoor air and the discharged cooled air change. If the space conditioning line has a sensible heat ratio of 0.90, the relationships between outdoor air state point o, discharge air points, and indoor space conditions point r are shown in Fig. 27.4a. In Fig. 27.4a upper diagram, when the outdoor air enters the indirect-direct tw o-stage evaporative cooler at a 93 °F (33.9°C) dry-bulb temperature and 70 °F (21.1°C) wet-bulb temperature, it is e vaporatively cooled indirectly to point ca,l with a dry-b ulb temperature of 78 °F (25.6°C) and a wet-b ulb temperature of 66 °F (18.9°C), and then e vaporatively cooled directly to 66.5 °F (19.2°C) with a relative humidity of 95 percent. Cooled air absorbs the system heat gain and the space cooling load, and is controlled to maintain a space temperature of 78 °F (25.6 °C) with a relati ve humidity of about 66 percent. This outdoor air condition (93 °F dry / 70°F wet or 33.9°C dry / 21.1 °C wet) is the upper limit of the space air condition within the ASHRAE summer comfort zone that a properly designed indirect-direct tw o-stage evaporative cooling system can provide. In Fig. 27.4 a lower diagram, when the outdoor temperature enters the tw o-stage e vaporative cooler at a dry-bulb value of 93°F (33.°C) and a 60°F (15.6°C) wet-bulb temperature, the cooled air is discharged from the tw o-stage evaporative cooler at 54 °F (12.2°C) dry-bulb and 52 °F (11.1°C) wet-bulb temperatures. The space condition can then be controlled and maintained at 75°F (23.9°C) with a relative humidity of 45 percent. When a two-stage evaporative cooling system uses all outdoor air to pro vide an indoor environment within the ASHRAE comfort zones, the key parameter that af fects the f nal indoor temperature and relative humidity is the entering outdoor wet-bulb temperature To.

Operating Characteristics Using Return Air as Wet Air and Outdoor-Return Air Mixture as Cooled Air Consider an indirect-direct tw o-stage e vaporative cooling system using return air as the wet air Suppose the following conditions hold: ●

●

●

●

.

The outdoor air dry-b ulb temperature is 93 °F (33.9°C), and the outdoor wet-b ulb temperature is 70°F (21.1°C). Thirty percent of outdoor air is mixed with the return air, and the mixture is used as the cooled air. The effectiveness of the indirect e vaporative cooler is 0.60, and the saturation ef f ciency of the direct evaporative cooler is 90 percent. Space temperature is controlled and maintained at 78°F (25.6°C).

For a mixture of 30 percent outdoor air and 70 percent return air, the temperature of the mixture Tm, in °F (°C), can be approximately calculated as Tm  0.3(93) 0.7(78)  27.9 54.6  82.5°F (28.1°C)

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CHAPTER TWENTY-SEVEN

Assume that the wet-b ulb temperature of the space air is 72 °F (22.2 °C). From Eq. (27.8), cooled air temperature leaving the indirect cooler Tca,l, in °F (°C), can be calculated as

the

Tca,l  Tca,e  in(Tca,e  Twet,e)  82.5  0.6(82.5  72)  76.2 F (24.6 C) From Fig. 27.4 b, draw line ca,l-ss from point ca,l that parallels the thermodynamic wet-b ulb temperature line. As the saturation ef f ciency is 0.9, point s, the condition of cooled air lea ving the direct evaporative cooler, can then be determined; Ts is 70.5°F (21.4°C). Draw line sr from point s with a sensible heat ratio of 0.9. Point r, which represents the condition of the space air , intersects line ro at r. The space air temperature is 78 °F (25.6°C), space relative humidity is 75 percent, and the space wet-b ulb temperature is 72 °F (22.2 °C), which is approximately equal to the assumed value. From the above analysis: ●

●

Because of the addition of the space latent load, the space humidity ratio and space relati ve humidity at point r must be higher than the humidity ratio and the relati ve humidity of the outdoor-return air mixture point ca,e in Fig. 27.4 b. The space humidity ratio and relati ve humidity are also higher when the outdoor -return air mixture is used instead of all outdoor air as sho wn in Fig. 27.4a upper diagram. All outdoor air supplied to the conditioned space signi f cantly improves the IAQ more than when the outdoor-return air mixture is used.

Energy Efficiency Ratio and Energy Use Intensities According to ASHRAE Handbook 1999, HVAC Applications, the energy eff ciency ratio (EER) for an indirect-direct, two-stage evaporative cooling system, including the energy use of fan and pumps in indirect-direct tw o-stage coolers, ranges between 12.9 and 38.7. Cities lik e Austin, Texas, have the lowest EER; and Denver, Colorado, and Albuquerque, New Mexico, both have the highest EER. Compared with a con ventional refrigeration system whose EER  10, an indirect-direct tw o-stage evaporative system installed in the Los Angeles area only needs 35 percent of the ener gy use of a conventional refrigeration system. However, a stand-alone e vaporative cooling system needs ducts, f lters, circulating pumps, and supply, recirculating, and exhaust fans. According to DOE / EIA Commercial Buildings Consumption and Expenditures 1995, the energy consumption of evaporative cooling systems in commercial buildings in 1995 in the United States w as 38.0 kBtu / ft2 yr (11.1 kWh / ft2 yr, or 119 kWh / m2 yr), as sho wn in Table 25.1, which is less than that of other air conditioning systems except for heat pump systems.

Case Study: A Two-Stage Evaporative Cooling System in Nevada’s College Scof eld (1998) reported an indirect-direct tw o-stage evaporating cooling system with DX coil refrigeration which was used in a new three-story building at Western Nevada Community College, at an elevation of 5100 ft (1555 m), in Carson City, Nevada. This new classroom and laboratory complex has a f oor area of about 70,000 ft2 (6500 m2).

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HVAC&R System. An indirect-direct tw o-stage evaporative cooling-VAV packaged system (ECVAVPS) with DX coil and air-cooled condenser was compared with a fan-powered VAV central system with w ater-side economizer . When only an indirect-direct tw o-stage e vaporative cooler is adopted without an y refrigeration as the f nal stage of cooling, occupants ha ve complained of higher indoor relati ve humidity. The f nal decision w as made to install an EC-V AVPS with heat recovery by heat-pipe heat exchangers.

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The 55°F (12.8°C) all outdoor cooled air from the two-stage evaporative cooler provides cooling for 95 percent of the cooling hours. Ho wever, this cooled air temperature is again lo wered to 50°F (10.0°C) by the DX coil so that the space relati ve humidity in summer can be maintained at 40 to 45 percent. This also reduces the supply volume flow as the air density is increased. Winter Heating. During winter heating, the heat-pipe heat e xchanger uses e xhaust heat to raise the outdoor air temperature from 0 °F ( 17.8°C) up to 42 °F (5.6 °C). As the supply v olume f ow rate is turned do wn 50 percent in winter , a DDC controller positions the outdoor and return themselves to introduce 76 percent outdoor air and 24 percent recirculating air . A 10 percent dif ference

TABLE 27.2 System Characteristics of Indirect-Direct Two-Stage Evaporative Cooling Systems System structure Zone thermal and sound controls Control zones Control method Control mode Sound control Indoor air quality Minimum ventilation control Filters Air systems Types Supply fan, types Supply fan total pressure Combined supply fan and motor drive eff ciency Supply fan volume f ow control Return / relief fan Return / relief fan total pressure Return fan volume f ow control Supply-relief fan combination Supply-return fan combination Heat recovery Operating modes

TSECS with or without DX refrigeration Single zones or multiple zones VAV or CV DDC, PI control mode, or on-off NC 30 to 45 All outdoor air, or 20% to 30% f xed amount of outdoor ventilation air Medium eff ciency f lters (MER 9 to 12) Two-stage evaporative VAV packaged system (VAV reheat) Centrifugal, unhoused or plug fans 3.0 to 4.0 in. WC 0.5 Inlet vanes, variable-speed drives Axial fan, centrifugal, unhoused or plug fans 0.6 to 1.2 in. WC Automatically balanced Space pressurization control Heat pipe heat exchanger, 0.65 effectiveness Cooling, heating, nighttime setback, purging, warm-up, optimum start and stop

Cooling systems Indirect evaporative cooler Direct evaporative cooler

Effectiveness 0.6 4- to 12-in. rigid media, saturation eff ciency 90%

Refrigeration systems Compressors Energy use Capacity control Evaporator Condensers

Scroll, or reciprocating EER 10 to 12 Multiple units, staging on / off DX coil Air-cooled or evaporatively-cooled

Heating systems Type Boiler Maintenance

Hot water reheat Gas and electric boilers Sumps equipped with side-stream centrifugal slit separators; periodical f ush of sumps

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in supply and return / exhaust volume f ow is used to maintain a positi ve space pressure. When the outdoor air temperature is raised abo ve 10°F (12.2°C), the heat-pipe heat e xchanger can furnish 100 percent outdoor air. During the dry winter season another alternative is to overheat the outdoor air by using a hot water heating coil. At the same time, use the 4-in.- (100-mm-) deep wetted media direct e vaporative cooler to add necessary humidi f cation and still maintain the required 50 °F (10°C) supply air temperature. When the outdoor air temperature is between 30 and 40°F (1.1 and 4.4°C), it is expected to maintain the indoor relative humidity at 30 percent. Cooling Mode Performance. The operating parameters of this indirect-direct tw o-stage evaporative cooling VAV packaged system with DX coil refrigeration are as follows: ●

●

●

●

●

Outdoor air enters the indirect e vaporative cooler at 96 °F (35.6 °C) dry-bulb and 62 °F (16.7 °C) wet-bulb temperatures. It is e vaporatively cooled indirectly with an ef fectiveness of 0.66 to a dry-b ulb temperature of 73.5°F (23.0°C) and a wet-bulb temperature of 54.3°F (12.4°C). It is sensibly cooled in a DX coil by refrigeration to a dry-b ulb temperature of 59.2 °F (15.1°C) and a wet-bulb temperature of 48.5° F (9.2°C). It is e vaporatively cooled in a direct e vaporative cooler to a dry-b ulb temperature of 50 °F (10.0°C) and a wet-b ulb temperature of 48.5 °F (9.2 °C) and is then supplied to the conditioned space with a saturation eff ciency of 90 percent. The cooled supply air absorbs the f an and duct heat, as well as the space cooling load, and f nally becomes the space air at a dry-b ulb temperature of 75 °F (23.9°C) and a wet-b ulb temperature of 60°F (15.6°C) with a space relative humidity of 45 percent.

Energy and Initial Cost Sav ed. For a duty cycle of 7 a.m. to 7 p.m., 6 d / week, there is a 78 percent reduction in annual ton-hour (kWh) consumption when an indirect-direct tw o-stage evaporative cooling VAV packaged system with DX coil refrigeration is used compared with a conventional VAV central system. The installation cost for the tw o-stage evaporative cooling VAV packaged system is $857,183 and for the conventional VAV central system is $971,016.

System Characteristics System characteristics of an indirect-direct listed in Table 27.2.

two-stage e vaporative cooling system (TSECS) are

27.5 ADD-ON EVAPORATIVE COOLERS Add-on Indirect-Direct Evaporative Cooler to a DX Packaged System In locations where outdoor air has a higher wet-bulb, temperature, evaporative cooling alone cannot provide the required cooling to maintain a desirable indoor en vironment. In man y applications, an add-on evaporative cooler to a packaged system, as shown in Fig. 27.5 a either with a component sequence: indirect cooler , DX coil, and direct cooler or with a sequence: indirect cooler , direct cooler, and DX coil, is often more economical than a packaged system alone. Anderson (1986) has compared a packaged system with an add-on indirect-direct cooler and a packaged system using a DX coil only: SH__ ST__ LG__ DF

Outdoor conditions Supply air Space temperature

100°F (37.8°C) dry, 70°F (21.1°C) wet 57.5°F (14.2°C) dry, 56.5°F (13.6°C) wet 78°F (25.6°C)

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In the evaporative cooler and DX coil combination, return air is used for the wet air, and outdoor air is used only for the cooled air . For cooling capacity control of such a combination, components are usually energized in the follo wing sequence: indirect cooler f rst, then direct cooler, and f nally the DX coil.

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AC SYSTEMS: EVAPORATIVE COOLING

FIGURE 27.5 Evaporative cooler and DX coil versus DX coil only: (a) Evaporative cooler and DX coil; (b) DX coil only.

Space relative humidity System pressure drop DX coil only Evaporative cooling and DX coil Wet air Fan eff ciency Indirect cooler effectiveness Direct cooler saturation eff ciency EER refrigeration Outdoor ventilation air Ratio of installation cost of evaporative coolers and DX coil to DX coil only Energy use (air and refrigeration side) Evaporative coolers and DX coil DX coil only

50 percent 2 in. WC (500 Pa) 2.75 in. WC (688 Pa) 0.85 in WC (213 Pa) 0.6 0.7 0.9 9 Btu / Wh (COP 2.64) 15 percent 2.25 1.10 kW / ton (COP 3.20) 1.79 kW / ton (COP 1.96)

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If the electric po wer rate is $0.05 / kWh, the number of payback hours for an indirect-direct evaporative cooler added onto a DX coil packaged system is 7948; if the rate is $0.075 / kWh, the number of payback hours is 5295; and if the rate is $0.1 / kWh, the number of payback hours is 3974 h.

Tower Coil and Rotary Wheel Combination

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When a cooling to wer is connected to a w ater cooling / heating coil as sho wn in Fig. 27.6 a, the tower coil becomes an indirect e vaporative cooler. During the cooling season, the condenser w ater from the tower is forced through the coil to cool the air f owing over it. Such a cooling coil is often used as a precooling coil, because there may be another cooling coil do wnstream. In the heating season, hot water from the condenser may f ow through the coil and heat the air. Using water from the cooling to wer to cool the air by means of a precooling coil (to replace all or part of the refrigeration) is often called a w ater-economizing process. The combination of a cooling tower and the connected w ater cooling coil is called a w ater economizer and is sho wn in Fig. 27.6a. If the tower coil is in series with a direct e vaporative cooler using a rotary wheel, the combination is actually an indirect-direct cooler which can maintain space conditions in summer similar to those achieved with refrigeration for areas where the outdoor wet-b ulb temperature is belo w 65°C (18.3°C), as discussed in Sec. 27.4. Water returns from the precooling coil, typically at a temperature of 78°F (25.6°C), enters the cooling tower, and is evaporatively cooled to about 70°F (21.1°C). Water is then drawn through the precooling coil, where it absorbs the heat from the air f owing over the coil. It is then pumped back to the to wer at a temperature of about 78 °F (25.6°C) to be evaporatively cooled again. Outdoor air at a dry-b ulb temperature of 100 °F (37.8 °C) and a wet-b ulb temperature of 65 °F (18.3°C) is dra wn through the precooling coil by the supply f an and is sensibly cooled to 75 °F (23.9°C). It then f ows through a rotary wheel type of direct cooler and is e vaporatively cooled, typically to 57.5 °F (14.2 °C) dry-bulb and 56 °F (13.6 °C) wet-bulb temperatures. After that, air is supplied to the conditioned space. Recirculating air may be used instead of outdoor air when it is more benef cial. In an ef fective tower coil, the approach of the cooling to wer should be around 5 °F (2.8°C). In order to ha ve such an approach, the cooling to wer must be 60 percent lar ger than a to wer with a 10°F (5.6°C) approach. The row depth and f n spacing of the precooling coil should be selected to sensibly cool the outdoor air from 100 °F (37.8°C) to 75 °F (23.9°C) at an entering w ater temperature of 70°F (21.1°C). Field experience and tests ha ve shown that a lo w face velocity through the wetted media of the rotary wheel results in a saturation ef f ciency above 0.90 and pre vents carryover. In these tests, the air velocity was about 700 fpm (3.5 m / s) with a pressure drop of 0.25 in. WC (63 P a). The rotary wheel revolved at a rate of approximately 1.5 r / min. In a system with an add-on evaporative cooler, the horsepower and number of operating hours of the fan always far exceed those of the w ater pump. To save energy and reduce operating costs for a large unit, a two-speed (or e ven three-speed) f an is often economical. The eff ciency of the cooler that consists of cooling to wer, cooling coil, and rotary wheel depends lar gely on the conditions of the f lls in the tower, the inner surface of the cooling coil, and the wetted media in the rotary wheel. If the wetted surf aces are clogged with dirt and scale, the eff ciency will decrease proportionally to the resulting drop in e vaporation and air f ow. Periodic bleedof f and other necessary w ater treatments are essential for good performance. Another type of spraying coil – rotary wheel combination is sho wn in Fig. 27.6 b. When a spraying coil is connected to a cooling coil, water that has been evaporatively cooled in the former can be pumped to the cooling coil to absorb heat from the ambient air . Such a spraying coil-coil combination is actually an indirect cooler . If this indirect cooler is connected with a rotary wheel or other type of direct cooler, the resulting combination has a system performance similar to that of a tower-

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FIGURE 27.6 (a) Tower coil and rotary wheel combination; (b) spraying coil-coil and rotary wheel combination; (c) evaporative cooling process.

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coil and rotary wheel combination. The heat transfer between tw o airstreams by means of a spraying coil-coil combination is al ways higher than that of a coil-to-coil arrangement, which is commonly located inside the outdoor air and e xhaust airstreams or between mak eup air and e xhaust airstreams, and is often called a runaround system.

Tower, Plate-and-Frame Heat Exchanger, and Coil Combination When the outdoor wet-b ulb temperature drops belo w a certain v alue (about 40 °F, or 4.4 °C), it is possible to use the condenser w ater from the cooling to wer to cool the air and thus replace all or part of the chilled w ater from the refrigeration plant, as shown in Fig. 27.7. To prevent the dirt and solid matter contained in the open-circuit to wer condenser w ater from scaling the inner surf ace of the chilled w ater coil, a plate-and-frame heat e xchanger is often used so that the to wer condenser water does not enter the coil directly. In Fig. 27.7, the condenser water from the tower at 45°F (7.2°C) enters the plate-and-frame heat exchanger. It cools all or part of the chilled w ater returning from the cooling coil to 50 °F (10.0°C). Chilled water from the plate-and-frame heat e xchanger is then mix ed with the w ater from the centrifugal chiller and enters the chilled water cooling coil at 45°F (7.2°C). To cool the chilled w ater at the heat e xchanger when the outdoor wet-b ulb temperature is 40 °F (4.4°C), the approach of the cooling to wer should be no greater than 5 °F (2.8°C) and the temperature difference at the heat e xchanger between the condenser w ater from the to wer and the chilled

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FIGURE 27.7 Tower, plate-and-frame heat exchanger, and coil combination.

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water from the coil also should not e xceed 5°F (2.8°C). Cost analysis is required to determine the optimum size of the cooling tower and heat exchanger and the f ow rates of the condenser water and chilled water.

Plate-and-Frame Heat Exchanger A plate-and-frame heat e xchanger, as shown in Fig. 27.8, is a liquid-to-liquid heat e xchanger. It consists of a number of corrugated metal plates that are usually made of stainless steel, titanium, or aluminum-brass alloy. Corrugated plates and gaskets form alternate passages that separate tw o different f uids. Gaskets are usually made of elastomers. Warm f uid f ows downward on one side of the corrugated plate, and cold fluid f ows upward on the other side in a counter f ow arrangement. The corrugated plates are compressed together by a f xed-end frame and a mo vable-end frame with clamping bolts. Proper selection of the gask et material and proper operating conditions are important to pre vent f uid leakage. The plates and mo vable-end frame are suspended from an upper carrying bar and lo wer guide bar . Fluid connections are located in the f xed-end frame. The gap between two adjacent plates is rather small, usually ranging from 0.1 to 0.2 in. (2.5 to 5 mm). Because the plates are corrugated, a high degree of turbulence is produced, which results in a high heat-transfer coef f cient. The following equation may be used for the calculation of the heat transfer: h  0.2536

 Lk Re g

0.4 0.65 L Pr

(27.14)

where k  thermal conductivity of f uids, Btu / h ft °F (W / m °C) Lg  spacing of gap, ft (m) The water velocity is usually between 60 and 200 fpm (0.3 and 1 m / s). The higher the f uid velocity, the higher the heat-transfer coef f cient and the greater the pressure drop of the liquid f owing through the heat e xchanger. For a typical plate-and-frame heat e xchanger, the overall heat-transfer coeff cient may be 740 Btu / h ft2 °F (4200 W / m2 °C) through the corrugated plate between the

FIGURE 27.8 Plate-and-frame heat exchanger.

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warm and cold liquids at a pressure drop of about 14.7 psi (101 kP a). The maximum working pressure for a high pressure model can be as high as 350 psig (2412 kPa). The plate-and-frame heat e xchanger has a high heat-transfer coef f cient on the surf aces of its corrugated plates, a smaller temperature dif ference between the f uids on the tw o sides of each plate, and a compact size. It is easily dismantled for cleaning and routine maintenance.

27.7 DESIGN CONSIDERATIONS Scope of Applications In locations where the outdoor dry-b ulb temperature To  95°F (35°C) and outdoor wet-b ulb temperature To  62°F (16.7 °C), a stand-alone direct-indirect tw o-stage e vaporative cooling system plus DX coil refrigeration can pro vide a comfortable indoor en vironment for occupants at a space temperature of 75°F (23.9°C) and a relative humidity of 45 to 50 percent, using evaporating cooling most of the time in the cooling season. In locations where the outdoor dry-bulb temperature To  95°F (35°C) and outdoor wet-bulb temperature To  65°F (18.3°C), a stand-alone indirect-direct tw o-stage evaporating cooling system can maintain a comfortable indoor environment for occupants at a space temperature of 75°F (23.9°C) and a relative humidity of 55 to 65 percent. In locations where the outdoor dry-b ulb temperature To  95°F (35°C) and the outdoor wet-b ulb temperature To  70°F (21.1°C), a stand-alone indirect-direct two-stage evaporating system can maintain an indoor en vironment within the summer comfort zones for occupants at a space temperature of 78°F (25.6°C) and a relative humidity of 55 to 66 percent. In locations where the outdoor wet-b ulb temperature To  72°F (22.2 °C), an add-on e vaporative cooler or an add-on e vaporative cooler plus DX coil refrigeration sa ve energy and is often economical. Life-cycle cost analysis is recommended, especially for small and medium-size systems or systems with a low number of operating hours during the cooling season. In addition to the comfort cooling in residential and commercial b uildings, evaporative cooling has been used in process cooling in industrial applications such as spot cooling, cooling of lar ge motors, cooling of gas turbines and generators, textile mills, cooling of w ood and paper products, laundries, animal barns and houses, and product storage cooling. Evaporative cooling is best suited to applications where both summer cooling and winter humidif cation are required.

Beware of Dampness, Sump Maintenance, and Water Leakage In direct e vaporative coolers, cooled air contacts directly with liquid w ater. Select rigid wetted media instead of atomizing humidi f ers in which w ater droplets may cause dampness or wetted surfaces at a certain distance downstream of the atomizing humidif ers. Sumps must be investigated and cleaned re gularly. Water bleedoffs and necessary w ater treatments are essential to maintain a clean and ef f cient evaporative cooling system. The quantity of bleedof f should be no greater than the rate of e vaporation. The water sump must be properly sealed so that w ater leaks are pre vented. These considerations should be taken into account during system design.

Selection of Summer Outdoor Design Conditions

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During the design of an evaporating cooling system, outdoor conditions should be selected to calculate the performance of the e vaporative coolers; dry-b ulb, wet-bulb,or dew point temperature; and whether this corresponds to 0.4, 1, or 2.5 percent annual accumulative frequency of occurrence and the mean coincident dry-bulb, wet-bulb, or dew point temperature.

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Figure 27.9 shows the performance of an indirect-direct two-stage evaporating cooling VAV system if 1 percent dry-bulb and mean coincident wet-bulb (1 percent DB / MWB) temperatures, 1 percent wet-b ulb and mean coincident dry-b ulb (1 percent WB / MDB) temperatures, and 1 percent dew point and mean coincident dry-b ulb (1 percent DP / MDB) temperatures of Den ver, Colorado, are selected. From ASHRAE Handbook 1997, Fundamentals, in Denver the outdoor design condition based on 1 percent wet-bulb temperature and the mean coincident dry-bulb temperature To has a dry-bulb temperature of 80 °F (26.7°C) and 63 °F (17.2°C). In an indirect e vaporative cooler, if its ef fectiveness is 0.60, the temperature of cooled air discharged from the cooler is Tca,l  80  0.6(80  63)  80  10.2  69.8 F(21.0 C) Draw a line from point ca, l that parallels the thermodynamic wet-b ulb temperature line. With the saturation ef f ciency of the direct e vaporating cooler sat  0.9, the cooled air lea ves this direct evaporating cooler with a dry-b ulb temperature of 60.3 °F (15.7 °C) and a relati ve humidity of 95 percent. If the space temperature is maintained at 75 °F (23.9 °C) and if the sensible heat ratio of the space conditioning line is 0.9, the space relative humidity can be determined from the psychrometric chart to be 60 percent. The supply temperature dif ferential used to calculate the supply v olume f ow rate is Ts  75  60.3  14.7°F (8.2°C). Similarly, the performance of the indirect-direct tw o-stage evaporating cooling system using 1 percent dry-b ulb temperature and the mean coincident wet-b ulb temperature, also 1 percent de w point temperature and the mean coincident dry-bulb temperature can be calculated as follows:

90

%

w, lb/lb

50

60

% 80

70

100

%

80

%

0.016

40

60 s s 50

s

%

r ca,l ca,l

To

30 r r

T o

20%

To

ca,l

50

60

70

0.012

80

90

0.08

0.04

100 To, F

FIGURE 27.9 Comparison of 1 percent dry-b ulb, 1 percent wet-bulb, and 1 percent dew point temperatures in Denver, Colorado.

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Description

1 percent DB / MWB

1 percent WB / MDB

1 percent DP / MDB

Supply cooled air temperature Ts, °F Space relative humidity, percent Supply air differential Ts, °F

53.5 44 21.5

60.3 60 14.7

56.5 52 18.7

From these analyses:

●

●

If 1 percent DB / MWB is adopted as the outdoor air design condition, the design supply differential is 21.5 °F (11.9 °C). When outdoor air of 1 percent WB / MDB occurs, the volume f ow r ate must increase 21.5 / 14.7  1.46, or 46 percent, to maintain a space temperature of 75°F (23.9°C). Otherwise,the space temperature Tr may increase to 75 21.5  14.7  81.8°F (27.7°C). From Fig. 27.9, the temperature of supply cooled air is lo wer than others if 1 percent DB / MWB is adopted as the outdoor air design condition. The 1 percent DB / MWB does not include all the outdoor air at 1 percent WB / MDB and 1 percent DP / MDB.

The annual cumulative frequency of occurrence is no longer 1 percent and is greater than 1 percent, or 1.5 percent, or even more. If 1 percent WB / MDB is adopted as the outdoor air design condition, from Fig. 27.9, it includes more outdoor air at 1 percent DB / MWB and 1 percent DP / MDB than 1 percent DB / MWB, and the annual cumulati ve frequency of occurrence will be less than 1 percent DB / MWB. The performance of an e vaporative cooling system is more closely related to the outdoor wetbulb temperature To than to other outdoor parameters. The value of To not only determines the temperature of cooled air lea ving the direct cooler, but also has a decisi ve effect on the lowest possible wetted surface temperature in an indirect cooler . Outdoor air at a high dry-b ulb temperature To and a low wet-bulb temperature To can easily be cooled in a direct cooler . Outdoor air with a high wetbulb temperature often needs refrigeration. For a stand-alone indirect-direct tw o-stage e vaporating cooling system with or without DX refrigeration that serv es a more demanding project, a wet-b ulb temperature corresponding to 0.4 percent annual cumulati ve frequency of occurrence and the mean coincident dry-b ulb temperature (0.4 percent WB / MDB) are recommended for the design outdoor air condition. For a stand-alone tw o-stage evaporative cooling system or an add-on e vaporative cooler, a wetbulb temperature corresponding to 1 percent annual cumulati ve frequenc y of occurrence and the mean coincident dry-bulb temperature (1 percent WB / MDB) are recommended for the design outdoor air condition.

REFERENCES

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Anderson, W. M., Three-Stage Evaporative Air Conditioning versus Conventional Mechanical Refrigeration, ASHRAE Transactions, 1986, Part IB, pp. 358 – 370. ASHRAE, ASHRAE Handbook 1996, HVAC Systems and Equipment, ASHRAE Inc., Atlanta, GA, 1996. ASHRAE, ASHRAE Handbook 1999, HVAC Applications, Atlanta, GA, 1999. Beaudin, D., Evaporative Cooling System for Remote Medical Center, ASHRAE Journal, no. 5, 1996, pp. 35 – 38. Brown, W. K., Fundamental Concepts Integrating Evaporative Techniques in HVAC Systems, ASHRAE Transactions, 1990, Part I, pp. 1227 – 1235. Brown, W. K., Application of Evaporative Cooling to Large HVAC Systems, ASHRAE Transactions, 1996, Part I, pp. 895 – 907. Colvin, T. D., Off ce Tower Reduces Operating Costs with Two-Stage Evaporative Cooling Systems, ASHRAE Journal, no. 3, 1995, pp. 23 – 24.

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DOE / EIA, 1998 Nonresidential Building Energy Consumption Survey: Commercial Buildings Consumption and Expenditures 1995, DOE / EIA-0318 (95). Dombroski, L., and W. I. Nelson, Two-Stage Evaporative Cooling, Heating / Piping Air / Conditioning, May 1984, pp. 87 – 92. Dowdy, J. A., and N. S. Karabash, Experimental Determination of Heat and Mass Transfer Coeff cients in Rigid Impregnated Cellulose Evaporative Media, ASHRAE Transactions, 1987, Part II, pp. 382 – 395. Dowdy, J. A., R. L. Reid, and E. T. Handy, Experimental Determination of Mass-Transfer Coeff cients in Aspen Pads, ASHRAE Transactions, 1986, Part II, pp. 60 – 70. McClellan, C. H., Estimated Temperature Performance for Evaporative Cooling Systems in Five Locations in the United States, ASHRAE Transactions, 1988, Part II, pp. 1071 – 1090. McDonald, G. W., M. H. Turietta, and R. E. Foster, Modeling Evaporative Cooling Systems with DOE-2.1D, ASHRAE Transactions, 1990, Part I, pp. 1236 – 1240. Meyer, J. R., Evaporative Cooling for Energy Conservation, Heating / Piping / Air Conditioning, September 1983, pp. 111 – 118. Mumma, S. A., C. Cheng, and F. Hamilton, A Design Procedure to Optimize the Selection of the Water-Side Free Cooling Components, ASHRAE Transactions, 1990, Part I, pp. 1250 – 1254. Peterson, J. L., An Effective Model for Indirect Evaporative Coolers, ASHRAE Transactions, 1993, Part II, pp. 392 – 399. Peterson, J. L., and B. D. Hunn, Experimental Performance of an Indirect Evaporative Cooler, ASHRAE Transactions, 1992, Part II, pp. 15 – 23. Scof eld, C. M., and N. H. DesChamos, Indirect Evaporative Cooling Using Plate-Type Heat Exchangers, ASHRAE Transactions, 1984, Part IB, pp. 148 – 153. Scof eld, M., Savings out of Thin Air, Eng. Systems, no. 9, 1998, pp. 98 – 125. Sun, T. Y., Design Experience with Indirect Evaporative Cooling, Heating / Piping / Air Conditioning, January 1988, pp. 149 – 155. Supple, R. G., and D. R. Broughton, Indirect Evaporative Cooling — Mechanical Cooling Design, ASHRAE Transactions, 1985, Part IB, pp. 319 – 328. Wang, T. A., and R. L. Reid, Surface Wettability Effect on an Indirect Evaporative Cooling System, ASHRAE Transactions, 1996, Part I, pp. 427 – 433. Watt, J. R., Nationwide Evaporative Cooling Is Here! ASHRAE Transactions, 1987, Part I, pp. 1237 – 1251. Wu, H., Performance Monitoring of a Two-Stage Evaporative Cooler, ASHRAE Transactions, 1989, Part I, pp. 718 – 725. Wu, H., and J. I. Yellott, Investigation of a Plate-Type Indirect Evaporative Cooling System for Residences in Hot and Arid Climates, ASHRAE Transactions, 1987, Part I, pp. 1252 – 1260. Yellott, J. I., and J. Gamero, Indirect Evaporative Air Coolers for Hot, Dry Climates, ASHRAE Transactions, 1984, Part IB, pp. 139 – 147.

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CHAPTER 28

AIR CONDITIONING SYSTEMS: SPACE CONDITIONING SYSTEMS 28.1 SPACE AIR CONDITIONING SYSTEMS 28.1 Applications 28.1 Advantages and Disadvantages 28.2 Induction Systems 28.3 28.2 FAN-COIL SYSTEMS AND FAN-COIL UNITS 28.3 System Description 28.3 Operating Characteristics 28.3 Fan-Coil Units 28.5 Volume Flow Rate 28.7 Cooling and Dehumidifying 28.8 Heating Capacity 28.9 Sound Power Level of Fan-Coil Units 28.9 28.3 FOUR-PIPE FAN-COIL SYSTEMS 28.9 General Description 28.9 Dedicated Ventilation System 28.10 Space Recirculation Systems 28.11 Temperature of Chilled Water Supplied to Coils 28.11 Exhaust Air to Balance Outdoor Ventilation Air Intake 28.12 Part-Load Operation 28.13 Zone Temperature Control and Sequence of Operations 28.13 System Characteristics 28.14 Calculation of Operational Parameters 28.14

Applications 28.20 28.4 TWO-PIPE FAN-COIL SYSTEMS 28.20 Two-Pipe Systems 28.20 Nonchangeover Two-Pipe Systems 28.20 Changeover Two-Pipe Systems 28.23 System Characteristics 28.24 Applications 28.24 28.5 WATER-SOURCE HEAT PUMP SYSTEMS 28.24 System Description 28.24 Operating Characteristics 28.24 Loop Temperatures 28.25 Water-Source Heat Pumps 28.26 Energy Performance and Energy Use Intensity of WSHPs 28.27 Closed-Circuit Evaporative Water Cooler 28.27 Water Heater 28.29 Storage Tanks 28.29 Air Systems and Maintenance 28.29 Controls 28.30 System Characteristics 28.31 Case Studies: Water-Source Heat Pump Systems 28.31 Design Considerations 28.32 28.6 PANEL HEATING AND COOLING 28.33 REFERENCES 28.33

28.1 SPACE AIR CONDITIONING SYSTEMS Applications As discussed in Sec. 1.5, space air conditioning systems, or simply space conditioning systems are one of the four major air conditioning systems (indi vidual, packaged, space, and central) that comprised about 8 percent of total floor area of commercial b uildings in 1992 in the United States. According to the surv ey of the Census Bureau, the shipments of f an-coil units comprised 215,000 units, and the shipment of w ater-source heat pumps comprised 99,000 units in 1994 in the United States. Space conditioning systems can be subdivided into the following air conditioning systems:

28.1

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Four-pipe fan-coil air conditioning systems, or simply four-pipe fan-coil systems Two-pipe fan-coil air conditioning systems, or simply two-pipe fan-coil systems Water-source heat pump systems

Advantages and Disadvantages Space conditioning systems have the following advantages compared to central and packaged systems: ●

●

●

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●

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Both heating and cooling de vices in the form of terminals, such as fan coils or w ater-source heat pumps, are installed directly above or within the conditioned space, or very near to it. There is no return duct in a space conditioning system e xcept the large core water-source heat pumps. Only a delivery outlet or a short recirculating duct supplies conditioned air to the perimeter zone. As discussed in Sec. 20.9, a separately dedicated outdoor ventilation system is always required to provide v entilation air for occupants as well as the necessary e xhaust air. At the same time, a space recirculating system is used to condition the recirculating air for f an-coil and w ater-source heat pump systems. The required amount of ventilation is always guaranteed, and a dedicated outdoor air ventilation system facilitates the adoption of demand-controlled ventilation. The air economizer cycle for free cooling is only limited to the volume fl w of the outdoor ventilation air. In space conditioning systems e xcept the large core water-source heat pumps, the outdoor ventilation air duct is probably the only main duct with comparati vely less headroom required than in the supply duct that crosses under the beams in the ceiling plenum. Both four -pipe f an coils and console w ater-source heat pumps pro vide indi vidual zone control with each f an coil or each console w ater-source heat pump serving an indi vidual zone. There is often an additional high-lo w speed or high-medium-lo w speed control on the v olume fl w of the supply fan. The four-pipe fan coil or the w ater-source heat pump can be changed from cooling mode to heating mode and vice v ersa automatically or manually , and it adjusts its capacity to meet the variation of the zone load during part-load operation. There is less cross-contamination between rooms and control zones. According to Zaidi and Howell (1993), the energy use of a four-pipe fan-coil system was only 73 percent compared with a VAV reheat system, and the energy use of a water-source heat pump was only 63 percent that of a VAV system. Anantapantula and Sauer (1994) found that the ener gy use of a four -pipe fan-coil system compared to a VAV reheat system w as even lower. Lower energy use in fan-coil and water-source heat pump systems is due to the following: First, there is far less energy required to transport the conditioned air and the recirculating air; also, no energy is needed to transport the return air. Second, a fan-coil system can transfer e xcess heat from the interior zone of a b uilding to the perimeter zone through a heat reco very system, as discussed in Sec. 13.3. A w ater-source heat pump operated in heating mode can absorb the heat rejected by another w ater-source heat pump that is operated in cooling mode and recovers it for heating in the perimeter zone. Space conditioning systems have the following disadvantages: Only low-efficien y filters are used in an coils and w ater-source heat pumps. Lo w efficien y air filters are unable to rem ve particulates of size  3 m and are inef fective at proving an acceptable IAQ for occupants as well as protecting the coils and components in the air distrib ution system from dust and contaminants. Fan-coil systems reduce the chilled w ater fl w rate to decrease their cooling capacity during part load. Therefore, this results in a significantly higher zone relat ve humidity at part-load operation. The zone relative humidity may increase to 70 percent at part load, especially when the fan coil is oversized. Because fan coils and w ater source heat pumps are often mounted directly abo ve or near to the conditioned space, they may create a noise problem, as discussed in Sec. 26.2, which is al ways

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28.3

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annoying. A receiver may receive a sound pressure level between 35 and 40 dBA in a conditioned space produced by a typical fan coil of 400 cfm (189 L / s). Fan coils and w ater-source heat pumps are scattered terminals. Both ha ve moving parts, filters and condensate pans and require site maintenance directly abo ve, near, or within the conditioned space, which causes dif ficulties inconvenience, and possible leaks of w ater and condensate. In addition, because only low-efficien y or coarse filters are used in the an coils and water-source heat pumps, more cleaning work is required for the coils, ducts, and air distribution components. The volume fl w rate of outdoor v entilation air supplied to the conditioned space is not suf ficien to pressurize the floor immediately ab ve or below the fire flo , to prevent smoke contamination in case of a building fire

Induction Systems In an induction system, recirculating air e xtracted through the coil is due to a ne gative pressure formed adjacent to the high-v elocity supply nozzle. An induction system sees a considerably greater energy use than fan-coil systems and is not recommended in new and retrofit projects

28.2 FAN-COIL SYSTEMS AND FAN-COIL UNITS System Description A four-pipe or a tw o-pipe fan-coil system includes boilers and chillers in the central plant, water system supplying chilled or hot w ater to the f an coils, a space recirculating system using f an coils to condition the space recirculating air, and a dedicated outdoor ventilation air system using an outdoor air AHU to condition the outdoor air , as shown in Fig. 28.1. Both the space recirculating and dedicated outdoor air systems have ducts, diffusers, inlets, controls, and accessories. Outdoor air is often cooled and dehumidified heated, or sometimes e ven humidified in a sepa rate outdoor air or makeup AHU. The outdoor air is then transported to the f an coils where outdoor air is mixed with the recirculated air and conditioned. If the conditioned outdoor air is supplied directly to the conditioned space via supply ducts in the dedicated v entilation system, then it is mixed with the conditioned recirculating air from the f an coil of the space recirculating system in the space. Outdoor (primary) air mixing with recirculating air in the outdoor air AHU is not economical because recirculating air must be transported back to the outdoor air AHU and conditioned there instead of in the fan coil in or near the conditioned space. Normally, there are no return fans or return ducts in four-pipe or two-pipe fan-coil systems. Chilled water or hot w ater is supplied to the f an coil through a tw o-pipe or four-pipe water system. In locations with a moderate winter , or where electric rates are f avorable in winter, an electric heating coil in each f an-coil unit is sometimes used for winter heating. This greatly simplifies th operation of fan-coil systems. Four-pipe and two-pipe fan-coil systems and w ater-source heat pump systems are the air conditioning systems that use dedicated ventilation and space recirculating air systems.

Operating Characteristics During summer cooling mode operation, outdoor air at point o in Fig. 28.1 is cooled and dehumidified and filtered in the outdoor air AHU. The outdoor air lea ves the cooling coil at point pc, f ows through the supply f an outlet at point pf in the dedicated v entilation system, and is transported by the supply duct to the fan coil in the space recirculating systems or directly to the conditioned space

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FIGURE 28.1 A four -pipe f an-coil system with outdoor air supplied to the mixing plenum of f (a) schematic diagram; (b) air conditioning cycle.

DF

28.4

an-coil unit:

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at point ps. At ps, the dedicated ventilation system is combined with the space recirculating system and forms the air system of the four-pipe or two-pipe fan-coil system. The outdoor air can be mixed with the recirculating air in three ways: ●

●

●

Outdoor air at point ps is supplied to the mixing plenum of the f an coil. It is then mix ed with the recirculating air, which is extracted by the fan in the fan-coil unit through the filter and forms th mixture m. Air pressure at the mixing plenum is slightly less than the space pressure, so that space air can be e xtracted into the plenum. The mixture then fl ws through the f an and coil in the f ancoil unit, cools and dehumidifies at the coil and leaves the coil at point fc. Air that has been conditioned in the fan coil is supplied to the conditioned space to of fset the space load. State point fc may or may not be on line pc-ps, but will be near point pf. Outdoor air at point ps is supplied to the conditioned space directly . In this case, space air is filtered and cooled and dehumidified (or heated) in the f an coil. The conditioned space air discharged from the fan-coil unit at point fc is mixed with the outdoor air in the space. It is assumed that the mixture at point m offsets the space load. In such an arrangement, the recirculating airstream will recirculate along its own circuit. Outdoor air at point ps is supplied just before the supply outlet of the f an-coil unit. It mix es with the conditioned air from the f an coil at point fc. The psychrometric c ycle of this arrangement is approximately the same as that of outdoor air supplied directly to the conditioned space.

Outdoor air supplied to the mixing plenum or before the supply outlet of the f an-coil unit provides a balanced distribution of outdoor air in the conditioned space. On the other hand, outdoor air supplied directly to the conditioned space has a shorter outdoor air supply duct, and the conditioned air is better distrib uted when one of the f an-coil units is turned of f during capacity control at partload operation, as multiple fan-coil units are installed in a large room. The temperature increase from fan power heat gain in the fan-coil unit is about 0.5°F (0.3°C) for a permanent-slit capacitor fan motor and about 0.8°F (0.5°C) for a shaded-pole f an motor. The supply duct in the ceiling plenum after the f an-coil unit is usually v ery short, so such a duct heat gain or loss is negligible. During winter heating mode operation, in extremely cold weather , outdoor air may be preheated to point ph by the heating coil in the outdoor air AHU. After absorbing the supply f an power to point pf and releasing the outdoor air supply duct heat loss to point ps in the dedicated ventilation system, the outdoor air is mix ed with recirculating air e xtracted by the f an coil at point r, and forms a mixture m, as shown by line ps-m-r in Fig. 28.1 b. The mixture is heated in the f an coil to point cf by the f an power heat gain and then to point s by the heat released from the coil in the space recirculating system, and the mixture is supplied to the conditioned space. Outdoor air can also be mix ed with heated air from the f an-coil unit when outdoor air is supplied directly to the conditioned space. In locations with a mild winter , only outdoor air is heated in the outdoor air AHU. Fan-Coil Units A fan-coil unit, or a fan coil, is a terminal unit installed directly inside the conditioned space or in the ceiling plenum just abo ve the conditioned space. A fan-coil unit includes a small motor dri ven centrifugal fan or tw o small centrifugal f ans connected in parallel, a finned coil a filte , an outer casing, and controls. Sometimes, a cooling coil and a heating coil may be connected in series along the airfl w, as shown in Fig. 28.2. A fan-coil unit can be a horizontal unit installed inside the ceiling plenum (Fig. 28.2 a and b) or a vertical unit mounted on the floor under the wind wsill (Fig. 28.2c) or a stack unit installed v ertically along the tw o sides of the windo w. Vertical and stack units are usually used to of fset the cold draft on the inner surf ace of windo w glass or on the e xternal wall during cold weather . Cold draft often fl ws do wnward along the glass because it is hea vier than the surrounding air . Warm air

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FIGURE 28.2 Fan-coil units: (a) horizontal unit; (b) horizontal unit with inner lined plenum; (c) vertical unit.

discharged from a f an coil during winter heating raises the inner surf ace temperature of windo w glass. Horizontal and v ertical fan-coil units are sho wn in Fig. 28.2. F an-coil units are a vailable in standard sizes 02, 03, 04, 06, 08, 10, 12, 16, and 20. Size 02 means a nominal fl w rate of 200 cfm (0.1 m3 /s), 04 means 400 cfm (0.2 m3 /s), and so on.

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Fan. Double-width, double-inlet, forward-curved centrifugal f ans are usually used because of their compact size and lo wer noise le vel. The fan wheels are usually made of aluminum or galv anized steel with a diameter less than 10 in. (250 mm) in most cases. F an housings are die-formed with integral scrolls and inlets.

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Fan Motor. Permanent-split capacitor (PSC) motors and shaded-pole (SP) motors are used. As in Sec. 26.4, PSC motors have an efficien y of 55 to 70 percent which is considerably higher than that of SP motors, which have 30 to 40 percent efficien y. Two-speed high-low switches or three-speed high-medium-low switches are used to vary the fan speed manually or automatically (by a DDC controller). F an motors are generally protected by a thermal overload protector. Periodic oiling of the bearing (twice per year) is required. Coils. Coils are usually made from copper tubes and aluminum fins. Coolin / heating coils usually have two, three, or four rows of fins depending on the coil’s cooling capacity and the sensible heat ratio of the cooling and dehumidifying process SHR c. Two-row coils or three-ro w coils are widely used. Four-row coils have a greater dehumidifying capacity. Usually, there is only one coil for both heating and cooling. A separate electric heating coil is sometimes used with two-stage step control in locations where the heating season is short or ener gy rates are low in winter. Manual air vents are installed to pre vent the formation of air pock ets inside the water circuit. A galvanized-steel pan with an insulating liner is often used to drain the condensate during dehumidification and to pr vent outer surface condensation on the fan-coil unit. To reduce the air -side pressure loss, the face velocity of the air fl wing through the coil is usually from 200 fpm (1 m / s) to 300 fpm (1.5 m/ s). Filters. Usually lo w-efficien y, low-pressure-drop permanent filters are used. They are easy to clean and replace periodically. Sometimes disposable, low-efficien y fibe glass filters are used Casing. The external cabinet is usually made of 18-gauge (1.3-mm) galv anized-steel sheet with a corrosion-resistant surf ace coating. The cabinet is lined with insulation to pre vent outer surf ace condensation. Although there are f ans, water coils, and filters in both an-coil units and air -handling units, a fan-coil unit is distinguished from an AHU by the following characteristics: Fan coil Classification of equipment Location Volume fl w, cfm (m3 / s) Fan total pressure, in. WC (Pa) Sound power level, dB Filter efficien y Coil row depth Fin spacing Selection of unit

A terminal unit

AHU Basic equipment in air system Fan room

Under window sill or in ceiling plenum  2000 (1)  0.6 (150)

1200 – 50,000 (0.6 – 25)  6 (1500)

Lower Low 2, 3, or 4 rows Fixed Based on cooling capacity

Higher Medium or high 2, 3, 4, 6, or 8 rows Custom-made Based on volume fl w rate

Volume Flow Rate The volume fl w rate of a fan-coil unit is affected by the following factors: ●

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The external pressure drop because of the filters the supply duct, and the diffusers in the space recirculating system The position of the fan switch, whether is high-low or high-medium-low The elevation of the fan coil above sea level

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The nominal fl w rate is the fl w rate of a f an coil whose e xternal pressure drop is at a specific alue when the fan switch is in the high position and the f an coil is at sea le vel. The higher the external pressure drop, the lower the v olume fl w rate. If the e xternal pressure drop increases from 0.06 to 0.3 in. WC (15 to 75 Pa), the volume fl w rate of the fan coil may decrease to 55 percent of its nominal v alue. For a typical fan coil, the volume fl w rate of fan switch position medium is about 80 percent of the high value, and low is only 70 percent of the high value.

Cooling and Dehumidifying Because the face velocity of a specific an-coil unit at nominal v olume fl w rate is nearly the same for various sizes, and because the outer surf ace area of a w ater cooling coil is di vided into dry and wet parts during cooling and dehumidifying, the coil capacity and the sensible heat ratio of the cooling and dehumidifying process SHRc depend on the following factors: ●

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Dry- and wet-bulb temperatures of entering air Entering water temperature Water temperature rise in the coil The surface area and number of rows, including both pipe surface and fin area of the coil

For a 04 f an-coil unit with an entering air temperature of 80°F (26.7°C) dry-b ulb and 67°F (19.4°C) wet-bulb, and an entering water temperature of 45°F (7.2°C), if the water temperature rise is 10°F (5.6°C), then its cooling and dehumidifying capacity Qcc varies from 11 to 14 MBtu / h (3.22 to 4.1 kW), and SHRc varies from 0.65 to 0.80. If the volume fl w rate of a f an-coil unit deviates from the nominal v alue because of a greater external pressure drop or a higher altitude, its total cooling capacity should be corrected as follows: Qc,c  CpCaQc,r

(28.1)

where Qc,c  corrected total cooling capacity, MBtu / h (kW) Qc,r  manufacturer’s catalog listed cooling capacity at a specific nominal olume fl w rate, MBtu / h (kW) Cp, Ca  total cooling capacity correction factor for excessive external pressure drop and high altitude, respectively The corrected sensible cooling capacity of the fan coil Qcs,c, in MBtu / h (kW), is Qcs,c  CpsCasqcs,r

(28.2)

where Qcs,r  manufacturer’s catalog-listed sensible cooling capacity at nominal volume fl w rate, MBtu / h (kW) Cps, Cas  sensible cooling capacity correction factor for excessive external pressure and high altitude, respectively Correction factors Cp, Cps, Ca, and Cas can be found in the manuf acturer’s catalog. If these data are not available, the following values can then be used for Cp and Cps: External pressure drop, in. WC (Pa)

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0.06 (15) 0.1 (25) 0.2 (50) 0.25 (63)

Cp

Cps

1 0.96 0.84 0.76

1 0.96 0.83 0.73

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If the sensible heat factor of the cooling and dehumidification process 0.  SHRc  0.95, for each 1000 ft (328 m) higher than sea level up to an altitude of 10,000 ft (3280 m); Ca  1  Cas  1 

0.01(altitude, ft) 1000 0.03(altitude, ft)

(28.3)

1000

Heating Capacity The heating capacity of the selected f an coil can be found in the manuf acturer’s catalog. Usually , for the same f an coil, a greater heating capacity can be pro vided at winter design conditions because the temperature dif ference between the hot w ater and heated air is higher than that between chilled water and cooled air.

Sound Power Level of Fan-Coil Units Because the fan-coil unit is usually located inside the ceiling plenum or directly under the windo w sill, the room ef fect and the short supply duct are often not suf ficient to attenuate an noise in the fan-coil unit. Therefore, sound power level is often an important factor to consider during the selection of a fan-coil unit. For a typical fan-coil unit of size between 02 and 12, the sound power level Lp rating, measured in a re verberant room according to ARI Standard 443-70 for an octa ve band with a middle frequency of 1000 Hz, varies from 45.5 to 52 dB. F or a fan-coil unit of size 04, because Lp  47.5 dB for an octave band with a middle frequenc y of 1000 Hz, if the room effect is 7.5 dB, the space NC level is about 40. The greater the size of the fan-coil unit, the higher the NC level. Refer to the manufacturer’s catalog for details.

28.3 FOUR-PIPE FAN-COIL SYSTEMS General Description The water systems used for the fan coils can be classified into t o categories: two-pipe systems and four-pipe systems. The three-pipe system w as discontinued because of the ener gy loss in the common return pipe. A four-pipe fan-coil system is equipped with tw o supply mains: a chilled w ater supply and a hot water supply. There are also two return mains: a chilled water return and a hot water return, as shown in Fig. 28.3. The finned coil may be a common coil as shown in Fig. 28.3a, or two separate coils, a cooling coil of two or three rows, and a heating coil of one row, as shown in Fig. 28.3b. In a common coil, the DDC controller admits the chilled w ater and hot w ater in sequence to tw o three-way valves and modulates their w ater fl w at part-load operation to maintain a preset space temperature. The chilled water stream never mixes with the hot water in these three-way valves. When two separate coils are used, a DDC controller admits chilled water to the cooling coil and hot water to the heating coil in sequence and modulates their w ater fl w at part load. Hot w ater, steam, or electric energy can be used as the heat source. A drainage pipe for condensate in the f an-coil unit is needed as the air is often cooled and dehumidified in the an coil, especially where conditioned space may have hot and humid infiltrated out door air. A four -pipe system is more fl xible, easy to operate because there are no troublesome

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FIGURE 28.3 Four-pipe fan-coil unit system: (a) common coil; (b) separate coil.

changeover problems, and lower in operating cost than a two-pipe system. On the other hand, it has a higher initial cost; and if reverse return pipe is used, many pipes must be squeezed into the ceiling plenum. Four-pipe fan-coil systems are most widely used.

Dedicated Ventilation System

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The dedicated v entilaton system in a four -pipe fan-coil system must pro vide minimum outdoor air V˙p, in cfm (L / s), specif ed for occupants according to ASHRAE Standard 62-1999, served by each fan coil. The estimated cooling capacity Qcpn, in Btu / h (W), provided by the minimum outdoor air to offset the space cooling load can be calculated as Q cpn  60V˙pps(h r  h ps)

(28.4)

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where ps  air density of outdoor air, lb / ft3 (kg / m3) hr, hps  enthalpy of space air and outdoor air, Btu / lb (J / kg) The estimated sensible cooling capacity of the minimum outdoor air Qs,pn, in Btu / h (W), that offsets the space sensible cooling load can be calculated as Q spn  60V˙ppscpa(Tr  Ts)

(28.5)

where Tr, Ts  temperature of space air and outdoor air, °F (°C). The total cooling capacity Qcp and the total sensible cooling capacity Qsp of the cooling coil in the outdoor air AHU, both in Btu / h, are the sum of the cooling capacities required to condition the outdoor air supplied to each of the f an coils served by this outdoor air AHU, and they can be calculated as n

Q cp  Q cpn 1

n

Q sp  Q spn 1

(28.6)

where n  number of fan coils served by the outdoor air AHU. Space Recirculation Systems Each fan-coil unit forms a space recirculation system. Its system total pressure v aries between 0.06 and 0.6 in. WC (15 and 150 Pa). Estimated Cooling Capacity . The size of a f an-coil unit is selected to meet the required cooling capacity and sensible cooling capacity to offset the space cooling load Qrc and sensible cooling load Qrs, both in Btu / h (W). As the conditioned outdoor air supplied to a speci f c fan coil or to the conditioned space serv ed by the f an coil of fsets part of the space load, the estimated total cooling capacity Qc,fc and sensible cooling capacity Qs,fc of a fan coil, both in Btu / h (W), for a specif c control zone or conditioned area it serves can be calculated as Q c,fc  Q rc  Q c,p  60V˙pps(h r  h ps) Q s,fc  Q rs  Q sp  60V˙ppscpa(Tr  Tps)

(28.7)

where hr, hps  enthapy of zone air and conditioned outdoor supply air, Btu / lb (J / kg) Tr, Tps  temperature of zone air and conditioned outdoor supply air, °F (°C) V˙p  volume f ow rate of outdoor (primary) air supplied to each fan coil, cfm [m3 /(60 s)] Sensible Heat Ratios. When a four-pipe fan-coil system is operated under steady equilibrium, the sensible heat ratio of the space conditioning line SHR s is usually approximately equal to the sensible heat ratio of the cooling and dehumidifying process of the f an coil SHR c,f because the cooling and dehumidifying process in the fan coil cf-fc and the space conditioning line fc-r form a cycle. During summer design conditions, the temperature of chilled water supplied to the fan coils and to the outdoor air AHU is usually the same, Tw,p  Tw,f, both in °F (°C). Also, the humidity ratios of the conditioned outdoor air at point ps and the air lea ving the f an-coil point fc are approximately equal to each other , wps  wfc, both in lb / lb (kg / kg). For fan-coil units serving the interior zone, more often sensible ratios SHR s  SHRc,f  0.8. F or f an-coil units serving a perimeter zone, SHRs  SHRc,f  0.95, as shown in Fig. 28.1b. Temperature of Chilled Water Supplied to Coils In a four -pipe fan-coil system, the space or zone temperature is a speci f ed design criterion within the ASHRAE comfort zones. In summer , it is usually between 75 and 78 °F (23.9 and 25.6 °C). The humidity ratios of the supply air from the outdoor air AHU wps and the supply air from the f an coil wfc, as well as the humidity ratio of the space air wr depend on the temperature of the chilled w ater

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39445 Wang (MCGHP) Ch_28

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28.12

SECOND PASS bzm 7/24/00

pg 28.12

CHAPTER TWENTY-EIGHT

supplied to the f an coils Tw,f and to the outdoor air AHU Tw,p. Also Tw,f and Tw,p have a direct in f uence on the cooling capacity of the fan coils and the outdoor air AHU. If Tw,p  Tw,f, then wps  wfc, and wps  wr. A considerable amount of both sensible and latent load is undertaken by the outdoor air during cooling mode operation. The required cooling capacities of the fan-coil units are considerably less than those of the space cooling load. If Tw,p  Tw,f, wps  wfc, and wps  wr, there is a certain amount of sensible load and a comparatively less amount of latent load undertak en by the outdoor air supplied from the outdoor air AHU during cooling mode operation. The required sensible cooling capacity of the f an coils is smaller than the space sensible cooling load. In the early days, the chilled w ater temperature supplied to the f an coils Tw,f was deliberately equal to or higher than the de w point of the entering air to the f an coil Tf,en, so that there w as no condensate in the fan coil. The coil was easier to maintain and clean. Only dehumidif ed outdoor air was used to of fset the space latent load. Ho wever, many four -pipe f an-coil systems are no w designed so that Tw,p  Tw,f , as shown in Fig. 28.1b, because of the following: ●

●

In locations where outdoor climate is hot and humid, the inf ltrated air most probably forms condensate in the fan coil It is less expensive as well as more convenient to control and operate when Tw,p  Tw,f.

Exhaust Air to Balance Outdoor Ventilation Air Intake First, as discussed in Sec. 20.3 and e xpressed in Eq. (20.8), at a steady state, the mass f ow rate of entering air must be equal to the mass f ow rate of leaving air, and m˙o  m˙rec  m˙inf  m˙rt  m˙ex

(28.8)

where ,m˙o m˙rec  mass f ow rate of outdoor air intak e and recirculating air , lb / min (kg / min), and m˙s  m˙o  m˙rec . If the designed entering air is exceeds the leaving air, the space will be pressurized to a higher space pressure, a portion of space air is then e xf ltrated, and a ne w balance is formed with an additional term  m˙inf . Second, for a typical of f ce b uilding, the required outdoor v entilation air intak e V˙o,sys is 0.14 cfm / ft2 (0.70 L / s m2) of f oor area. The e xhaust in the rest rooms is about 0.03 cfm / ft (0.15 L / s m2), which results in an excessive outdoor ventilation air of 0.14  0.03  0.11 cfm / ft2 (0.56 L / s m2) of f oor area. If there is no e xhaust system to exhaust this excessive amount of outdoor air to outdoors, then all the excessive outdoor air will be squeezed out through the leakage area on the exterior walls at a pressurized space-outdoor pressure difference pro. For an a verage leaky off ce building that has an e xterior wall area of 10,000 ft 2 (929 m 2) and a f oor area of 20,000 ft 2 (1858 m 2), as discussed in Sec. 23.2, its effective leakage area is 2.08 ft 2 (0.193 m 2). From Eq. (23.4), the space and outdoor pressure dif ferences to e xf ltrate all the e xcessive outdoor ventilation air are as follows:

Average leaky Leaky

Ae,l, ft2 (m2)

pro, in. WC (Pa)

2.08 (0.193) 4.16 (0.387)

0.07 (17) 0.017 (4.3)

Therefore, the following is recommended when a four-pipe fan coil system is used: ●

●

SH__ ST__ LG__ DF

For leaky buildings, outdoor ventilation air exhaust is not required. For average leaky buildings, the space pressurization due to excessive outdoor ventilation air should be calculated and analyzed. If the pressurized space-outdoor pressure difference pro  0.03 in. WC (7.5 Pa), outdoor ventilation exhaust due to excessive outdoor ventilation air is not recommended.

39445 Wang (MCGHP) Ch_28

SECOND PASS bzm 7/24/00

pg 28.13

AIR CONDITIONING SYSTEMS: SPACE CONDITIONING SYSTEMS

28.13

__RH

Part-Load Operation At cooling mode part-load operation as the zone sensible cooling load decreases, the zone temperature decreases accordingly. In a four-pipe fan-coil system, when the zone temperature sensor senses this temperature decrease, a DDC controller closes the tw o-way valve in a f an coil, reducing the chilled w ater f ow rate entering the coil. As the chilled w ater f ow reduces, the temperature of chilled water leaving the f an coil Twf rises accordingly, which raises both the temperature and humidity ratio of the conditioned (cooled and dehumidi f ed) air lea ving the f an coil Tfcp and wfcp as well as the humidity ratio of the zone air wrp, as shown in Fig. 28.1 b. As soon as the reduction in the sensible cooling ef fect of the conditioned air due to the increase of Tfcp in Qs,fcp  60V˙fcscpa (Tr  Tfcp) is just equal to the reduced sensible zone cooling load, a new balance is formed and the zone temperature is maintained at the indoor design temperture Tr. At cooling mode part-load operation, the chilled water temperature entering the f an coil and the cooling coil of the outdoor air AHU Twfe is better reset 5 to 7 °F (2.8 to 3.9 °C) higher, according to either the outdoor air temperature or the system load. Chilled water temperature reset can save compression energy and reduces the sensible cooling ef fect of the conditioned air . On the contrary , it also increases the zone relative humidty. Analyses should be done to select an appropriate value. During cooling mode part-load operation, reducing the chilled water f ow rate entering a cooling coil to match the reduction in zone or space load al ways causes a higher zone relative humidty. But reducing the supply air v olume f ow rate to match the reduction in zone or space load in VAV systems causes the zone relative humidity to remain approximately the same.

Zone Temperature Control and Sequence of Operations The zone temperature control of a four -pipe fan-coil system includes a f an-coil unit FC-1, using a separate cooling coil and a heating coil that serv e a control zone rx1 in the perimeter zone, and the corresponding outdoor air AHU which supplies the conditioned outdoor ventilation air to the fan coil FC-1. The operating mode can be divided into cooling mode, deadband mode, and heating mode. 1. When the time-of-day clock signals the four -pipe fan-coil system in the of f position, and the outdoor air AHU and f an-coil unit serving control zone rx1 in the perimeter zone is shut of f, the two-way valves for both cooling and heating coils in the f an coil FC-1 should be closed. Also the outdoor air damper and the supply fan in the outdoor air AHU (OAAHU) are shut off. 2. When the time-of-day clock signals the four -pipe fan-coil system to the on position, and the perimeter zone temperature sensor senses a zone temperature Trx 75°F (23.9 °C), both the outdoor air AHU and fan coil FC-1 call for cooling. A DDC system controller then opens the tw o-way valves, resets the discharge air temperature Tdis, and starts the supply fan motor of the OAAHU with the outdoor damper closed. After a time delay, such as 30 s, the outdoor damper of the O AAHU is fully opened. 3. A DDC terminal controller ne xt opens the tw o-way valve of the cooling coil of the f an coil, starts the fan motor and the small fan, and modulates the valve opening to maintain a zone temperature of 75°F (23.9°C). The two-way valve of the heating coil remains closed. 4. During cooling mode part-load operation, the zone sensible cooling load reduces. When the drop in zone temperature Trx1p is sensed by the temperature sensor of control zone rx1, a DDC terminal controller reduces and modulates the opening of the tw o-way valve of the cooling coil in the fan coil, to maintain a preset zone temperature of 75°F (23.9°C). Because of the reduction in the chilled w ater f ow in the cooling coil, the result is an increase in the temperature of chilled water leaving the cooling coil as well as in the temperature and humidity ratio of the cooled conditioned air lea ving the coil, which results in a higher zone relati ve humidity

rp, as shown in Fig. 28.1b. Usually, the sensible heat ratio of the space conditioning line (SHR sp) at part load is smaller than that at summer design conditions. This further increases the zone relative humidity.

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28.14

SECOND PASS bzm 7/24/00

pg 28.14

CHAPTER TWENTY-EIGHT

5. When the space cooling load is reduced to 50 percent of the design load, a DDC system controller resets the chilled w ater temperature in linear proportion from 45 to 50 °F (7.2 to 10.0 °C) according to a preset range of outdoor temperature To. Also, a DDC terminal controller reduces the fan speed in the fan coil from high to low. 6. When the zone temperature drops within the range 72.0 °F  Trx1p  75°F (22.2°C  Trx1p  23.9°C), the f an coil FC-1 is operated in deadband mode. In deadband mode, the DDC terminal controller actuates the following: ●

●

Both two-way valves for cooling and heating coils in the fan-coil are closed. The fan is still operating at low speed to provide air movement and f ltration.

7. When the zone temperature Trx1  72.0°F (22.2°C), the fan coil FC-1 is operating in heating mode. In heating mode operation: ●

●

●

The modulation of the hot w ater f ow by the DDC terminal controller should be re verse-acting; i.e., the higher the sensed zone temperature Trx1, the smaller the output; and the lo wer the sensed Trx1, the greater the output. The DDC controller modulates the tw o-way valve opening of the heating coil to maintain a zone temperature Trx1 of 72°F (22.2°C). The fan in the fan coil is operating at low speed.

8. In heating mode, the heating process in the fan coil is a horizontal line ms with a constant humidity ratio (as shown in Fig. 28.1b). The discharge temperature of the OAAHU To,dis is often reset between 80 and 100 °F (26.7 and 37.8 °C) according to outdoor temperature To through a DDC system controller. The lower the outdoor temperature To, the higher To,dis. When the outdoor temperature To drops to a preset value, the fan in the fan coil raises its speed to high speed. During heating mode part-load operation, if the outdoor humidity ratio wo is higher than the winter design conditions, the humidity ratio of the mixture of outdoor and recirculating air wm is raised accordingly. The result is a higher zone relative humidity which is more comfortable to occupants in winter Systems Characteristics System characteristics of four-pipe fan-coil (4PFC) systems are listed in Table 28.1. Calculation of Operational Parameters Consider a large off ce in an off ce building using a four-pipe fan-coil system. The outdoor ventilation air from the outdoor-air AHU is supplied to the mixing plenum of the f an-coil unit. At summer and winter design loads, the operating parameters for this four -pipe fan-coil system in this of f ce room are as follows: Summer

SH__ ST__ LG__ DF

Indoor space temperature, °F Space relative humidity, percent (wet-bulb temperature) Space cooling or heating load, Btu / h Space sensible cooling load, Btu / h Space latent load, Btu / h Part-load, operation, cooling load, Btu / h Part-load operation, sensible cooling load, Btu / h Outdoor temperature, °F Outdoor relative humidity, percent Outdoor air, cfm

78 45 (63.5°F) 29,000 26,200 15,900 13,100 90 50 200

Winter 72 9500 2700 20 80 200

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pg 28.15

AIR CONDITIONING SYSTEMS: SPACE CONDITIONING SYSTEMS

28.15

__RH

TABLE 28.1 System Characteristics of Space Conditioning Systems

Zone thermal and sound controls Control zone Control methods Control modes Heating-cooling changeover Sound control Indoor air quality Minimum ventilation air control Filters: FC or WSHPs Outdoor-air AHU Air systems Space recirculating system Type of supply fan Combined fan-motor-drive eff ciency Volume f ow control Space recirculating system total pressure, in. WC Dedicated outdoor ventilation system Type of supply fan Combined fan-motor drive eff ciency Volume f ow control Outdoor ventilation system total pressure, in. WC Refrigeration systems Type Refrigerants Evaporator Condenser Cooling tower Refrigerant f ow control Minimum performance, compressor, kW / ton Heating system Type Gas-f red boiler eff ciency Maintenance Fault detection and diagnostics

4PFC

2PFC

Multizone DDC or electric PI or on / off Automatic

Multizone DDC or electric PI or on / off Nonchangeover; changeover, auto matic or manual 35 – 45 dBA

35 – 45 dBA

WSHP Multizone DDC or electric PI or on / off Automatic 40 – 45 dBA

Guarantee minimum outdoor ventilation air supply Low eff ciency Low eff ciency Low eff ciency Medium to high eff ciency Medium to high eff ciency Fan coil Forward-curved centrifugal 25%

Fan coil Forward-curved centrifugal 25%

High-low or high-mediumlow fan speed 0.06 – 0.6

High-low or high-mediumlow fan speed 0.06 – 0.6

Constant-volume or DCV Centrifugal 45% Constant / variable-speed 3– 4

Constant-volume or DCV Centrifugal 45% Constant / variable-speed 3– 4

Centrifugal or screw HCFC-123, HCFC-22, or HCFC-22, HFC-134a Chiller Water-cooled Open-circuit

Centrifugal or screw HCFC-123, or HFC-134a Chiller Water-cooled Open-circuit

Orif ce, f oat, or Ori expansion valve 0.5 – 0.7

f ce, f oat, or Capillary expansion valve 0.5 – 0.7

Hot water or electric heating AFUE 80 to 93% More site maintenance Chiller

Hot water or electric heating AFUE 80 to 93% More site maintenance Chiller

Water-source heat pump Forward-curved centrifugal Console, 25% Core, 30 to 40% Console, multi speed Core, constant speed Console, 0.06 – 0.6 Core,  3.0 Constant-volume or DCV Centrifugal 45% Constant / variable-speed 3– 4

Reciprocating, rotary, or scroll HCFC-22, HFC-407C, or HFC-410A DX coil Water-cooled Open-circuit, or closed-circuit tube 10 – 15 EER

WSHP AFUE 80 to 93% More site maintenance Core WSHPs

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28.16

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pg 28.16

CHAPTER TWENTY-EIGHT

1. Assume that the temperature of outdoor air supplied from the O AAHU Tps is 61°F (16.1°C) and the relative humidity is 81 percent. From the psychrometric chart, hps  24.8 Btu / lb (57.7 kJ / kg). At summer design conditions, the enthalpy of space air is 28.8 Btu / lb (67.0 kJ / kg), and outdoor air density ps  0.075 lb / ft3 (1.2 kg / m3); from Eq. (28.7) the estimated fan-coil cooling capacity is Qc,fc  Qrc  60 V˙ pps(hr  hps)  29,000  60(200)(0.075)(28.8  24.8)  25,400 Btu / h (7442 W) The estimated fan-coil sensible cooling capacity is Qs,fc  Qrs  60V˙ ppscpa(Tr  Tps)  26,200  60(200)(0.075)(0.243)(78  61)  22,482 Btu / h (6587 W) If the chilled water supplied to the f an coils and the outdoor air AHU are both at 45 °F (7.2°C) during summer design conditions, for an entering air of 78 °F (25.6°C) dry b ulb and 63.5 °F (17.5°C) wet bulb temperatures and a chilled water temperature rise of 10°F (5.6°C), one manufacturer’s catalog gives the following capacities for a size 03 fan-coil unit: Total cooling capacity Sensible cooling capacity

7.2 MBtu / h (2.1 kW) 6.3 MBtu / h (1.8 kW)

Because the external pressure drop in the supply duct after the f an coil is 0.06 in. WC (15 Pa) and the fan coils are installed at sea level, no volume f ow corrections are needed. Four 03 fan-coil units are selected, so that the total cooling capacity is Qc,fc  4  7200  28,800 Btu / h (8438 W) which is greater than the estimated cooling capacity of 25,400 Btu / h (7442 W) at summer design load. The sensible cooling capacity of fan coils is Qs,fc  4  6300  25,200 Btu / h (7384 W) which is also greater than the estimated sensible cooling capacity of 22,482 Btu / h at summer design load. 2. Draw line r-fc from the conditioned zone point r with known SHR s  26,200 / 29,000  0.9, and a fc  1 / 13.7  0.73 lb / ft3. The temperature of supply air lea ving the fan-coil unit Tfc, in °F (°C), can be calculated as Tfc  Tr 

Q rs 60V˙fcfccpa

 78  26,200 / 60  4  300  0.073  0.243  57.5°F (14.2°C)

(28.9)

V˙fc  where corrected volume f ow rate of fan coil, cfm (L / s) fc  air density at fan outlet in fan coil, lb / ft3 (kg / m3) Point fc can then be determined. From the psychrometric chart, hfc  23.3 Btu / lb (54.2 kJ / kg). 3. Let Twp be the w ater temperature rise in the cooling coil of the outdoor air AHU. It can be assumed that the temperature of air lea ving the cooling coil of the outdoor air AHU Tpc, in°F (°C), is SH__ ST__ LG__ DF

Tpc  Tw,p  Twp For Twp  10°F, Tpc  45  10  55°F (12.8°C).

(28.10)

39445 Wang (MCGHP) Ch_28

SECOND PASS bzm 7/24/00

pg 28.17

AIR CONDITIONING SYSTEMS: SPACE CONDITIONING SYSTEMS

28.17

__RH

Generally, the cooling coil may ha ve six or eight ro ws. The relative humidity of air lea ving the cooling coil pc  98 percent, and the humidity ratio wpc can thus be determined. F or Tpc  55°F (12.8°C) and pc  98 percent, wpc  0.0092 lb / lb (0.0092 kg / kg) and hpc  23.2 lb / lb (54.0 kJ / kg). For simplicity, the fan power heat gain for the OAAHU can be assumed as 2°F, and the duct heat gain can be estimated at 4°F, so the supply temperature of the OAAHU Tps, in °F, is Tps  Tpc  2  4  61°F (16.1°C)

(28.11)

Because wps  wpc, state point ps can also be determined. 4. For a f an coil whose outdoor air is mix ed with recirculating air in the mixing plenum of the fan coil, draw line r-ps. From Fig. 28.1b, V˙p m-r 200    0.17 ps-r ˙ 4  300 Vfc

(28.12)

Point m can thus be determined. In this case, Tm  75.0°F (23.9 °C), and wm  0.0092 lb / lb (0.0092 kg / kg). Because Tcf  Tm  0.5  75  0.5  75.5°F (24.2 °C), and wcf  wm  0.0092 lb / lb (0.0092 kg / kg), point cf can also be determined. From the psychrometric chart, hcf  28.2 Btu / lb (65.6 kJ / kg). 5. If cf  1 / 13.7  0.073 lb / ft3 (1.168 kg / m3), the cooling coil load Qc,fc and sensible cooling coil load Qs,fc of the fan-coil unit, both in Btu / h, can be calculated as Qc,fc  60V˙ fccf (hcf  hfc)

(28.13)

 60  4  300  0.073(28.2  23.3)  25,754 Btu / h (7546 W) Qs,fc  60V˙ fccf cpa (Tcf  Tfc)

(28.14)

 60  4  300  0.073  0.243(75.5  57.5)  22,990 Btu / h (6736 W) where hcf,hfc  enthalpy of air at fan outlet and supply outlet, as shown in Fig. 28.1a, Btu / lb (J / kg) Tcf,Tfc  temperature of air at fan outlet and supply outlet, °F (°C) Check the calculated cooling coil load and sensible load against the selected f an coil’s cooling and sensible capacity. The calculated coil load Qc,fc of 25,754 Btu / h (7546 W) is v ery near to the estimated capacity of 25,400 Btu / h (7442 W), and both are less than the selected capacity Qc,fc of 28,800 Btu / h (8438 W). Similarly, the calculated sensible coil load Qs,fc is 22,990 Btu / h (6736 W). It is slightly greater than the estimated 22,482 Btu / h (6587 W) because of the f an power heat gain. Both are less than the selected capacity Qs,fc of 24,000 Btu / h (7032 W). 6. Calculate the cooling coil load in the outdoor air AHU Qcp, in Btu / h (W). From the psychrometric chart, To  90°F, o  50 percent, and ho  38.5 Btu / lb, so Qcp  60V˙ pps(ho  hps)  60  200  0.075(38.5  23.2)  13,770 Btu / h (4035 W)

(28.15)

The cooling coil load due to the outdoor air in the O AAHU Qcp is about one-half of the f an coil’s cooling coil load. This includes the cooling and dehumidifying load of outdoor air , the system heat gain of the dedicated outdoor ventilation air system, and a portion of the zone cooling load. 7. During cooling mode part-load operation, the zone temperature is still maintained at 78 °F (25.6°C); that is, Trp  Tr. The fan in the f an coil can be switched from high speed to lo w speed while chilled water is reset to a higher temperature. If the zone sensible load ratio is reduced to 50 percent of the full-load value and the volume f ow rate of the fan at low speed is only 70 percent of the high speed, then the temperature of air lea ving the fan coil at part load Tfc,p, in °F (°C), can be

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28.18

pg 28.18

CHAPTER TWENTY-EIGHT

calculated as Tfc,p  Trp 

 78 

Q rsp

(28.16)

60  0.7V˙fccf cpa 0.5(26,200) 60  4  300  0.7  0.073  0.243

 63.3F (17.4C)

where Qrsp  zone sensible cooling load at part load, Btu / h (W) Trp  zone air temperature at part load, °F (°C) At part-load operation, the relative humidity of supply air lea ving the f an coil fc,p may vary between 80 and 90 percent. If fc,p  88 percent, point fcp can be plotted on the psychrometric chart as shown in Fig. 28.1b. From the chart, hfcp  27.2 Btu / lb (63.3 kJ / kg). 8. During cooling mode part-load operation, the sensible heat ratio of space conditioning line SHRs  13,100 / 15,900  0.82. Draw line fcp-rp from fcp with a sensible heat ratio of the space conditioning line at part load SHR sp  0.82. Line fcp-rp intersects the zone temperature line Trp at rp. Therefore, point rp can be determined. From the psychrometric chart, rp  57.5 percent. 9. When the space cooling load is reduced to 50 percent of design load, if the temperature of chilled water supplied to fan coils Tw,f and that to the outdoor air AHU Tw,p are reset to 50°F (10°C) at part-load operation, and the chilled water temperature increase in the cooling coil of the outdoor air AHU is still 10 °F, then the condition of supply air from the outdoor air AHU at part-load point psp can be determined as in full-load operation, Tpsp  60  2  4  66°F (18.9°C), and wpsp  0.0110 lb / lb (0.0110 kg / kg). Draw line psp-rp. The condition of the mixture at part load, point mp, can be determined from the following relationship: V˙p rp-mp 200    0.24 rp-ps ˙ 4  300  0.7 Vfc

(28.17)

From the psychrometric chart, Tmp  75.3°F (24.1 °C), and wmp  0.0116 lb / lb (0.0116 kg / kg). Because Tcfp  Tmp  0.5°F  75.3  0.5  75.8°F, and wcfp  wmp  0.0116 lb / lb, from the psychrometric chart, hcfp  31.0 Btu / lb (72.1 kJ / kg). 10. Fan-coil cooling coil load at part load Qcfcp and sensible coil load at part load Qsfcp, both in Btu / h (W), can be calculated as Q cfcp  60V˙fccf (h cfp  h fcp )

(28.18)

 60(4)(300)(0.7)(0.073)(31.0  27.2)  13,980 Btu / h (4096 W) Q sfcp  60V˙fccf cpa(Tcfp  Tfcp)

(28.19)

 60(4)(300)(0.7)(0.073)(0.243)(75.8  63.3)  11,176 Btu / h (3275 W) 11. At winter heating mode design load operation, if only tw o fan-coil units are operated and hfg,o  1061 Btu / lb (2468 kJ / kg), the humidity ratio dif ference between zone air and the f an-coil supply air is wr  ws  SH__ ST__ LG__ DF



Q rl

(28.20)

60 V˙fc hf h fg,o 2700 60  2  300  0.073  1061

 0.00097 lb / lb (0.00097 kg / kg)

39445 Wang (MCGHP) Ch_28

SECOND PASS bzm 7/24/00

pg 28.19

AIR CONDITIONING SYSTEMS: SPACE CONDITIONING SYSTEMS

28.19

__RH

where ws  humidity ratio of f an-coil supply air, lb / lb (kg / kg). As shown in Fig. 28.1 b, ws  wm  wfc, and from the psychrometric chart, the humidity ratio of outdoor air wo  0.0017 lb / lb. Therefore, V˙p wr  ws 200    0.33 wr  wo 600 ˙ Vfc

(28.21)

So wr  wo 

wr  ws 0.33

 0.0017 

0.00097 0.33

 0.0046 lb / lb (0.0046 kg / kg) and ws  wm  wr  0.00097  0.0046  0.00097  0.0036 lb / lb (0.0036 kg / kg). 12. Because at heating mode operation Tr  72°F (22.2°C), from the psychrometric chart, the relative humidity at winter design condition r  28 percent. 13. If air lea ves the preheating coil of the outdoor air AHU at Tph  83°F and Tpf  Tph  fan temperature rise  83  2  85°F, because wo  wpf  0.0017 lb / lb, from the psychrometric chart, air density at the f an supply outlet in the outdoor -air AHU pf  1 / 13.75  0.073 lb / ft3. The preheating coil load Qcph, in Btu / h (W), can be calculated as Q cph  60 V˙ppf cpa(Tph  To )

(28.22)

 60  200  0.073  0.243(83  20)  13,411 Btu / h (3929 W ) 14. If the heat loss of outdoor air supply duct is 0.5 °F, Tps  85  0.5  84.5°F, and wps  wpf  0.0017 lb / lb, then point ps can be determined. Dra w line ps-r. Because wm  0.0036 lb / lb (0.0036 kg / kg), point m can be plotted on the psychrometric chart. From the chart, Tm  76.2°F (24.5°C). 15. The temperature of the f an-coil supply air at winter design load Ts, in °F (°C), can be calculated as Q rh Ts  Tr  (28.23) 60 V˙fccf cpa  72 

9500 60  2  300  0.073  0.243

 72  14.9  86.9F (30.5C)

As ws  wm  0.0036 lb / lb (0.0036 kg / kg), point s can be determined, and lines r-s and m-s can be drawn. The temperature dif ference Ts  Tr  86.9  72  14.9°F (8.3 °C), which is smaller than the supply air temperature dif ference for w arm air during heating mode 15 °F (8.3 °C) specified by ASHRAE handbooks to pre vent an e xcessive b uoyancy ef fect, as discussed in Sec. 18.6. 16. The air temperature at the f an outlet in the f an coil Tcf  Tm  0.5  76.2  0.5  76.7°F. The heating coil load in the fan-coil unit Qch, in Btu / h (W), can be calculated as Q ch  60V˙fccf cpa(Ts  Tcf )

(28.24)

 60  2  300  0.073  0.243(86.9  76.7)  7152 Btu / h (2095 W) 17. At winter heating mode part-load operation, the warm supply air from the f an coil, point sp, moves along the horizontal line m-sp and is supplied at a lower temperature Tsp in order to maintain the required zone temperature Trp.

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Applications Four-pipe f an-coil systems with a dedicated v entilation and space recirculating air system are widely used in hotels and motels where the e xhaust air system for bathrooms can easily balance part of the outdoor ventilation air supply. They are also used in hospitals, schools, and off ces.

28.4 TWO-PIPE FAN-COIL SYSTEMS Two-Pipe Systems A two-pipe system equipped with a supply pipe and a return pipe is sho wn in Fig. 28.4. In such a system, chilled water is supplied to the coil to cool and dehumidify the air during cooling mode operation. In heating mode operation, chilled water is changed o ver to hot w ater and then supplied to the coil to heat the air. A two-way valve is usually installed before the coil inlet because it costs less, is easier to install, and sa ves pump po wer at part-load operation when the w ater f ow rate is reduced. A DDC controller is often used to modulate the water f ow at part load.

Nonchangeover Two-Pipe Systems When a tw o-pipe fan-coil system is used to serv e a perimeter zone in a b uilding, changeover from chilled water to hot water or vice versa is a troublesome process and may take several hours. Therefore, in locations where winter weather is moderate, a nonchangeover two-pipe fan-coil system may be used. In a nonchangeo ver tw o-pipe system, chilled w ater is supplied to the f an coil throughout the year when the space is occupied. Warm outdoor air is supplied to the f an coil in winter to offset the space heating load in the perimeter zone, and in spring and f all when space heating is required. In such an arrangement, various zones at dif ferent orientations in the b uilding that need cooling and heating simultaneously during spring and fall can be served, as shown in Fig. 28.5a.

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FIGURE 28.4 Supply and return mains in a two-pipe fan-coil system.

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In a typical nonchangeo ver two-pipe fan-coil system, as shown in Fig. 28.5 a, chilled water Tel enters the coil at a temperature of about 45 °F (7.2 °C) at summer design conditions, whereas the temperature of outdoor air supply Tps may be maintained at 62 °F (16.7°C). At cooling mode partload operation, Tel supply for both fan-coil units and the outdoor air AHU can be reset gradually, up to 52 °F (11.1 °C), when the space sensible cooling load drops to 30 percent of the design v alue. When outdoor temperature To drops below 70 °F (21.1 °C), the outdoor air supply temperature Tps begin to rise. The lower To, the higher Tps. If the space heating load of an y room or control zone in the perimeter zone is of fset entirely by the heated outdoor air in a nonchangeo ver two-pipe fan-coil system, the required volume f ow rate of heated outdoor air V˙p, in cfm [m3 /(60 s)], can be calculated as V˙p 



Q rh 60pscpa (Tps  Tr) (AexUm  60V˙infocpa)(Tr  To) 60pscpa(Tps  Tr)

(28.25)

where Qrh  room heating load, Btu / h (W) Aex  total area of building shell in that room, ft2(m2)

FIGURE 28.5 Operating parameters of air and w (b) changeover.

ater in a tw o-pipe f an-coil system: (a) nonchangeo ver;

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CHAfYfER

TWENTY-EIGHT

140

120

~ 0

~ ~ '§ & § .. "

100

~ ~ -g '" .. "~

80

~ E "t: ~

60

40 0

20

60

80 Outdoor

100 temperature

1;, , of

(b) FIGURE 28.5

(Continued)

Urn = weighted average of overall heat-transfer coefficient of building shell in that room, Btulh .ft2.oF (W Im2. OC) Vinf = volume flow rate of infiltrated air, cfm [m31(60 s)] PO'Pps= density of outdoor air and outdoor air supply, lblft3 (kglm3) Tr = zone air temperature, oF (OC) Tps = supply temperature of conditioned outdoor air, oF (OC) The air transmission ratio is defined as the ratio of outdoor air volume flow rate supplied to a room in the perimeter zone Vp, in cfm [m31(60 s)], to the transmission loss per degree of outdoor-indoor temperature difference ~T tran'Btulh .oF (W 1°C). According to Eq. (28.25), it can be calculated as

-= Vp ~Ttran

Vp AexUm

+ 60VinfPoCpa

(28.26) Tr60ppscpa(Tps

To -Tr)

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For all the rooms in the perimeter zone supplied by the same outdoor v entilation air system in a two-pipe fan-coil system, the temperature difference between room air and outdoor air Tr  To is a constant at any time during operation. If the temperature drop due to the heat loss of the outdoor air supply duct is ignored, for the same outdoor v entilation air system, the supply air temperature difference Tps  Tr is again a constant. Because air density ps and specif c heat of supply air cpa can also be considered constant, and if the zone heating load in the perimeter zone is of fset entirely by the conditioned outdoor air , the air transmission ratio V˙p / Ttranfor all the control zones in the perimeter zone in the same outdoor v entilation air system for a tw o-pipe nonchangeover f an-coil system must be a constant. The current trend is to ha ve an adequate amount of outdoor air supply for a better indoor air quality. Because the conditioned outdoor air supply must be higher than the outdoor v entilation air requirement, this often means that a nonchangeo ver two-pipe fan-coil system has suf f cient conditioned outdoor air for winter heating. In massi vely constructed buildings, the air transmission ratio can be smaller, such as 0.7 V˙p / Ttran , because of the heat storage effect of the building shell. One of the primary dra wbacks of a nonchangeo ver two-pipe fan-coil system is the w aste of energy during simultaneous heating and cooling in some control zones during winter heating. In a cold winter in the northern hemisphere, for a control zone f acing south, warm outdoor air heating, solar radiation, and some internal loads often combine to result in a higher zone temperature than that for a control zone facing north in the same outdoor ventilation air system. Chilled water is then admitted to the f an coil to cool the zone air simultaneously . Simultaneous heating and cooling for the same control zone must be a voided. The waste of energy of the conditioned outdoor v entilation air system in a nonchangeo ver two-pipe fan-coil system is similar to the w aste of ener gy in an air skin system, discussed in Sec. 21.3.

Changeover Two-Pipe Systems During the cooling mode operation of a changeo ver tw o-pipe f an-coil system, the space cooling load is offset by the combined cooling ef fect of the cooled outdoor v entilation air and chilled water supplied to the f an coil, as shown in Fig. 28.5 b. As discussed in Sec. 7.6, ASHRAE / IESNA Standard 90.1-1999 specif es that two-pipe fan-coil changeover systems are acceptable when the design dead band width between changeo ver from one mode to the other is at least 15 °F (8.3°C) outdoor air temperature. It also will allo w operation in one mode at least four hours before changeo ver to the other mode. At the changeover point, reset controls allow heating and cooling supply temperatures to be no more than 30°F (16.7°C) apart. During the fall season, before the changeover from cooling mode operation to heating mode operation, when outdoor temperature To is higher than changeover temperature Tco, the supply temperature of outdoor v entilation air increases according to To when To falls below 70°F (21.1°C). The heating effect of the conditioned outdoor v entilation air is able to of fset the individual zone heating load in the perimeter zone. Chilled w ater is used to pro vide cooling and dehumidifying when needed. Meanwhile, the chilled water supplied temperature Tel should be reset between 45 and 52°F (7.2 and 11.1°C) at part-load operation before changeover. During changeover, the chilled w ater may be reset to 52 °F (11.1°C), and it is changed o ver to hot water at a temperature higher than 80 °F (26.7°C). Meanwhile, the warm air supply is changed over to cold air supply , typically at a temperature of 50 °F (10.0°C). The changeover process w as discussed in Sec. 7.6, and the outdoor changeover temperature Tco can be calculated from Eq. (7.5). Changeover may require se veral hours. To prevent two to three changeo vers within the same day , operation in one mode lasts at least four hours before changeover to the other mode, and a tolerance of about  2°F (1.1°C) is often used. There are many factors that inf uence the changeover temperature Tco in actual practice. Therfore, calculated Tco should be modi f ed according to actual operating experience. When To  Tco, a changeover two-pipe fan-coil system is in heating mode operation. The hot water temperature is reset according to outdoor temperature To. Cold outdoor air is supplied, typically at 50 °F (10.0°C). Any zone cooling load of an indi vidual control zone in the perimeter zone can be of fset by the combined ef fect of the cold outdoor supply air and the transmission and

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CHAPTER TWENTY-EIGHT

inf ltration loss. Zone heating loads are of fset by the hot w ater supplied to the f an coils. During heating mode operation, the refrigeration system that serv es the perimeter zone is usually shut down. However, the changeover two-pipe fan-coil system still operates with a cold source of outdoor air and a heat source of hot water simultaneously. In spring, heating mode operation changes o ver to cooling mode operation at outdoor temperature Tco only when the combined ef fect of outdoor v entilation air cooling and the transmission and inf ltration loss does not of fset the cooling load of e very individual control zone in the perimeter zone. In a tw o-pipe f an-coil system, except for the changeo ver from heating mode to cooling mode operation or vice versa, operating characteristics of the fan-coil unit and the space recirculating system, exhaust air balance, and the zone temperature control are similar to those in a four -pipe fancoil system. System Characteristics System characteristics of a two-pipe fan-coil system (2PFC) are listed in Table 28.1. Applications Because of the w aste of ener gy in nonchangeo ver two-pipe fan-coil systems and the dif f culties in changeover operations, the applications of two-pipe fan-coil systems are comparatively less.

28.5 WATER-SOURCE HEAT PUMP SYSTEMS System Description In a water-source heat pump system, some water-source heat pumps located in the shady side of the perimeter zone of a b uilding may e xtract heat from a w ater loop to heat the supply air , and other water-source heat pumps in the core part of the b uilding may reject heat to the w ater loop to cool the supply air . Excess heat is therefore transferred from the core of the b uilding to the perimeter zone of the building. A water-source heat pump system conserv es more energy than many other air conditioning systems only when simultaneous heating and cooling occur in a building and therefore excessive heat from the core part is reco vered and transferred to the perimeter zone of the b uilding. A typical closed-circuit w ater-source heat pump system is illustrated in Fig. 28.6. It consists of water-source heat pumps, an evaporative water cooler, a boiler or w ater heater, two water circulating pumps, an expansion tank, piping, necessary accessories, and controls. One of the w ater circulating pumps is the lead pump, the other is a standby. A storage tank is optional. Operating Characteristics

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During hot summer, when the outdoor wet-b ulb temperature is 78 °F (25.6°C), all the water-source heat pumps are operting in cooling mode. The condenser water leaves the closed-circuit evaporative water cooler at 95°F (35.0°C) and enters the water loop. After all heat is rejected from water-source heat pumps, the condenser w ater returns to the e vaporative water cooler at a temperature of 105 °F (40.6°C) and cools to 95°F (35.0°C) again. During moderate weather, water-source heat pumps serving the shady side of the b uilding are in heating mode, whereas those serving the sunny side of the building are in cooling mode. The rise or fall in condenser water temperature depends on the ratio of the number of cooling units to the total number of units, or, more exactly, the ratio of heat rejection to heat e xtraction. If heat rejected from the cooling units is greater than heat e xtracted by the heating units, the average temperature of the

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FIGURE 28.6 A typical closed-circuit water-source heat pump system: (a) schematic diagram; (b) water-source heat pump.

water loop tends to rise, and vice versa. When the condenser water temperature entering the evaporative water cooler Te,ev rises to 90°F (32.2°C), the cooler starts to operate and brings the condenser water temperature down to a lower value (say, 87°F, or 30.6°C). During cold weather, some water-source heat pumps serving the core of the b uilding still operate in cooling mode. If the heat e xtracted by the units operating in heating mode is greater than the heat rejected, the average temperature of the w ater loop drops. When the condenser water temperature Te,ev falls below 60°F (10.6°C), the boiler (or w ater heater) is ener gized. Heat is added to the water loop to raise Te,ev to 63°F (17.2°C) or higher. A water-source heat pump system reco vers the heat rejected to the w ater loop by the core units or by the units on the sunn y side and supplies it to the units serving the perimeter zone in which heating is required to offset heat losses. ASHRAE / IESNA Standard 90.1-1999 specif es that in a water-source heat pump system, watersource heat pumps are connected to a common w ater loop with central de vices of heat rejection (such as a cooling to wer) and heat addition (such as a boiler). Such a system shall ha ve controls to provide heat pump w ater supply temperature with a dead band of at least 20 °F (11.1°C) between initiation of heat rejection and heat addition by the cooling to wer and boiler (for e xample, 90  60  30°F or 16.7°C). Loop Temperatures Water-source heat pumps (WSHPs) are rated according to Air-Conditioning and Refrigeration Institute (ARI) Standard 320 at the follo wing entering condenser w ater temperatures: for cooling mode

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performance 85°F (29°C) and for heating mode performance 70 °F (21°C). Pietch (1991) listed the performance of WSHPs of a WSHP system at various loop temperatures: Loop temperature, °F (°C) 90 (32.2) 85 (29.4) 80 (26.7) 75 (23.9) 70 (21.1) 65 (18.3) 60 (15.6)

Rating mode Cooling Heating

Relative cooling capacity

EER / EERc,ra

Heating capacity / ARIrated c cap

COP / COPh,ra

0.98 1.00 1.02 1.04 1.06 1.08 1.10

0.94 1.00 1.06 1.12 1.18 1.24 1.30

1.50 1.45 1.40 1.35 1.30 1.25 1.20

1.08 1.06 1.04 1.02 1.00 0.98 0.96

Here EER c,ra indicates the ener gy eff ciency ratio of the WSHP at the cooling mode rated conditions, ARIrated c cap indicates the ARI-rated cooling capacity , and COP h,ra indicates the COP of the WSHP at heating mode rated conditions. WSHPs at cooling mode operation require lo wer condenser water temperature (CWT) to reduce energy use, and WSHPs at heating mode need higher CWT for a higher COP. Optimizing loop temperatures depend on the condition of zone loads, the number of WSHPs in cooling and heating mode operations, and the outdoor weather. Because a water loop may have WSHPs in both cooling and heating mode operations, the water loop temperature operating limit is usually maintained between 60 and 90°F (15.6 and 32.2°C). Gottfried et al. (1997) suggested when the CWT has been lo wered to a rather lo wer value in a late afternoon, this allows the CWT to rise to a le vel that maintains WSHP eff ciencies and at the same time keep evaporative cooler (cooling to wer) energy use to a minimum during the last w orking hours of the day.

Water-Source Heat Pumps Three types of WSHPs are widely used in WSHP systems: ●

●

●

SH__ ST__ LG__ DF

Console type for perimeter zones. Most ha ve a cooling capacity of 1 ton (12,000 Btu / h, or 3.5 kW) or less. Vertical WSHPs often installed in mechanical rooms. They are used to serv e the core (interior zone) of the b uildings and are connected with ducts. Vertical units often ha ve cooling capacity from 10 up to 24 tons. Ceiling-mounted horizontal WSHPs, often for special rooms. Ceiling units ha ve a cooling capacity, usually between 0.5 and 5 tons.

Most WSHPs have a heating capacity about 10 percent higher than their cooling capacity. A water-source heat pump usually consists of an air coil that is a f nned coil to condition the air; a double-tube water coil to reject or e xtract heat from the w ater loop; a forw ard-curved centrifugal fan, which is often located downstream of the air coil; single or twin hermetic compressors; a short capillary tube; a reversing valve; an outer casing; controls; and accessories. A two-speed fan motor is often used for better capacity control. A typical w ater-source heat pump is sho wn in Fig. 28.6 b. In a vertical water-source heat pump, the centrifugal fan is usually located at the top outlet, and the hermetic compressor is often mounted in the ottom of the unit. In a horizontal unit, the centrifugal fan is usually located at the end of the unit. During cooling mode operation, the air coil acts as an e vaporator and the w ater coil as a condenser. Air in a typical w ater-source heat pump is cooled and dehumidi f ed at the air coil from an entering dry-b ulb temperature of 80 °F (26.7 °C) and wet-b ulb temperature of 67 °F (19.4 °C) to a

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leaving condition of 60°F (15.6°C) dry-bulb and 57°F (13.9°C) wet-bulb temperatures. The suction temperature is about 43°F (6.1°C). On the other hand, in heating mode operation, the air coil acts as a condenser and the w ater coil as an e vaporator. During cold weather air is heated at the air coil from 70°F (21.1°C) to an off-coil temperature of 100 °F (37.8°C). The suction tmperature in winter is about 45°F (7.2°C). Sometimes, electric resistance heating, in steps, is added between the air coil and the centrifugal fan instead of using a centralized water heater.

Energy Performance and Energy Use Intensity of WSHPs ASHRAE Standard 90.1-1999 mandatorily speci f es the minimum ef f ciency requirements for w ater-source heat pumps as follows: Operation Cooling

Heating

Capacity, Btu / h

Rating conditions

Qrc  65,000

85°F entering water 86°F entering water 17,000  Qrc  65,000 85°F entering water 86°F entering water 65,000  Qrc  135,000 85°F entering water 86°F entering water Qh  135,000 (heating capacity) 70°F entering water 68°F entering water

Minimum eff ciency 9.3 EER 9.3 EER 10.5 EER 3.8 COP

Eff ciency as of 10 / 29 / 2001 11.2 EER 12.0 EER 12.0 EER 4.2 COP

Gottfried et al. (1997) reported a new high-eff ciency WSHP with an EER of 14.9 (4.33 COP with a cooling capacity from 21,000 ro 57,000 Btu / h, or 6.2 to 16.7 kW) and a high-eff ciency motor with a minimum eff ciency of 90 percent. According to DOE/EIA Commercial Buildings Consumption and Expenditures 1995, the energy use intensity (EUI) for heat pumps (including mainly w ater-source heat pumps and air -source heat pumps) in commercial buildings in 1995 in the United States was 28.1 kBtu / ft2 yr (8.2 kWh / ft2 yr or 88 kWh / m2 yr) which is the smallest of all the air conditioning systems.

Closed-Circuit Evaporative Water Cooler A closed-circuit e vaporative w ater cooler (CCEWC), sometimes called a closed-circuit cooling tower, resembles a cooling to wer. However, condenser water in a cooling to wer is an open-circuit system, whereas in an e vaporative water cooler the condenser w ater is a closed-circuit system (see Fig. 28.6a). Condenser water f ows through a coil onto which recirculating water is sprayed. Heat is rejected from the condenser w ater through the tube w all and is absorbed by the v aporized liquid at the outer surf ace of the coil. A closed-circuit e vaporative water cooler is usually located outdoors, most probably on the rooftop. Both centrifugal fans and propeller fans can be used in a closed-circuit evaporative water cooler. The capacity of a centrifugal f an is more easily controlled with a damper and mak es less noise; therefore, it is more frequently used. The capacity of a closed-circuit e vaporative w ater cooler can be modulated by the follo wing methods: ●

●

●

●

Convective cooling — opening of the top damper only Natural-draft e vaporative cooling — spraying of recirculating w ater and opening of the top damper Forced-draft evaporative cooling — operating the fan and the water spraying Forced-draft evaporative cooling with damper modulation

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CHAPTER TWENTY-EIGHT

Cooling tower (open-circuit) 3

2

1 Cooling tower pump

T3

4 5

Plate-and-frame heat exchanger

Roof

Gas-fired water heater T1

T2 Optional storage tank

Closed-circuit water-loop pumps

Console WSHP Vertical WSHP

Vertical WSHP

T4 T5

Console WSHP

FIGURE 28.7 A water-source heat pump system using an open-circuit cooling tower.

The motive for a closed-loop w ater system is to pre vent fouling of the w ater coils in w ater-source heat pumps. Closed-circuit evaporative water coolers were widely used in the early years of the application of WSHP systems. However, CCEWCs have the following disadvantages: 1. According to Cooper (1994), the maintenance of spray nozzles, pumps, and heat exchangers of a CCEWC required twice amount of service of an open-circuit cooling tower. 2. According to Kush (1990), f eld experience in an installation in Stamford, Connecticut, showed that the heat and pumping ener gy losses from the outdoor e vaporative water cooler and piping work were about one-half of the energy input to the electric boiler in January 1988. 3. For locations where the outdoor temperature drops belo w 32°F (0°C), freeze protection of the outdoor portion of the water loop must be considered. The current trend is as follows: ●

●

●

SH__ ST__ LG__ DF

Divide the w ater loop of a WSHP system into tw o portions: closed-circuit and open-circuit. A plate-and-frame heat e xchanger is used to connect these tw o part and transfers heat ener gy from the closed circuit to open circuit, as shown in Fig. 28.7. Use a cooling to wer (open-circuit) instead of a closed-circuit e vaporative water cooler, and two sets of water circulating pumps, one for the closed-circuit and the other for the open-circuit cooling to wer and accessories. Set only the cooling to wers and the necessary pipe work and accessories in the open circuit outdoors. Properly insulate the outdoor pipe work and equipment. Pro vide freeze protection for the outdoor open-circuit water loop, including the addition of a certain percentage of ethylene glycol solution to the water loop. Ethylene glycol is expensive. Inhibitors must be added at the same time to resist corrosion. Ethylene glycol solution reduces the heat-transfer coeff cient as well as the heating and cooling capacities of the equipment.

ASHRAE / IESNA Standard 90.1-1999 speci f es that for climates with HDD65 greater than 1800 hours annually, if a closed-circuit tower is used, then either an automatic v alve shall be installed to bypass all but a minimal volume f ow of water around the tower for freezing protection or low-leakage positive closure air dampers shall be pro vided. If an open-circuit to wer is used directly in the

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heat pump water loop, then an automatic valve shall be installed to bypass all heat pump water f ow around the tower. If an open-circuit to wer is used in conjunction with a separate heat e xchanger to isolate the tower from the heat pump w ater loop, then the heat loss shall be controlled by shutting down the circulating pump on the outdoor cooling to wer loop during cold winter when the to wer is not in use.

Water Heater A water heater or a hot w ater boiler is used to pro vide heat energy to the w ater loop to maintain a preset operating temperature. Electric boiler is most widely used. It is easiest to install and maintain. In locations where electric rates are high, electric boilers require higher operating costs than gas- or oil-f red boilers. For better capacity control, electric load in an electric boiler is often di vided into stages, as discussed in Sec. 8.4. The heating capacity is controlled in steps by sensing the w ater temperature entering the electric boiler. Gas- or oil- f red heaters are comb ustion de vices containing comb ustion chambers. Water at a temperature of 60 °F (21.1°C) from the w ater loop cannot directly contact one side of the comb ustion chamber and a f ame on the opposite side. It will cause condensation and deteriorate the w ater heater. Water contacting one side of the comb ustion chamber must be recirculated at a temperature of 140°F (60.0°C) or above. Loop water can then be heated by mixing.

Storage Tanks Whether a storage tank should be used in a WSHP system depends on the local electric rate structures and a careful analysis of initial costs and operating cost. The principle of thermal storage is covered in Chap. 31. Because of the electric dere gulation as well as the application of real-time pricing (RTP) rate structures, the provision of a storage tank in a WSHP system is often cost-ef fective in man y applications. Heat rejected to the w ater loop at a temperature between 60 °F (15.6°C) and 90°F (32.2°C) can be stored in the storage tank to of fset nighttime heat losses if the w ater loop operates at night with a setback space temperature. Many utilities offer lower electric rates at of f-peak times. A cooling tower or evaporative cooler can be used to cool the water to a lower temperature. It can then be stored in the tank for use during on-peak hours in summer to reduce costs. An electric boiler also may be ener gized at night to raise the water temperature to 180 °F (82.2°C) or higher . High-temperature stored w ater can be used in winter peak hours to reduce electricity costs.

Air Systems and Maintenance A WSHP system must have a dedicated outdoor ventilation system to provide ventilation air for occupants. Outdoor ventilation air can either be supply to the return plenum mixing with the recirculating air or supply to the conditioned space directly . The centrifugal supply f an in a WSHP itself forms a space recirculating system. Console WSHPs have no ducts. The operating characteristics of the air systems in a WSHP system are similar to those in a four-pipe fan-coil system. According to Cooper (1994), maintenance in a WSHP system consists of mainly cleaning and changing f lters and replacing belts and seal bearings. Sometimes a WSHP is remo ved and the condenser is acid-cleaned. There is also maintenance required for cooling towers, pumps, and water heaters. F or the 3 million ft 2 (278,800 m 2) of commercial of f ce b uildings that ha ve been maintained, records sho wed that on WSHPs wih a cooling capacity of 1 ton (3.5 kW) and less, the replacement percentage is less than 0.5 percent per year . For WSHPs of 2 to 5 tons (7 to 17.6 kW), the rate is slightly higher, for WSHPs of 6 to 10 tons (21.1 to 35.0 kW), slightly higher still.

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Controls Water-Loop Temperature Control. In a WSHP system with an open-circuit cooling to wer and a plate-and-frame heat-exchanger, the water temperature lea ving the plate-and-frame heat e xchanger is controlled between 60 and 90 °F (15.6 and 32.2°C) when the WSHP system is in operation. F or a typical WSHP system, a temperature sensor T1 may be located at the e xit of the plate-and-frame heat exchanger, as shown in Fig. 28.7. When the microprocessor -based DDC controller recei ves a signal from T1, it actuates the following: At 57°F (13.9°C) At 60 to 68°F (15.6 to 20.0°C) At 84°F (28.9°C) At 86°F (30.0°C) At 88°F (31.1°C) At 90°F (32.2°C) At 105°F

Alert by alarm. WSHP, cooling tower, water heater, and pumps all shut down. Start the water-loop water pumps. Energize the water heater. Start the cooling tower water circulating pumps. Open the damper of the cooling tower. Turn on the f rst-stage tower fans. Turn on the second-stage tower fans. Alert by alarm. System will shut down.

An outdoor air sensor T3 resets the w ater heater ener gizing temperature T2 in such a manner that when the outdoor temperature To drops from 60 to 0 °F (15.6 to 17.8°C), T2 increases linearly from 60 to 68°F (15.6 to 20.0°C). To maintain space air at 75 °F (23.9°C), 50 percent relati ve humidity, and a de w point of 55 °F (12.8°C) in summer , it is critical to maintain a minimum w ater-loop temperature Twl  57°F (13.9°C). If Twl  57°F (13.9°C), condensate may form on the outer surface of uninsulated pipes in the conditioned space and cause damage. The actuating temperatures for turning of f the second-stage and f rst-stage tower fans, shutting off the cooling tower water pump, and closing the damper in the cooling tower as the water temperature decreases should be successi vely 2°F (1.1°C) lower to prevent short cycling [e.g., turn off the second-stage tower fans at 90°F (32.2°C), turn off the f rst-stage tower fans at 88°F (31.1°C)]. Capacity Control of Water Heater. If the water heater is a gas-f red hot water boiler or a hot water heat exchanger, a temperature sensor T2 senses the water temperature leaving the water heater. The heating capacity can be controlled by modulating the f ow rate of gas or hot w ater by means of a DDC controller. Because most hot w ater boilers contain a limited amount of w ater, if the temperature sensor is located at the e xit side of the boiler , the temperature response to modulation is f ast. Therefore, a broadband DDC controller should be used to prevent short cycling. If an electric boiler is used, step control in appropriate stages can be performed through a sensor located at the inlet to sense the temperature of entering water. Step control often causes temperature f uctuation at the leaving side. Safety Control. In addition to the high and lo w operating limit control of the w ater-loop temperature, a f ow switch or pressure differential sensor should be installed across the water-loop pump. In case of pump f ailure, the entire system, including the water-source heat pumps, cooling tower, and circulating pump, or water heater shuts down. After a delay of 10 to 15 s, the standby pump is energized. At the same time, an alarm sounds with an indicating light. When water f ow is resumed, the cooling to wer and circulating pumps, or the w ater heater and w ater-source heat pumps, are restarted, with a delay between each stage to limit any sudden increase in starting current. SH__ ST__ LG__ DF

Water-Source Heat Pump Control. When the pushb utton is engaged, the electric circuit is energized and the outdoor air damper is opened. When the zone temperature rises abo ve the cooling mode set point, the temperature sensor T4 (Fig. 28.7) calls for cooling. The DDC controller then

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starts the f an. F or small console WSHPs used in the perimeter zone, a multispeed f an is usually used. Fan is at either the high or low position, which is set manually. After a 30- to 50-s delay, if the high-pressure and lo w-pressure safety control circuits are closed, the compressor is started. F or large vertical or horizontal WSHPs, the fan is belt-driven. Fan motor starts the f an after safety control circuits are closed. If there are multiple compressors, the leading compressor is started. WSHP is now in cooling mode operation. When the zone temperature drops belo w the cooling mode set point and is sensed by T4, the DDC controller then stops the compressor, or one of the compressors is deener gized. The f an should be operated continuously when the space is occupied, to pro vide outdoor ventilation air to the space. When the zone temperature drops below the heating mode set point and is sensed by the temperature sensor T5, the DDC controller starts the compressor . WSHP is operating at heating mode. When the zone temperature rises abo ve the heating mode set point, the DDC controller stops the compressor. There is always a deadband of 1 to 3 °F (0.6 to 1.7°C) if the cooling mode is automatically changed to heating mode or vice versa. In either heating or cooling mode operation, if the discharge pressure rises above the high-pressure limit, or the suction pressure drops belo w the lo w-pressure limit, the high- or lo w-pressure control opens the electric circuit, lights an alarm lamp, and stops the compressor. ASHRAE / IESNA Standard 90.1-1999 speci f es that each w ater-source heat pump shall ha ve a two-position valve for a water system having a total pump system power exceeding 10 hp (7.5 kW).

System Characteristics System characteristics of water-source heat pump systems are listed in Table 28.1.

Case Studies: Water-Source Heat Pump Systems Gottfried et al. (1997) introduced a w ater-source heat pump system with an open-circuit cooling tower in Ridgehaven Building, San Diego (RBSD), California, as discussed in Sec. 25.3. RBSD is an off ce building of 73,000 ft 2 (6784 m 2). In 1995, the new WSHP system installed in each of the two wings of the Ridgehaven Building included the following primary components: ●

●

●

●

72 high eff ciency WSHPs having an EER of 14.9 (4.33 COP) 100-ton (open-circuit) cooling tower 400,000 Btu / h (117 kW) gas-f red hot water boiler Condenser water isolation valve for each WSHP to cut off condenser water f ow when the WSHP is shut off

Ridgehaven Building provides 20 cfm (10 L /s) outdoor ventilation air for each occupant. Using specif ed high-eff ciency f lters for outdoor -air AHU, the RBSD tar get energy use is 9 kWh /ft2 yr (97 kWh / m2 yr). The f rst month’s operating data sho wed that the annual ener gy use for a WSHP system with an open-circuit cooling to wer prior to commissioning is approximately 11 kWh / ft2 yr (118 kWh / m2 yr). For an of f ce building in Stamford, Connecticut, Kush (1990) analyzed and reported on the actual performance of a closed-circuit w ater-source heat pump system that serv ed a total area of 72,000 ft 2 (6691 m 2) of of f ce space on three f oors. This system has 140 one-ton perim (console) WSHPs, twelve 10-ton core WSHPs, a closed-circuit e vaporaive cooler , and a 300-kW electric boiler. Stamford has a 97.5 percent winter design temperature of 9 °F (12.8°C) and 5617 heating degree-days. In summer, 2.5 percent design dry-bulb temperature is 84°F (28.9°C) with a mean coincident wet-bulb temperature of 71°F (21.7°C). According to the results measured between No vember 1987 and October 1988, the breakdown of annual HVAC&R energy use was as follows:

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CHAPTER TWENTY-EIGHT

Electric boiler Evaporative water cooler Water-loop pumps Perimeter WSHPs Core WHPs Fan Cooling Heating Others

32 percent 4 percent 11 percent 18 percent 8 percent 15 percent 2 percent 9 percent

The energy use for heating in core WSHPs is mainly caused by excessive outdoor ventilation air. The monthly HVAC&R energy use of this b uilding between November 1987 and October 1988 is shown in Fig. 28.8. The highest HVAC&R energy use, 110,050 kWh, occurred in January 1988. Of this, 66 percent was consumed by the electric boiler. The highest in summer months was that for August 1988 — 65,715 kWh. About one-half of August’s energy was used by core WSHPs. Yearround, the lo west ener gy use, 32,330 kWh, occurred in May 1988, when the core WSHPs consumed 40 percent of the energy input.

Design Considerations ●

Cooper (1994) recommended the following rules for occupant satisfaction: Provide core WSHPs from mechanical rooms. Provide console WSHP with individual control for control zones in perimeter zone. Provide adequate outdoor ventilation air. Give the tenant means to use the HVAC system in off hours. Decentralize to the points free of major shutdowns. Adjust tenant space conveniently.

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FIGURE 28.8 HVAC&R monthly ener gy use of a w ater-source heat pump system.

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AIR CONDITIONING SYSTEMS: SPACE CONDITIONING SYSTEMS ●

●

●

●

●

28.33

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If ceiling units are used, good design of ducting, return plenum, vibration hangers, and piping f ex is mandatory and requires careful consideration for noise attenuation. When console WSHPs are used, they should not oversized. An oversized WSHP creates more noise and is not energy-eff cient. A water f ow rate of 2.5 to 3.5 gpm (0.16 to 0.22 L / s), typically 3 gpm (0.19 L / s) per cooling ton, is appropriate for water-source heat pump systems. Outdoor ventilation air rates should be adequate and should not be greater than that speci f ed in ASHRAE Standard 62-1999, as listed in Table 5.9. Outdoor air damper should be closed when the conditioned space is not occupied. Fan power in water-source heat pumps has a certain in f uence on energy use. Both supply volume f ow rate and total pressure loss should be carefully calculated in order to select the proper unit. Variable-speed water-loop pumps and cooling to wer fans are often cost-ef fective in lar ge WSHP systems.

28.6 PANEL HEATING AND COOLING Panel heating and cooling in volves the control of the mean radiant temperature of the conditioned space, which is closely related to the thermal comfort of the occupants. Floor-panel heating systems have been used by man y residences in Europe since the 1980s, as discussed in Sec. 8.8. Ho wever, panel heating and cooling is expensive and has a very limited application in the United States. If the conditioned space has a lar ge glaze area, metal ceiling panel should be used. A dedicated v entilation system should be used to pro vide the ASHRAE Standard 62-1999 specif ed amount of outdoor ventilation air. To prevent any possibility of condensation on the panels, the air system must maintain a space dew point temperature and humidity levels lower than the design conditions all the time when cooling panel is used. Panel heating and cooling requires greater f eld e xperience in design and operation. Refer to ASHRAE Handbook 1996, HVAC Systems and Equipment, chapter 6 for detailed information.

REFERENCES Anantapantula, V. S., and H. J. Sauer, Heat Recovery and the Economizer for HVAC Systems, ASHRAE Journal, no. 11, 1994, pp. 48 – 53. ASHRAE, ASHRAE Handbook 1996, HVAC Systems and Equipment, ASHRAE Inc., Atlanta, GA, 1996. Carrier Corp., Products and Systems 1992 / 1993 Master Catalog, Carrier Corporation, Syracuse, NY. Clark, S. J., and B. Eversion, Non-HVAC Water Lines Provide Cost Effective Four-Pipe Fan-Coil Systems for Motel, ASHRAE Journal, no. 3, 1994, pp. 29 – 31. Cooper, W. S., Operative Experience with Water Loop Heat Pump Systems, ASHRAE Transactions, 1994, Part I, pp. 1569 – 1576. DOE / EIA, 1998 Nonresidential Buildings Energy Consumption Survey: Commercial Buildings Consumption and Expenditures 1995, DOE / EIA -0318 (95). Friberg, E. E., Case History — Low-Rise Off ce Building Using Water-Source Heat Pump, ASHRAE Transactions, 1988, Part I, pp. 1708 – 1725. Gottfried, D. A., E. A. Schoichet, and M. Hart, Green Building Environmental Control: A Case Study, HPAC, no. 2, 1997, pp. 71 – 78. Howell, R. H., and J. H. Zaidi, Analysis of Heat Recovery in Water-Loop Heat Pump Systems, ASHRAE Transactions, 1990, Part I, pp. 1039 – 1047. Hughes, P. J., Survey of Water Source Heat Pump System Conf gurations in Current Practice, ASHRAE Transactions, 1990, Part I, pp. 1021 – 1028.

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Kilkis, B. I., Radiant Ceiling Cooling with Solar Energy: Fundamentals, Modeling, and a Case Design, ASHRAE Transactions, 1993, Part II, pp. 521 – 533. Kochendorfer, C., Standardized Testing of Cooling Panels and Their Use in System Planing, ASHRAE Transactions, 1996, Part I, pp. 651 – 658. Kush, E. A., Detailed Field Study of a Water-Loop Heat Pump System, ASHRAE Transactions, 1990, Part I, pp. 1048 – 1063. Kush, E. A., and C. A. Brunner, Optimizing Water-Loop Heat Pump Design and Performance, ASHRAE Journal, no. 2, 1992, pp. 14 – 19. Lefebvre, R. R., New HVAC System Reduces Operating Costs, ASHRAE Journal, no. 4, 1993, pp. 20 – 23. Meckler, M., Integrating Water Source Heat Pumps with Thermal Storage, Heating / Piping / Air Conditioning, July 1988, pp. 49 – 64. Mulroy, W. J., The Effect of Short Cycling and Fan Delay on the Eff ciency of a Modif ed Residential Heat Pump, ASHRAE Transactions, 1986, Part I B, pp. 813 – 826. Pietch, J. A., Optimization of Loop Temperatures in Water-Loop Heat Pump Systems, ASHRAE Transactions, 1991, Part II, pp. 713 – 726. Simmonds, P., Practical Applications of Radiant Heating and Cooling to Maintain Comfort Conditions, ASHRAE Transactions, 1996, Part 1, pp. 659 – 666. The Singer Company, Electro Hydronic Systems, The Singer Co., Carteret, NJ, 1977. The Trane Company, Water Source Heat Pump System Design, The Trane Company, La Crosse, WI, 1981. The Trane Company, Fan-Coil Units, La Crosse, WI, 1990. Virgin, D. G., and W. B. Blanchard, Cary School — 25 Years of Successful Heat Pump / Heat Reclaim System Operation, ASHRAE Transactions, 1985, Part I A, pp. 40 – 45. Wang, S. K., Air Conditioning, vol. 4, Hong Kong Polytechinc, Hong Kong, 1987. Weinstein, A., L. D. Eisenhower, and N. S. Jones, Water-Source Heat Pump System for Mount Vernon Unitarian Church, ASHRAE Transactions, 1984, Part I B, pp. 304 – 312. Zaidi, J. H., and R. H. Howell, Energy Use and Heat Recovery in Water-Loop Heat Pumps, Variable-Air-Volume, Four-Pipe Fan-Coil, and Reheat HVAC Systems, Part 1 and Part 2, ASHRAE Transactions, 1993, Part II, pp. 13 – 39.

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CHAPTER 29

AIR CONDITIONING SYSTEMS: PACKAGED SYSTEMS AND DESICCANT-BASED SYSTEMS 29.1 PACKAGED SYSTEMS 29.2 Comparison between Packaged and Central Systems 29.2 Applications 29.3 Types of Packaged Systems 29.4 29.2 SINGLE-ZONE CONSTANT-VOLUME PACKAGED SYSTEMS 29.4 System Description 29.4 Air and Water Economizers 29.4 Supply Volume Flow Rate, and Cooling and Heating Coil Loads 29.4 Controls 29.5 Energy Use Intensities 29.5 System Characteristics 29.5 29.3 SINGLE-ZONE VAV PACKAGED SYSTEMS 29.7 System Description 29.7 System Calculations 29.7 Controls 29.7 System Characteristics 29.8 29.4 COMPLIANCE WITH STANDARD 90.1 – 1999 THROUGH SIMPLIFIED APPROACH OPTION FOR SMALL AND MEDIUM HVAC&R SYSTEMS 29.8 29.5 VAV COOLING PACKAGED SYSTEMS 29.9 System Description 29.9 Supply Volume Flow Rate and Coil Load 29.10 Pressure Characteristics and Duct Static Pressure Control 29.10 System Characteristics 29.12 29.6 VAV REHEAT PACKAGED SYSTEMS 29.12 System Description 29.12 Supply Volume Flow Rate and Coil Load 29.12 Night Setback and Morning Warm-up 29.14 Evenly Distributed Airflow at DX Coils 29.14 Discharge Air Temperature Control for Packaged Systems 29.15

Fan Modulation 29.16 Air-Cooled, Water-Cooled, and Evaporatively-Cooled Condensers 29.17 Controls and Sound Problems for Rooftop Packaged Units 29.17 Case Study: VAV Reheat Packaged System for Precision Manufacturing 29.17 System Characteristics 29.18 29.7 PERIMETER-HEATING VAV PACKAGED SYSTEMS 29.18 System Characteristics 29.18 29.8 FAN-POWERED VAV PACKAGED SYSTEMS 29.18 System Description 29.18 Supply Volume Flow Rate and Coil Load 29.19 Controls 29.20 Case Study: A Fan-Powered VAV Packaged System with Rooftop Packaged Unit 29.20 System Characteristics 29.22 29.9 DESICCANT-BASED AIR CONDITIONING SYSTEMS 29.22 Desiccant-Based Air Conditioning 29.22 Desiccant Dehumidification and Sensible Cooling 29.22 Desiccant-Based Air Conditioning Systems 29.24 Desiccants 29.24 Rotary Desiccant Dehumidifiers 29.27 29.10 CASE STUDY: A DESICCANT-BASED AIR CONDITIONING SYSTEM FOR A SUPERMARKET 29.27 Loads in Supermarkets 29.27 System Description 29.28 Space Conditioning Line 29.28 Operating Parameters in Rotary Desiccant Dehumidifier 29.29 Heat-Pipe Heat Exchanger 29.29 Mixing of Process Air and Recirculating Air 29.30 Indirect Evoporative Cooler or Refrigeration 29.30

29.1

29.2

CHAPTER TWENTY-NINE

Gas Heater 29.30 Operating Parameters of the Desiccant-Based Air Conditioning Cycle 29.30 Part-Load Operation and Controls 29.31 29.11 CASE STUDY: A DESICCANT-BASED AIR CONDITIONING SYSTEM FOR RETAIL STORES 29.31 System Description 29.31 Operating Characteristics 29.31 Performance 29.32 29.12 CASE STUDY: A DESICCANT-BASED AIR CONDITIONING SYSTEM FOR OPERATING ROOMS 29.32

Indoor Environment of Operating Rooms 29.32 System Description 29.33 29.13 APPLICATIONS OF DESICCANT-BASED AIR CONDITIONING SYSTEMS 29.34 Comparison between Conventional Vapor Compression Refrigeration System and Desiccant-Based Air Conditioning System 29.34 Conditions to Apply Desiccant-Based Air Conditioning Systems 29.34 System Characteristics 29.35 REFERENCES 29.35

29.1 PACKAGED SYSTEMS Comparison between Packaged and Central Systems The differences in construction and operational characteristics between packaged and central systems include these: ●

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●

●

A packaged system (PS) al ways uses a DX coil to cool the supply air directly in the PU, whereas a central system (CS) adopts chilled w ater as the cooling medium. Refrigerants cool the chilled water in the chiller , and the chilled w ater cools the supply air in the AHU. If both packaged and central systems are using w ater-cooled condensers, a central system will need a 3 to 7°F (1.7 to 3.9°C) lower evaporating temperature than a packaged system. In a packaged system, the size of the air system, refrigeration system, and air conditioning system that serves a specific area a typical floo , an area of se veral floors or the entire b uilding is the same. In a central system, the size of the conditioned area of an air and refrigeration system is usually different. In a packaged system gas-fired furnace electric heaters are often used to heat the air , and DX coils, often air-cooled, or sometimes e vaporative condensers are used. Man y packaged units are rooftop units and installed outdoors. Scroll and reciprocating compressors are usually used in packaged systems. In a central system, there is often a central plant where boilers and chillers are installed indoors to supply hot and chilled w ater to the AHUs and to cool and heat the supply air there. Centrifugal, screw, and reciprocating chillers and w ater-cooled condensers are used in central systems. AHUs are usually installed indoors in the fan rooms. In a packaged system, the controls of the heating and cooling system are often a part of the discharge air temperature control in the packaged unit. In a central system, there are separate w ater system controls, heating system controls, and refrigeration system controls in the central plant. In general, low- , medium- , and high-efficien y filters are used in packaged systems; and usuall , medium- and high-efficien y filters are used in central systems Packaged units used in packaged systems are f actory-fabricated and -assembled, whereas some components in an AHU may be custom-built in the field The modulation of cooling capacity in a packaged system with scroll and reciprocating compressors is achieved by cycling of the cylinders or step controls. The control actions of the modulation

AC SYSTEMS: PACKAGED AND DESICCANT-BASED SYSTEMS

29.3

of capacities in air, water, and refrigeration systems of a central system are often stepless, continuous modulation controls. Because of these differences, packaged systems and central systems are compared as follows: First, consider the IAQ of the conditioned space that the air conditioning systems serve: ●

●

●

Both the PS and the CS can provide the minimum ventilation air to dilute the air contaminants. Both the PS and CS can pro vide the humidity control and pre vent wet surfaces and mold gro wth in the ducts and conditioned space. Because of the lo wer fan total pressure pro vided by the small and medium-size packaged units, low- and medium-ef ficien y filters are often used in small and medium-size PUs instead o medium- and high-efficien y filters in AHUs

Second, the energy efficien y of the centrifugal and scre w chillers is higher than that of reciprocating and scroll compressors in packaged units, and the combined f an, motor, and drive efficien y is higher in AHUs in CS than in PUs in PS. Ho wever, a higher fan total pressure in AHUs, greater system heat gains, and an additional cooling medium require a greater e vaporating-condensing pressure lift than a direct-e xpansion (DX) system if PS and CS both are using either w ater-cooled or evaporatively cooled condensers. All these cause the energy consumption of packaged systems to be generally less than that of the central systems. According to the DOE /EIA 1995 Commercial Buildings Consumption and Expenditures, the annual energy use intensities of the PS and CS in 1995 in the United States are as shown below: EUI, kBtu / ft2  yr (kWh / m2  yr) Packaged systems Water source heat pumps (HPs), air source HPs, and other HPs VAV systems Single-zone, constant-volume

28.1 (8.2) 37.1 (10.9) 44.0 (12.9)

Central systems

44.7 (13.1)

Third, packaged systems need more maintenance than central systems. Fourth, rooftop packaged systems may be located directly abo ve the occupied area and create potential noise problems. Proper design and attenuation are important. Fifth, chillers and chilled w ater systems in a central system ha ve greater reliability than more scattered packaged units in packaged systems. The fl xibility of expansion and change of layout are similar for both packaged and central systems. Sixth, most of the packaged systems and the central systems need return air to mix with the outdoor ventilation air, and install mechanical e xhaust systems if required. Central systems are more easily coordinated with smoke-control systems in multistory buildings than packaged systems. Finally a packaged system has lower initial cost and often requires less space; factory-built units are easier and faster to install than central systems.

Applications Most current medium-size and lar ge packaged units are equipped with microprocessor -based DDC control systems, higher-efficien y scroll compressors, medium- and sometimes high-efficien y filters with many energy-efficient features such as VAV boxes, air economizer, evaporative condenser, fault detection and diagnostics. All this mak es the packaged system the most widely used system in small and medium-size air conditioning (HVAC&R) systems in both commercial b uildings and households

29.4

CHAPTER TWENTY-NINE

in the United States, as discussed in Sec. 1.5. In commercial b uildings in 1995, the packaged system had 48 percent of the floor area among the air conditioned uildings, and in households, 55 percent. The packaged unit is the primary equipment in a packaged system, as discussed in Sec. 16.5. In a packaged system, air is often directly heated in the gas furnace or electric heater and is cooled by DX coils. These are the primary dif ferences between an AHU and a PU, and between a central and a packaged system. On the other hand, because of the use of the microprocessor-based DDC system in the packaged system, today the dif ference in indoor en vironmental control between a custombuilt central system and a packaged system is less than ever before.

Types of Packaged Systems Packaged systems can be subdi vided according to their configuration and operating characteristic into the following air conditioning systems: ●

●

●

●

●

●

Single-zone constant-volume packaged system (SZCVPS) Single-zone VAV packaged system (SZVAVPS) VAV cooling packaged system (VAVCPS) VAV reheat packaged system (VAVRPS) Perimeter-heating VAV packaged system (PHVAVPS) Fan-powered VAV packaged system (FPVAVPS)

29.2 SINGLE-ZONE CONSTANT-VOLUME PACKAGED SYSTEMS System Description A single-zone, constant-volume packaged system is an air conditioning system that uses a packaged unit to supply and return a constant-v olume fl w rate of conditioned air to and from a single-zone conditioned space. It maintains a predetermined zone parameter at design and part-load conditions by controlling the capacity of the gas furnace or refrigeration compressors. Single-zone packaged systems are widely used in residences, indoor stadiums, arenas, and many industrial applications. A typical single-zone constant-v olume packaged system is simlar to that sho wn in Fig. 20.16 a. It consists of mainly an upfl w gas furnace and a DX refrigeration system with the following: ●

●

●

●

●

●

●

A supply fan, most often of forward centrifugal type. Centrifugal fans with airfoil blades are usually used in large packaged units. A gas furnace with induced comb ustion, primary and secondary heat e xchangers. An air-source heat pump or an electric heater may be used to provide winter heating instead of a gas furnace. Low- and medium-efficien y filters DX coil connected with an outdoor condensing unit. Supply and return ducts, diffusers, and return inlets. Space, functional, and safety controls. A heating element humidifier and an xhaust / relief / return fan (both are optional).

Supply Volume Flow Rate, and Cooling and Heating Coil Loads For a single-zone, constant-volume packaged system, the supply v olume fl w rate V˙s, in cfm [m3 /(60 s)], required to offset the space sensible cooling load at summer design conditions can be

AC SYSTEMS: PACKAGED AND DESICCANT-BASED SYSTEMS

29.5

calculated from Eq. (20.69), as discussed in Sec. 20.17, V˙s 

Q rs 60scpa(Tr  Ts)

(29.1)

The supply v olume f ow rate required to of fset the space heating load at winter design conditions can be calculated from Eq. (20.70) V˙s 

Q rh 60scpa(Ts  Tr)

(29.2)

Usually, the supply volume f ow rate required to of fset the summer design sensible cooling load is greater than the winter heating load. F or convenience, the larger V˙s is used for both summer and winter. The DX coil load required at summer design conditions can be calculated from Eq. (20.48b) Q cc  60V˙ss(h ae  h cc)

(29.3)

and the heating coil load required to of fset the heating load at winter design conditions can be calculated from Eq. (20.39) Q ch  60V˙sscpa (T1  T2)

(29.4)

Controls For a single-zone constant-volume system used in most residences in the United States, the primary control is the zone temperature control; and a dual thermostat with tw o set points, for separate cooling and heating mode operation during year -round zone temperature control, is used, as discussed in Sec. 20.19. For single-zone constant-v olume systems used in man y industrial and commercial b uildings where morning w arm-up heating mode should be automatically changed o ver to cooling mode when the space is occupied, a single-set-point heating and cooling mode automatic changeo ver thermostat should be used. There is often a deadband of 1 to 3°F (0.6 to 1.7°C) between the heating and cooling modes, or vice versa. In addition to the zone temperature control, there are speci f c controls such as ignition control, capacity conrols, head pressure controls, and safety controls. The burner is al ways energized f rst and then the supply f an. Cooling and heating capacity controls include on / off and tw o-speed; and several heating-stage step controls are often used.

Energy Use Intensities According to DOE /EIA 1998 Commercial Buildings Consumption and Expenditure 1995, the energy use intensity of the single-zone constant-v olume packaged systems (or residential central systems), in 1995 in the United States w as 44 kBtu / ft2 yr (139 kWh / m2 yr). This EUI was only smaller than that in central air conditioning systems using chilled w ater as the cooling medium.

System Characteristics System characteristics of a single-zone, constant-volume packaged system (SZCVPS) are listed in Table 29.1.

29.6

CHAPTER TWENTY-NINE

TABLE 29.1 System Characteristics of Packaged Systems SZCVPS Zone thermal and sound control Control zone Control methods Control modes Heating-cooling mode changeover Sound control Indoor air quality (IAQ) Minimum ventilation control Filters Humidity control Types of humidity control Air systems Types Supply fan Fan total pressure Combined fan, motor, and drive eff ciency Volume f ow control Relief / return fan total pressure Combined fan, motor, drive eff ciency Economizer Cooling system Refrigeration compressor Capacity control Refrigerants Evaporator Condenser Refrigerant f ow control Energy performance Heating system Types of heating system AFUE Maintenance Fault detection and diagnostics N/A means not available.

SZVAVPS

VAVCPS, VAVRPS, PHVAVPS

Single-zone Electric, two-stage thermostat On / off Manual, automatic 35 – 45 dBA

Single-zone DDC, electric

Multizone DDC

PI, PID, on / off Automatic NC 30 – 45

PI, PID, on / off Automatic NC 30 – 45

Constant-volume Low- or mediumeff ciency Optional Heating element, wetted element

DCV, MPC Low-, medium-, higheff ciency Optional Heating element, wetted element

DCV, MPC Low-, medium-, high-eff ciency Optional Heating element, wetted element

Constant-volume air mixing Forward-curved centrifugal 2.5 in. WC 40%

VAV, air mixing

VAV, air mixing

Centrifugal, forward, airfoil 4.0 in. WC 40%

Centrifugal, forward, airfoil 4.5 in. WC 45%

Inlet vanes, variablespeed drive (VSD) 0.4 – 1.0 in. WC 35%

Inlet vanes, (VSD)

Air, water

Air, water

Air, water

Scroll, reciprocating On / off cycling

Scroll, reciprocating On / off cycling, hotgas bypass HCFC-22, HFC-407C, HFC-410A DX coil Air-cooled, evaporatively cooled TXV, restrictor

Scroll, reciprocating Cycling, hotgas bypass HCFC-22, HFC-407C, HFC 410A DX coil Air-cooled, evaporatively cooled TXV, electric expansion valve 10 – 12 EER

Gas furnace, conventional, 78%, condensing 93%

Gas furnace, heat pump, electric heating Gas furnace, conventional 78%, condensing 93%

VAVCPS N/A; VAVRPS and PHVAVPS: hot water, electric heating Boiler, conventional 80%, condensing 93%

Packaged unit

Packaged unit

Packaged unit

Constant-volume N/A

HCFC-22, HFC-407C, HFC-410A DX coil Air-cooled Expansion valve (TXV), capillary tube  5 tons, 9.3 EER Gas furnace, heat pump

0.4 – 1.0 in. WC 45%

AC SYSTEMS: PACKAGED AND DESICCANT-BASED SYSTEMS

29.7

29.3 SINGLE-ZONE VAV PACKAGED SYSTEMS System Description In a single-zone VAV packaged system, zone temperature is maintained by a DDC controller through variation of the opening of the inlet guide v anes at the supply f an inlet or the supply f an speed through a variable-speed drive. Therefore, the supply volume f ow rate is varied to match the zone load v ariation. A typical single-zone VAV packaged system is similar to that sho wn in Fig. 21.1 except for the following: ●

●

●

The AHU is replaced by a packaged unit. The water-cooling coil is replaced by a DX coil. The heating coil is replaced by a gas furnace with primary and secondary heat exchangers.

An air economizer or a water economizer is often used to save energy. The cooling mode operation at design conditions and part-load operation and the heating mode operation at design and partload operation are similar to that discussed in Sec. 21.2.

System Calculations The supply volume f ow rate to offset the zone sensible load at cooling mode design conditions can be calculated by Eq. (21.1) as V˙s 

Q rs 60scpa(Tr  Ts )

(29.5)

The DX coil load required at cooling mode design conditions can be calculated by Eq. (21.2) as Q cc  60V˙ss(h rn  h cc)

(29.6)

The heating capacity of the gas furnace required to of fset the single-zone heating load during heating mode design conditions can be calculated by Eq. (21.3) as Q ch  60V˙sscpa(Thc  Ten)

(29.7)

Controls Zone temperature control and sequence of operations are similar to those discussed in Sec. 21.2 e xcept for the following: ●

●

●

Cooling capacity is varied by cycling multiple compressors, or cylinder unloaders, or hot gas pass at low part-load operation. Cooling capacity control is discussed in later sections. Heating capacity is varied by controlling the gas valve in on / off or two-stage control. Discharge air temperature control is still required for a better control quality and ener gy saving. However, for a small packaged unit with only on / off cooling capacity control, zone temperature control in cooling mode operation is actually the same whether the single-zone VAV packaged system is employed with or without a discharge air temperature control.

Minimum v entilation control, head pressure control, and safety controls are similar to those discussed in Chap. 23. Shire y (1995) recommended an impro ved fan cycling strategy based on the f eld operating experience in the Salvador Dali Museum, St. Petersburg, Florida. The museum areas (storage

29.8

CHAPTER TWENTY-NINE

area and lobby) are primarily under thermostat control. Prior to the impro vements, all indoor fans operated continuously, regardless of compressor operation, to maintain v entilation and air circulation for occupant comfort. Continuous f an operation resulted in lar ge humidity f uctuations and increased indoor humidity le vels at night when space load w as reduced. Continuous f an operation also consumed more energy and needed additional compressor operation to of fset fan heat. If there are wetted surf aces around the coil and condensate drain pan when the compressors are cycling off, moisture will be evaporated from these wetted surfaces, extracted by the indoor fan, and supplied to the conditioned space. An alternati ve impro ved control strate gy “automatic f an mode ” was adopted. The indoor f an energized only when the compressors and heaters were operating. When the thermostat and humidistat set points are met, the indoor fan stops and the moisture drains from the packaged unit. The indoor fan that serves the storage room in the Salvador Dali Museum was changed to automatic fan mode control. The new control strategy provided a better storage indoor en vironment with signif cant energy savings. For the lobby, daytime continuous fan operation for constant air circulation was used when the lobby was open to the public. Automatic fan mode was used at nighttime.

System Characteristics System characteristics of a single-zone VAV packaged system are listed in Table 29.1. In Table 29.1, DCV represents demand-controlled v entilation and MPC indicates mixing plenum pressure control.

29.4 COMPLIANCE WITH STANDARD 90.1 – 1999 THROUGH SIMPLIFIED APPROACH OPTION FOR SMALL AND MEDIUM HVAC&R SYSTEMS Single-zone, constant-volume and single-zone, variable-air-volume packaged systems are the tw o widely used air conditioning systems in residential, commercial, and industrial b uildings in the United States. F or air conditioning systems in b uildings two stories or less in height and with less than 25,000 ft 2 (11,800 m 2) gross f oor area shall be considered in compliance with the requirements of HVAC in ASHRAE / IESNA Standard 90.1-1999 by means of a simpli f ed approach if the following criteria are met: 1. The system serves a single zone. 2. A single or split-packaged unit is used which is either air-cooled or evaporatively cooled and meets minimum eff ciency requirements listed. 3. An air economizer and control shall be installed as required in Sec. 21.2 with either barometric or powered relief sized to prevent space overpressurization. 4. Heating shall be provided by packaged heat pumps, PTHPs, gas-f red warm air furnaces, electrical heaters, or hot water heating systems and boilers with all applicable eff ciency requirements met. 5. Outdoor air supplied by the system shall be equal to or less than 3000 cfm (1420 /L/s) and less than 70 percent of the supply air quantity at minimum outdoor air design conditions unless an energy recovery ventilation system is provided. 6. The system is controlled by a manual changeover or dual set point thermostat. 7. Controls shall be provided to prevent auxiliary electric heater operation when the heat pump alone can meet the required heating load. 8. Controls shall not permit reheat or simultaneous heating and cooling. 9. For spaces other than those requiring continuous operation having a heating or cooling capacity greater than 65,000 Btu/h (30.7 kW), and having a supply fan motor power greater than 3/4 hp (0.56 kW), a timeclock shall be provided a. To start and stop the system under different schedules, b. That is capable of retaining programming and time setting during a loss of power of at least 10 hours c. That has an accessible manual override allowing temporary operation up to two hours

AC SYSTEMS: PACKAGED AND DESICCANT-BASED SYSTEMS

10. 11. 12. 13. 14.

29.9

d. That is capable of being set back down to 55°F (12.8°C) and up to 90°F (32.2°C) during off hours HVAC&R piping, ducts, and plenums shall be insulated and meet the requirements listed in Secs. 7.2, 17.2, and 17.4. Duct systems shall be air balanced to within 10 percent of design airf ow rates. When separate heating and cooling equipment serve the same control zone, thermostats shall be interlocked to prevent simultaneous heating and cooling Exhausts with a volume f ow rate over 300 cfm (140 L/s) that do not operate continuously shall be equipped with gravity or motorized dampers that will automatically shut when the exhaust systems are not in use. Air systems with a design volume f ow rate greater than 10,000 cfm (4720 L /s) shall have optimum start controls.

29.5 VAV COOLING PACKAGED SYSTEMS System Description A VAV cooling packaged system is a multizone air conditioning system that pro vides conditioned air without heating. It uses a packaged unit with DX cooling coils to condition the air and distrib ute it to v arious control zones through VAV boxes, ducts, distributing devices, and controls. The zone supply volume f ow rate is modulated by the damper in the VAV box to match the v ariation of the zone sensible load to maintain a preset zone temperature. A typical VAV cooling packaged system with its system pressure characteristics is sho wn in Fig. 29.1. It consists of a packaged unit, many

FIGURE 29.1 A VAV cooling packaged system.

29.10

CHAPTER TWENTY-NINE

VAV boxes, ducts, diffusers, and controls. The air conditioning c ycle of a VAV cooling packaged system is the same as the air conditioning c ycle in cooling mode operation for the interior zone in Fig. 21.5a. VAV cooling packaged systems are used for interior zones in b uildings where the heating load is negligible.

Supply Volume Flow Rate and Coil Load The supply volume f ow rate for a control zone in the interior zone Vsin, in cfm [m 3 /(60 s)], or the supply volume f ow rate of interior zone V˙si, in cfm [m 3 /(60 s)], can be calculated by Eq. (21.1) as V˙sin  V˙si 

Q rin 60scpa(Tr  Ts)

(29.8)

Q ri 60scpa(Tr  Ts)

where Qrin, Qri  sensible cooling load of control zone n and interior zone, Btu / h (W). The DX coil load Qcc, in Btu / h (W), of the packaged unit that serv es the interior zone at cooling mode design conditions can be calculated by Eq. (21.2) as Q cc  60V˙ss(h m  h cc)

(29.9)

Pressure Characteristics and Duct Static Pressure Control In a VAV cooling packaged system with a system total pressure loss of 4.75 in. WC (1188 P a), as shown in Fig. 29.1, the total pressure loss, in in. WC, between various sections of the system at peak supply volume f ow may have values as follows: Flow, cfm Return system PU Supply main duct VAV box Flexible duct and slot diffuser

Peak f ow

0.5 Peak f ow

0.40 2.25 1.15 0.60 0.40

0.10 0.56 0.29 0.90 0.10

The purpose of duct static pressure control is to achieve the following: ●

●

●

Limit the maximum static pressure in the supply main ducts and maintain a preset static pressure at the point where the static pressure sensor is located. Provide the required volume f ow rate of supply air for any branch takeoff at both design load and part-load conditions, and prevent starving of any VAV box. Minimize fan energy use at part load.

A starving VAV box is a VAV box with a supply v olume f ow rate lo wer than the amount required to offset the zone load. This is because the set point of the duct static pressure control is not properly set, the static pressure sensor is not properly located, or the supply duct and branch tak eoffs are not properly designed and sized. If the static pressure sensor is located at the end of the main supply duct just before the last branch tak eoff at point D, shown in Fig. 29.1, at peak supply

AC SYSTEMS: PACKAGED AND DESICCANT-BASED SYSTEMS

29.11

volume f ow rate when the total pressure loss of the VAV box, f exible ducts, and slot diffuser in a branch tak eoff p  1.0 in. WC (250 P a), the static pressure at point D is 1.0 in. WC (250 P a). Point C is located upstream of point D. At peak supply v olume f ow rate, the mean air v elocity inside the main supply duct at point B, denoted by vB , is 1500 fpm (7.5 m / s), at point C the mean velocity Vc is 1200 fpm (6 m / s), and at point D the mean velocity vD is 1000 fpm (5 m / s). The total pressure loss between points B and C along the main supply duct is 0.3 in. WC (75 P a), and between points C and D it is 0.25 in. WC (63 Pa). During part-load operation, when the v olume f ow rate has been reduced to 0.5 of peak supply volume f ow, the static pressure at point D is maintained at 1.0 in. WC (250 Pa). From Eq. (17.11), pt  ps  pv

(29.10)

pt1  pt2  pf

(29.11)

and from Eq. (17.12) If the duct static pressure sensor is relocated at point C, then the pressure characteristics in the supply main duct at points B, C, and D at peak and 0.5 of peak supply v olume f ow rate, in in. WG (Pa), can be calculated as follows:

Peak supply volume f ow rate Static pressure ps Velocity pressure pv Total pressure pt  ps  pv pt1  pt2  pf Total pressure loss pf 0.5 peak supply volume f ow rate Static pressure ps Velocity pressure pv Total pressure pt  ps  pv pt1  pt2  pf

Point D

Point C

Point B

0.78 0.062

1.00 0.09 1.09

1.25 0.14

0.084

0.25

0.944 0.015

1.00 0.022 1.022

0.959

Total pressure loss between C and D, pf

0.30

0.063

1.39

1.062 0.035 1.097 0.075

From these analyses, the following can be noted: ●

●

●

●

At peak (design) supply v olume f ow rate, points upstream of point C where the duct static pressure sensor is relocated (such as point B), both total pressure pt,up and static pressure ps,up of the airstream inside the duct, are higher than duct total pressure and static pressure at point C, denoted by ptC and psC. Points downstream of point C (such as point D) pt,do are lower than ptC, and ps,do are lower than psC. At part-load operation, such as 0.5 of peak supply volume f ow rate, pt,up ptC, ps,up psC, pt,do  ptC, and ps,do  psC. But their differences are smaller than in design load. When duct static pressure sensor is located at point C, compared with the static pressure sensor located at point D, at part-load operation of 0.5 of peak supply v olume f ow rate, the duct static pressure at point D or at points do wnstream from point D, ps,do  1.0 in. WG. That is, the VAV box connected to points do wnstream from where the static pressure sensor is located may starv e. It is the total pressure of the airstream inside the supply main duct (duct total pressure) that of fsets the total pressure loss in the branch tak eoff. However, duct static pressure is steadier than the duct total pressure, and duct static pressure is usually sensed in a supply main duct for duct static pressure control.

As discussed in Sec. 23.8, the set point of the duct static pressure control is the duct static pressure plus the associated v elocity pressure, i.e., the total pressure required to o vercome the

29.12

CHAPTER TWENTY-NINE

maximum total pressure loss in an y of the branch tak eoffs connected to the supply main duct. And 0.1 in. WG (25 Pa) should be added as a safety factor. For a supply duct system with similar con f guration in branch tak e-offs, the static pressure sensor of a duct static pressure control should be located near the remote end of the main duct as well as at a location where steady static pressure can be properly measured, as shown in Fig. 29.1. If there are tw o or three supply main ducts, two or three static pressure sensors should be installed. Each should be located near the remote end of each main duct. A comparator is used so that the DDC controller can pick the lo west static pressure as the feedback v alue to modulate the v ariable-speed drive to provide a duct static pressure higher than the preset value in all main ducts.

System Characteristics System characteristics of a VAV cooling packaged system are listed in Table 29.1.

29.6 VAV REHEAT PACKAGED SYSTEMS System Description A VAV reheat packaged system is a multizone system that uses a packaged unit with DX coil and f lters to condition the air and that supplies the conditioned air to v arious control zones in the perimeter zone of a building through reheating VAV boxes, ducts, and diffusers, and to various control zones in the interior zones through VAV boxes, ducts, and diffusers. The zone supply v olume f ow rate of the cold supply air is modulated to match the v ariation of the zone loads during partload operation. Heating provided by the reheating coil is used to of fset the zone heating load in winter in the perimeter zone as well as to pre vent the zone temperature drops belo w a preset temperature when the zone supply v olume f ow rate has been reduced to the minimum setting during cooling mode part-load operation. A reheating coil is not required in the VAV boxes in the interior zone. As discussed in Sec. 21.3, for energy saving, the zone supply volume f ow rate must be reduced to a minimum setting, such as 30 percent of the peak supply v olume f ow rate during cooling mode part-load operation before the reheating coil is ener gized. Figure 29.2 sho ws a VAV reheat packaged system with a rooftop packaged unit. The air conditioning cycle for cooling mode operation is the same as that sho wn in Fig. 21.5 b. Usually, the same rooftop packaged unit is used to supply conditioned air to both perimeter and interior zones. Ho wever, in winter, if two packaged units are used to serve the perimeter and interior zone separately , the mixing air temperature for the perimeter zone can be raised to a higher value, to save energy. VAV reheat packaged systems are simple and ef fective. Ho wever, simultaneous cooling and heating processes should be minimized. VAV reheat packaged systems are used in man y commercial buildings.

Supply Volume Flow Rate and Coil Load In a VAV reheat packaged system, the zone peak supply v olume f ow rate V˙sn, in cfm [m 3 /(60 s)], can be calculated from Eq. (21.4) as V˙sn 

Q rsn 60scpa(Tr  Tsn)

(29.12)

AC SYSTEMS: PACKAGED AND DESICCANT-BASED SYSTEMS

29.13

FIGURE 29.2 A VAV reheat packaged system.

From Eq. (21.5), the block supply volume f ow rate for perimeter zone V˙sx, in cfm [m 3 /(60 s)], can be calculated as V˙sx 

Q rsx 60scpa(Tr  Tsx)

(29.13)

From Eq. (21.6), the block supply volume f ow rate for interior zone V˙si, in cfm [m 3 /(60 s)], can be calculated as V˙si 

Q rsi 60scpa(Tr  Tsi)

(29.14)

29.14

CHAPTER TWENTY-NINE

From Eq. (21.8), the DX coil load Qcc, in Btu / h (W), can be calculated as Q cc  60V˙ss(h m  h cc)

(29.15)

The reheating coil load depends primarily on the winter zone heating load. From Eq. (21.10), the reheating coil load for a control zone in the perimeter zone Qchxn, in Btu / h (W), can be calculated as Q chxn  Q rhxn  Q venxn  Q rhxn  60V˙sxnscpa(Trxn  Tenn)

(29.16)

Night Setback and Morning Warm-up In multizone VAV reheat rooftop packaged systems, the heating of the zone air during the occupied period in the perimeter zone in winter is often pro vided by zone electric reheating coils, or sometimes by water heating coils. The gas furnace, heat pump, or electric heater in the rooftop packaged unit is only used to maintain a night setback temperature and for morning warm-up purposes. The reasons to supply w arm air to maintain an indoor night setback temperature (such as 55 °F, or 12.8°C) are ●

●

●

To prevent freezing of w ater pipes and w ater surfaces in areas where the outdoor temperature at night is below 32°F (0°C) To provide an acceptable indoor temperature for emergency access To reduce the time required to w arm up to a required temperature, say, 68 or 70 °F (20.0 or 21.1°C), prior to an occupied period the next morning

During the night setback and morning warm-up periods, the following hold: 1. Outdoor dampers and exhaust dampers should be completely closed. 2. Recirculating dampers, inlet vanes, and inlet cones of the supply fan should be fully open; or the variable-speed drive should be running at full speed. 3. All VAV boxes or reheating boxes should be fully open. 4. Both the refrigeration compressor and the relief fan should be turned off. 5. When the zone temperature e xceeds a certain limit, both the supply f an and the furnace, heat pump, or heater in the packaged unit can be turned of f. The heating device in the packaged unit will be energized again when the zone temperature drops below a certain limit. 6. The termination of the night setback period is the be ginning of the w arm-up period. The warm air supply temperature from the rooftop packaged unit during the w arm-up period is generally between 100 and 120°F (37.8 and 48.9°C). Evenly Distributed Airflow at DX Coils In a VAV reheat packaged system, if the air f owing through the DX coil is not e venly distributed over the entire coil surf ace, then liquid slugging of the reciprocating compressor , hunting of the thermostatic e xpansion v alve, and a decrease in the DX coil capacity may all occur at the same time. Refrigerant enters the various refrigerant circuits of the DX coil and the e vaporator, typically as a mixture of 75 percent liquid and 25 percent v apor, after passing through the thermostatic e xpansion v alve and the distrib utor tubes. If some of the refrigerant circuits ha ve hea vy refrigeration loads and others ha ve only v ery light loads, the refrigerant in circuits with hea vy loads e xpands rapidly to v apor, resulting in a greater v apor velocity and greater pressure loss. The refrigerant in circuits with v ery low loads remains in a liquid state and f ows to the compressor in the form of

AC SYSTEMS: PACKAGED AND DESICCANT-BASED SYSTEMS

29.15

liquid slugging. Liquid slugging may cause reciprocating compressor f ailure. Therefore, an e ven airf ow for the DX coil and thorough upstream air mixing must be maintained for both full- and part-load operations, especially using reciprocating compressors. Liquid refrigerant lo wers the temperature of the sensing bulb of the thermostatic expansion valve and causes hunting.

Discharge Air Temperature Control for Packaged Systems For a VAV reheat packaged system using reciprocating rooftop unit with a cooling capacity greater than 20 tons (70 kW), its compressors are often controlled by using four -step capacity controls. Multiple compressors are usually equipped in medium-size and lar ge packaged units. Each compressor may ha ve two or more c ylinders, so that tw o can be loaded and tw o unloaded at the same time (see Fig. 11.23). Air-Side Economizer Mode. For a typical DDC controller -activated dischar ge air temperature control, as sho wn in Fig. 5.6, if the set point of the dischar ge temperature Tdis is set at 53 °F (11.7°C), which gives a supply temperature at the slot dif fuser of Ts  55°F (12.8°C), there will be a control band of 4 °F (2.2°C). When a VAV reheat packaged system operates in an air economizer mode and 51  Tdis  55°F (10.6  Tdis  12.8°C), the system will f oat within the control band by mixing outdoor air with recirculating air. Initiation of Cooling Stages. If Tdis 53°F (11.7°C) and due to the air economizer alone can no longer balance the sensible coil load, then Tdis f oats to the upper limit of the control band. Once it reaches point 1, 55°F (12.8°C) in Fig. 5.6, the f rst-stage cooling is ener gized and Tdis drops below 55°F (12.8°C). First-stage cooling is most lik ely provided by the cooling capacity of tw o cylinders in a compressor. It cycles on and off, and Tdis f oats within the control band, with proportional-integral (PI) control mode to maintain Tdis  53°F (11.7 °C) with the least de viation. When the f rststage cooling is turned of f during c ycling, it needs a time delay of at least 4 min before it can be turned on again, in order to prevent hunting and possible damage to the compressor motor . If, after the f rst-stage cooling has been energized, the DX coil capacity still cannot balance the sensible coil load, then Tdis continues to rise until it reaches point 2, which is 1°F (0.56°C) higher than the upper limit of the control band. If the time interv al between the time the f rst stage turns on and the instant when Tdis reaches point 2 is greater than 4 min, the f rst-stage cooling is then locked on and the second-stage cooling is ener gized to c ycle on and of f in an attempt to maintain Tdis at 53 °F (11.7 °C). As in the f rst stage, there must be a time delay of 4 min between the time the second-stage cooling turns of f and the monent it can be turned on again. When Tdis f oats within the control band, it may drop to point 3 because of the lo w sensible coil load. If the cooling capacity of ha ving the f rst-stage cooling locked on and the second-stage cooling c ycling still cannot of fset the sensible coil load, Tdis will rise until it reaches a v alue 1°F (0.56°C) higher than the upper limit of the control band, such as point 4. At point 4, the third-stage cooling is c ycling, and the f rst- and second-stage cooling is locked on. In this type of packaged unit, the greatest cooling capacity is pro vided when the fourth-stage cooling is cycling and the f rst-, second-, and third-stage cooling is locked on. As Tdis f oats within the control band, any deviation from the set point of 53 °F (11.7°C) is integrated over time as part of the proportional-integral control mode, in an attempt to reduce the deviation to zero, i.e., to maintain a near -constant temperature at the set point of 53 °F (11.7 °C), e.g., points 5 and 6 in Fig. 5.6. If Tdis is being controlled in air economizer and cooling mode with refrigeration, the greater the deviation from the set point, the shorter the time needed for correcti ve action in order to pro vide control stability. When a decrease in the coil load causes Tdis to drop to a value 1°F (0.56°C) lower than the lower limit of the control band (e.g., point 7), the currently cycled stage is lock ed off and the ne xt-lower cooling stage becomes the cycling stage.

29.16

CHAPTER TWENTY-NINE

Successive decreases in Tdis to 1°F (0.56°C) below the lo wer limit of the control band cause repeated locking off of cooling stages until the second stage is lock ed off and the f rst stage becomes the cycling stage. When the f rst stage is cycled off, the liquid line solenoid valve is deenergized, but the compressor still operates to pump do wn the refrigerant to the condenser . When the suction pressure in the DX coil becomes lo wer than the limit of lo w-pressure control, it opens the compressor circuit, and both the compressor and the condenser fans are stopped. A fully intertwined DX coil will adjust its activating refrigerant circuits accordingly during step control of the compressor capacity. At design load, the evaporating temperature inside the DX coil is usually between 40 and 45 °F (4.4 and 7.2 °C). Hot gas bypass is only used to pre vent the DX coil from frosting in case Tev falls below 32°F (0°C) because of a sudden load decrease. Reset. For a DDC-controlled VAV packaged system that uses a proportional-inte gral control mode, Tdis can be reset based on either the space air temperature f oating within the deadband or the outdoor temperature To. When Tdis is reset based on the space temperature, it is preferable to ha ve several space temperature sensors connected in series. Their average temperature Trm, in °F (°C), is used to reset Tdis. As Trm f oats within the deadband between 70 and 75 °F (21.1 and 23.9 °C), each 1°F (0.56°C) drop in Trm corresponds to a 1 to 2°F (0.56 to 1.1°C) increase in Tdis. During cooling mode, if Tdis is reset based on the outdoor temperature To, then Tdis will not be reset when To 70°F (21.1°C). After To falls below 70°F (21.1°C), for every 5°F (2.8°C) that To drops, Tdis will increase 1°F (0.56°C). For instance, at an outdoor temperature To  50°F (10.0°C), the set point of Tdis will be 57 °F (13.9°C). For larger packaged units with multiple scroll compressors, cooling capacity is often divided into four capacity steps such as when reciprocating compressors are used.

Fan Modulation Three types of supply f an modulation are widely used in lar ge VAV packaged systems with rooftop packaged units to maintain a supply duct static pressure near the most remote branch tak eoff in the supply main duct: (1) a forw ard-curved centrifugal f an with inlet v anes, (2) an airfoil centrifugal fan with a v ariable-speed drive, and (3) an airfoil centrifugal f an with inlet cone modulation. Because of the smaller f an inlet in an airfoil centrifugal f an, it is not recommended that inlet v anes be used, in order to prevent an extremely high inlet velocity, such as the actual maximum inlet velocity which exceeds 5000 fpm (25 m / s). Both forw ard-curved centrifugal f ans with inlet v ane modulation and airfoil centrifugal f ans with inlet cone modulation have a lower cost than airfoil fans that use a variable-speed drive. However, variable-speed drive f an modulation is more ener gy-eff cient at part-load operations. A cost analysis is always helpful to determine whether ha ving a variable-speed drive or having inlet vanes is optimum. In a VAV packaged system using inlet v anes for its forw ard-curved centrifugal f an modulation, consider the following: ●

●

●

●

Because all inlet v anes are subject to a certain amount of air leakage at the completely closed position, the fan volume f ow rate will still be about 10 to 30 percent of the design f ow when the inlet vanes are completely closed. A modulation range from 100 to about 30 percent is usually suff cient for a VAV packaged system. A forward-curved centrifugal f an has a smaller f an surge area at a lo wer fan total pressure and f ow rate than an airfoil centrifugal fan. Fan modulation resulting in an e xternal total pressure drop greater than 4 in. WC (100 P ag) in a VAV packaged unit may cause damage to the unit. A discharge damper should not be installed because closing the VAV boxes can provide the same shutoff function. A discharge damper may cause e xcessive noise due to o verpressurization of the VAV boxes. It may also overpressurize the ducts.

AC SYSTEMS: PACKAGED AND DESICCANT-BASED SYSTEMS

29.17

A VAV packaged system is shut of f during nighttime unoccupied hours in summer and of f seasons. In locations where there is a cold winter , it operates to provide warm air and maintain a night setback indoor space temperature during unoccupied hours. Air-Cooled, Water-Cooled, and Evaporatively Cooled Condensers In rooftop packaged units using reciprocating compressors, the ener gy eff ciency ratio (EER) for air-cooled, water-cooled, and evaporatively cooled condensers at an outdoor dry-b ulb temperature of 95°F (35°C) and a wet-bulb temperature of 75°F (23.9°C) is as follows: Air-cooled condenser Water-cooled condenser Evaporative-cooled condenser

9 – 12 11 – 14 14 – 16

The EER for VAV packaged units using scroll compressors is higher than that for those using reciprocating compressors. Air-cooled and evaporatively cooled condensers often ha ve a higher installation cost than water-cooled condensers and cooling towers. Although water-cooled and evaporatively cooled condensers are considerably more ener gy-eff cient at an outdoor condition of 95 °F (35 °C) dry-b ulb and 75 °F (23.9 °C) wet-b ulb temperatures than air -cooled condensers, in man y locations the annual a verage dry-b ulb temperature is rather low. Shaffer (1987) used a manuf acturer’s computer program to compare the annual ener gy use of air-cooled and w ater-cooled reciprocating chillers for multistory b uildings using VAV reheat systems in Boston and San Die go. The annual ener gy use of a w ater-cooled condenser and cooling tower is about 10 percent higher than that of an air -cooled condenser, mainly due to the ener gy use of condenser pumps. Comparing the energy use and costs of air-cooled, water-cooled, and evaporatively cooled condensers is recommended in order to make an energy-eff cient and cost-effective selection. Controls and Sound Problems for Rooftop Packaged Units Microprocessor-based speci f c controls, safety controls, and diagnostics for a 1997 manuf actured rooftop packaged unit with a cooling capacity between 20 to 130 tons (70 to 457 kW) are listed in in Sec. 16.5. Noise considerations for rooftop top packaged units are discussed in Sec. 19.9. Case Study: VAV Reheat Packaged System for Precision Manufacturing Desmone and Frank (1992) reported an addition of 10,000 ft 2 (930 m 2) for sophisticated precision manufacturing for Oberg Industries, a manufacturer of high-quality carbide and steel stamping dies in Freeport, Pennsylvania. The required HVAC&R criteria are as follows: ●

●

●

Temperature is maintained at a constant le vel of 68  1°F (20  0.56°C) not only horizontally but also vertically from the f oor to 13 ft (3.96 m) high. Humidity should be maintained at 50 percent. Airf ow noise had to be limited and drafts and turbulence should be tightly constrained.

The selected HVAC&R system is a VAV reheat packaged system with tw o rooftop packaged units that feed cold air year -round into a medium-pressure duct loop. Machine oil duct e xhaust, grinding duct exhaust, conditioned air return, medium-pressure supply duct, and the lo west VAV boxes and reheating hot water piping are arranged in layers to facilitate servicing. There are 19 pneumatically controlled VAV boxes located throughout the e xtended plant with a two-stage modulation in sequence. The modulation of the supply air v olume f ow by the damper in

29.18

CHAPTER TWENTY-NINE

the VAV box provides the f rst-stage control. If the zone cooling load drops below the minimum setting of air delivery by the VAV box, the pneumatic controller opens the hot w ater valve, forcing the hot water to enter the reheating coil, and modulates the hot w ater f ow to maintain the required e xact zone temperature. The air economizer cycle is adopted to sa ve energy. The HVAC&R system is designed with a 75 percent backup capacity for reliability.

System Characteristics System characteristics of VAV reheat packaged systems are listed in Table 29.1.

29.7 PERIMETER-HEATING VAV PACKAGED SYSTEMS A perimeter -heating VAV packaged system is a VAV multizone system that uses packaged units with DX coils to condition the air , During cooling mode operation, the conditioned air is supplied to the control zones in the perimeter and to interior zones through VAV boxes, ducts, and diffusers. In heating mode operation, heating is provided to the perimeter zone by a perimeter -heating system using either hot water f nned-tube heaters or electric baseboard heaters to of fset zone heating loads. Ventilation is provided by the cold ventilation air supplied from the VAV boxes at 30 percent of the peak supply volume f ow rate minimum settting. Perimeter VAV packaged systems are suitable for multizone commercial b uildings in locations with long, cold winters. Similar to that in Sec. 29.5, for VAV reheat packaged systems, the supply v olume f ow rate for control zones in the perimeter and in interior zones V˙sn, in cfm [m 3 /(60 s)], can be calculated from Eqs. (29.12). The supply volume f ow rates for perimeter or interior zone Vsx or Vsi , in cfm [m / (60 s)], can be calculated from Eq. (29.13) or (29.14). The heating load for perimeter zone Qchx, in Btu / h (W), can be calculated from Eq. (29.16).

System Characteristics System characteristics of a perimeter-heating VAV packaged system are listed in Table 29.1.

29.8 FAN-POWERED VAV PACKAGED SYSTEMS System Description A fan-powered VAV packaged system (FPVAVPS) is a multizone system using a packaged unit to condition the air and to distrib ute the conditioned air through f an-powered VAV boxes, ducts, diffusers, and controls in the perimeter zone for an air system with con ventional air distrib ution as shown in Fig. 21.13 except that a packaged unit is used to condition the air; or it distributes the conditioned air through the f an-powered VAV boxes, ducts, diffusers, and controls in either perimeter or interior zones for an air system with cold air distribution. After the single-blade volume control damper has been closed to a minimum setting, such as 30 percent, the function of a f an-powered box is, f rst, to extract recirculating w arm plenum air and mix it with the cold primary air and, second, to energize a reheating coil to maintain a preset zone temperature during part-load cooling mode operation. In winter heating mode, an electric heater in several capacity stages is controlled to maintain a preset zone temperature. There are tw o types of f an-powered boxes: parallel fan-powered boxes and series f an-powered boxes. P arallel f an-powered box es sa ve f an ener gy more than series f an-powered box es and are

AC SYSTEMS: PACKAGED AND DESICCANT-BASED SYSTEMS

29.19

more widely used. F an-powered VAV packaged systems with parallel f an-powered boxes using an electric heater or hot water heating coils have been used in many commercial buildings.

Supply Volume Flow Rate and Coil Load As discussed in Sec. 21.5, to calculate the zone peak supply v olume f ow rate at summer design conditions V˙sn,d, in cfm [m 3 /(60 s)], one must consider whether it is a control zone in the perimeter zone or in the interior zone, and whether a conventional air distribution is used or a cold air distribution is used. ●

For conventional air distribution in the perimeter zone with a cooling supply temperature differential (Tr  Ts)  15 to 20°F (8.3 to 11.1 °C), from Eq. (21.18), the zone peak supply v olume f ow rate from the fan-powered box V˙sxn,d , in cfm [m3 /(60 s)], can be calculated as V˙sxn,d 

Q rsxn 60scpa(Tr  Ts)

(29.17)

For conventional air distribution in interior zones, the zone peak supply v olume f ow rate from the VAV boxes V˙sin, in cfm [m3 /(60 s)], can be calculated as V˙sin  ●

Q rsin 60scpa(Tr  Ts)

(29.18)

For cold air distribution in the perimeter zone with a cooling supply temperature dif ferential typically Tr  Tsc, dis  32°F (17.8°C), the peak zone supply v olume f ow rate of the cold primary air from the fan-powered box V˙sxn,c, in cfm [m3 / (60 s)], can be calculated as V˙sxn,c 

Q rsxn 60scpa(Tr  Tsc,dis)

(29.19)

As discussed in Sec. 21.5, the ratio between the recirculating plenum air from the f an-powered box and the peak zone supply v olume f ow rate V˙sxn,r /V˙sxn,c  0.67. Also, V˙sxn,c  V˙sxn,r  V˙sxn. Therefore, V˙sxn,r  V˙sxn  V˙sxn,c

(29.20)

where V˙sxn, V˙sxn,r  peak zone supply volume f ow rate and peak zone supply volume f ow rate of recirculating plenum air of control zone xn in perimeter zone, cfm [m3 /(60 s)] ●

●

●

For cold air distrib ution in the perimeter zone during winter heating design conditions, the peak zone supply volume f ow rate V˙sxn should be calculated by Eq. (21.18) so that the supply temperature dif ferential will be equal to or less than 15 °F (8.3 °C), to pre vent e xcessive b uoyancy and stratif cation. For cold air distribution in the interior zone, each control zone can be either with or without a fanpowered box, as discussed in Sec. 18.5. Supply of low-temperature air with high induction nozzle diffusers is recommended. When f an-powered box es are used for cold air distrib ution in the interior zone, the supply v olume f ow rate of cold primary air V˙sin,c, in cfm [m 3 /(60 s)], can be similarly calculated by Eq. (29.19) e xcept that the zone sensible cooling load in the perimeter zone Qrsxn should be replaced by zone sensible cooling load in the interior zone Qrsin, both in Btu / h (W). As the ratio V˙sin,r / V˙sin,c is equal to 0.67, the zone peak supply v olume f ow rate of the recirculating air from the fan-powered box V˙sin,r , in cfm [m3 /(60 s)], is therefore determined. From Eq. (29.13), the supply v olume f ow rate for the perimeter zone at summer design conditions V˙sx can be calculated. From Eq. (29.14), the supply volume f ow rate for the interior zone at summer design conditions V˙si can be similarly calculated, both in cfm [m3 /(60 s)].

29.20

CHAPTER TWENTY-NINE ●

From Eq. (29.15), the DX coil load Qcc in the packaged unit of a fan-powered VAV packaged system can be calculated, and from Eq. (29.16) the reheating coil load Qchxn in each of the f an-powered boxes in the perimeter zone can be similarly calculated, both in Btu / h (W).

Controls A fan-powered VAV packaged system has tw o air mixings: outdoor air and recirculating air mixing in the mixing plenum of the packaged unit, and the cold primary air and recirculating plenum air mixing in the f an-powered box. Minimum outdoor v entilation air control is critical. Both demandcontrolled ventilation using CO 2 sensors and mixed plenum pressure control are discussed in Chap. 23 and should be carefully in vestigated during the peak supply v olume f ow rate as well as when the v olume f ow rates of the cold primary air in control zones ha ve been reduced to minimum settings. Zone controls and the sequence of operations of f an-powered boxes are discussed in Sec. 21.5. Microprocessor-based speci f c safety controls and diagnostics for a late-1990s manuf acturered packaged unit are similar to those listed in Sec. 16.5.

Case Study: A Fan-Powered VAV Packaged System with Rooftop Packaged Unit This case study is a retro f t project. In it, an HVAC&R system is used to serv e a 48,000-ft 2 (4461m2) medical of f ce building in Little Rock, Arkansas. The building was constructed in 1979. It is owned by a limited partnership of 16 physicians. The renovation of the HV AC&R system of this project won a 1992 ASHRAE Technology Award, second place, of existing commercial b uildings. As described by Tinsley et al. (1992), prior to the 1990 reno vation, the HVAC&R system was a constant-volume electric terminal reheat rooftop packaged system. There were six rooftop packaged units with a total capacity of 120 tons (422 kW) and 162 electric duct heaters with an approximate total capacity of 450 kW. For the reno vation, a fan-powered VAV packaged system with a rooftop packaged unit w as designed and installed. One rooftop packaged unit of 133 tons (468 kW) with 100 electric heated parallel fan-powered units was equipped to serve this medical off ce building. The renovation mainly includes the following: ●

●

●

●

●

●

●

The air economizer and the compressors are controlled in sequence by DDC discharge air temperature control in order to maintain a supply air temperature at 55 °F (12.8°C). When the outdoor air temperature To 60°F (15.6°C), all compressors are locked out. An energy-eff cient evaporatively cooled condenser is used. The energy use of the compressors, tower fans, and spray pump at full load and at an outdoor wet-b ulb temperature of 80 °F (26.7°C) is 0.79 kW/ton (4.45 COP). An airfoil centrifugal fan with inlet cone modulation is used because it uses less fan energy than a forward-curved centrifugal fan with inlet vanes. A relief fan is used instead of a return f an. This relief fan operates only when the system is in air economizer mode. A lower duct static pressure control set point of 0.75 in. WG (188 Pag) is used. The fan-powered VAV packaged system operates only when an y one of the 16 suites is occupied. When any suite is unoccupied, the primary air damper in the VAV box is completely closed. The fan and the electric heater in the parallel f an-powered box are controlled in sequence to maintain a zone temperature 10 °F (5.6 °C) lo wer than the zone temperature set point during unoccupied hours. All supply ducts have been sealed and insulated.

AC SYSTEMS: PACKAGED AND DESICCANT-BASED SYSTEMS

TABLE 29.2 System Characteristics of Fan-Powered VAV Packaged System and Desiccant-Based Air Conditioning Systems FPVAVPS Zone thermal and sound control Control zone Control methods Control modes Heating-cooling mode changeover Sound control Indoor air quality (IAQ) Minimum ventilation air control Filters Humidity control Air systems Types Supply fan Supply fan total pressure Combined fan, motor, drive eff ciency Volume f ow control Air economizer Fan-powered box fan Fan total pressure Combined fan, motor, drive eff ciency Relief / exhaust fan Fan total pressure Combined fan, motor, drive eff ciency Cooling systems Refrigeration compressor

Multizone DDC or thermostat; high-mediumlow fan speed On / off, PI, or PID Automatic 35 – 45 dBA

VAV air mixing Centrifugal, forwardcurved or airfoil 4.5 in. WC 40 – 45% Inlet vanes, inlet cone, variable-speed drive (VSD) Air, temperature, or enthalpy Centrifugal, forward-curved 0.5 in. WC 25% Axial or centrifugal 0.4 – 1.0 in. WC 25 – 35%

Constant-volume or VAV Centrifugal, forwardcurved or airfoil 4.5 – 6 in. WC 45 – 55% N/A

Scroll or reciprocating

Centrifugal, screw, scroll, or reciprocating Centrifugal HCFC-123, screw HCFC-22, scroll HFC-407C, HFC-410A Centrifugal 0.5 – 0.55 kW / ton, screw 0.7 – 0.75 kW / ton, scroll 10 – 12 EER DX coil, liquid cooler Air-cooled, water-cooled, or evaporatively cooled Water f ow, or multiple compressors Gas heater, electric heater, or hot-water coil Conventional 78%, condensing 93%

Energy performance

10 – 12 EER

Evaporator Condenser

DX coil Air-cooled, or evaporatively cooled On / off, multiple compressors, hot-gas bypass Gas furnace, hot-water coil, or electric heater Conventional 78%, condensing 93%

AFUE Maintenance Fault detection and diagnostics

On / off, PI, or PID Automatic, manual NC 30 – 45 Demand-controlled ventilation, or mixing plenum pressure control Medium- or high-eff ciency Heating element humidif er and desiccant dehumidif er

HCFC-22, HFC-407C, HFC-410A

Heating system

Single-zone or multizone Thermostat, humidistat, or DDC

Demand-controlled ventilation, or mixing plenum pressure control Medium- or high-eff ciency Optional

Refrigerants

Capacity control

DBACS

Packaged unit

N/A N/A Centrifugal 1– 1.5 in. WC 30 – 40%

29.21

29.22

CHAPTER TWENTY-NINE ●

The zone temperature can be maintained between 70 and 80 °F (21.1 and 26.7 °C) in an y control zone by the DDC system.

The annual energy use for the HVAC&R system in this b uilding before renovation was 169,150 Btu / ft2 yr (533 kWh / m2 yr), and the annual ener gy cost w as $2.92 / ft2  yr ($31.4 / m2 yr). After renovation, the energy use intensity had dropped to 55,890 Btu / ft2 yr (176 kWh / m2 yr) and the energy cost was $1.40 / ft2 yr ($15.1 / m2 yr), drops of 67 and 52 percent, respectively. System Characteristics System characteristics of fan-powered VAV packaged systems are listed in Table 29.2.

29.9 DESICCANT-BASED AIR CONDITIONING SYSTEMS Desiccant-Based Air Conditioning A desiccant-based air conditioning process is a combination of desiccant dehumidi f cation, evaporative cooling, supplementary compression refrigeration, and the regeneration or reactivation of the desiccant by means of w aste heat or gas heating. A desiccant-based air conditioning system is a hybrid system of desiccant dehumidi f cation, evaporative cooling, refrigeration, and re generation systems to cool and dehumidify the space air and maintain it at a required temperature and relati ve humidity with adequate outdoor ventilation air, at the same time improving the eff ciency of energy use. In 1983, ASHRAE awarded the Willis Carrier Prize to Nanc y Banks, a 23-year-old engineer who proposed a desiccant dehumidi f cation process with evaporative cooling and necessary supplementary refrigeration in commercial b uildings. It w as a ne w idea at that time that considerably improved the ener gy use of refrigeration and started a ne w era of commercial desiccant-based air conditioning. Desiccant Dehumidification and Sensible Cooling When moist air f ows over a bed of either solid or liquid desiccant, absorption occurs at the bed of sorbents. Absorption is a moisture soption process associated with physical or chemical changes, while adsorption is one without physical or chemical changes. F or liquid absorbents, such as liquid lithium chloride, the absorption of moisture is mainly caused by the v apor pressure dif ference between the moist air and the surface of the liquid. Solid adsorbents such as silica gels attract moisture because of the vapor pressure difference and the electric f eld at the desiccant’s surface. During the process of sorption, heat is released. This released heat, often called the heat of sorption, is the sum of the following: ●

●

Latent heat due to the condensation of the absorbed water vapor into liquid Heat of wetting when the surf ace of the solid sorbent is wetted by the attached w ater molecules, or the heat of solution when moisture is absorbed by the liquid sorbents

Heat of wetting v aries from a relati vely large value when new desiccant f rst absorbs moisture to a very low value when it is nearly saturated. During the desiccant dehumidi f cation process, the humidity ratio of the lea ving moist air is decreased. Although there is a heat loss through the outer casing of the dehumidi f er to the ambient air, it is rather small compared with the release of the heat of sorption and an y residual heat remaining after the regeneration process. Regeneration uses heat to dri ve off the accumulated moisture so that the sorbent can be reused. The temperature of the dehumidi f ed air is considerably increased because of the heat released

AC SYSTEMS: PACKAGED AND DESICCANT-BASED SYSTEMS

29.23

during sorption.The desiccant dehumidi f cation process line deviates from the thermodynamic wetbulb temperature line and is represented by line di-do on the psychrometric chart in Fig. 29.3. The decrease in humidity ratio during the desiccant dehumidif cation process is given by Difference in humidity ratio  wdi  wdo

(29.21)

where wdi, wdo  humidity ratio at dehumidif er inlet and outlet, lb / lb (kg / kg). The reduction of latent heat Ql, in Btu / h (W), can be calculated as Q l  60V˙oo(wdi  wdo)h fg

(29.22)

where vV˙o  olume f ow rate of air entering dehumidif er, cfm [m3 /(60 s)] o  air density of entering air, lb / ft3 (kg / m3) hfg  latent heat of vaporization, Btu / lb (J / kg) Desiccant dehumidif cation is a dehumidi f cation and heating process on the psychrometric chart, and it is illustrated by a straight line inclined downward, deviating slightly from the thermodynamic wet-bulb temperature line, as shown in Fig. 29.3. It has a f atter negative slope. The acute angle between the desiccant humidi f cation process and thermodynamic wet-b ulb temperature depends mainly on the desiccant material and the w ater-holding capacity of the desiccant. When the desiccant approaches saturation, angle is smaller than that when a ne w desiccant f rst absorbs moisture from the surrounding air . F or a rotary desiccant dehumidi f er impre gnated with lithium chloride, if manufacturer’s data are not available, angle can be estimated at 5°. The regeneration process remo ves the absorbed moisture from the desiccant. F or the re generation airstream, it is also a humidifying and cooling process. Molecular sie ves require a higher regeneration temperature than silica gels and lithium chloride.

FIGURE 29.3 Desiccant dehumidif cation and sensible cooling.

29.24

CHAPTER TWENTY-NINE

After desiccant dehumidifcation, the extremely high-temperature air is usually sensibly cooled in two stages in order to save energy. This process is represented by line do-sc1 on the psychrometric chart in Fig. 29.3. The f rst stage uses an indirect evaporative cooling process, or is cooled by an outdoor airstream in an air -to-air heat exchanger, or is cooled by the e vaporator end of a heat-pipe heat exchanger. In an indirect e vaporative cooling process, the hot air is sensibly cooled by another evaporatively cooled airstream without direct contact. The f rst-stage sensible cooling load Qsc1, in Btu / h (W), can be calculated as Q sc1  60V˙oocpa(Tdo  Tsc1)

(29.23)

where Tdo, Tsc1  temperature of dehumidi f ed air entering and lea ving indirect evaporative cooler, air-to-air heat exchanger, or heat pipe heat e xchanger, °F (°C). The end state of f rst-stage sensible cooling sc1 can be determined according to local outdoor dry- and wet-b ulb temperatures to use fully the indirect evaporative cooling to save refrigeration in second-stage sensible cooling. In the second stage, dehumidif ed air may be mix ed with the recirculated air and then enter a DX coil for further sensible cooling. The second-stage sensible cooling load Qsc2, in Btu / h (W), can be calculated as Q sc2  60V˙oocpa(Tsc1  Tsc2)

(29.24)

where Tsc2  temperature of dehumidif ed air leaving coil, °F (°C).

Desiccant-Based Air Conditioning Systems In an air conditioning system, the coil load can be di vided into sensible load and latent load. F or comfort systems in commercial b uildings, the latent load v aries from 5 to 35 percent of the total coil load. In applications such as supermarkets, it may exceed 50 percent of the total coil load. A desiccant-based air conditioning system is a system in which latent cooling is performed by desiccant dehumidif cation and sensible cooling by e vaporative cooling or refrigeration, as shown in Fig. 29.4. Thus, a considerable part of the e xpensive vapor compression refrigeration is replaced by inexpensive evaporative cooling. There are two airstreams in a desiccant-based air conditioning system: a process airstream and a regenerative airstream. Process air can be all outdoor air or a mixture of outdoor and recirculating air . Process air is also the conditioning air supplied directly to the conditioned space or enclosed manuf acturing process, or to the packaged unit or sometimes to the air-handling unit, or terminal for further treatment. A regenerative airstream is a high-temperature airstream used to reactivate the desiccant. A typical desiccant-based air conditioning system, or simply a desiccant-based system, consists of mainly the follo wing components: rotary desiccant dehumidi f er, heat-pipe heat e xchanger, indirect evaporative cooler, scroll or reciprocating v apor compression unit with DX coil or w atercooling coil, gas-f red heaters, fans, pumps, f lters, controls, outer casing, ducts, and piping. Recently, the heat-pipe heat e xchanger has been used instead of a rotary heat e xchanger in the earlier desiccant-based systems because a heat-pipe heat e xchanger has no cross-contamination or mo ving parts. Any of these components may be replaced by other components of similar function. In addition to these components, a desiccant-based system may need a return air system, smoke control systems, and mechanical exhaust systems.

Desiccants Either solid or liquid desiccant absorbs or releases moisture because of the difference in vapor pressure between the surface of the desiccants pdes and the surrounding air psur. In a desiccant dehumidif cation process, when pdes  psur, the desiccant absorbs moisture from the ambient air. In a regenerative process, pdes psur, so moisture is released from the desiccant to the surrounding air.

FIGURE 29.4 A desiccant-based air conditioning system for a supermark et: (a) schematic diagram; ( b) desiccantbased air conditioning cycle.

29.26

CHAPTER TWENTY-NINE

Desiccants can be classif ed as adsorbents, which absorbs moisture without accompanying physical and chemical changes, and absorbents, which absorb moisture accompanied by physical or chemical changes. Three kinds of desiccants are widely used in desiccant-based air conditioning systems: silica gel, lithium chloride, and molecular sieves. Silica Gels. These are solid desiccants and adsorbents. Structurally, they contain numerous pores and capillaries in which water is condensed and contained. Silica gel has a high capacity to absorb moisture and releases it at a higher temperature. They are lo w in cost and a vailable in sizes from 316-in. (4.8-mm) beads to powderlike grains. Recently, titanium silica gel is used for its stable property. Lithium Chloride (LiCl). This is an absorbent. It is in dry form when each LiCl molecule holds two water molecules. If each LiCl molecule holds more than two water molecules, it becomes a liquid and continues to absorb moisture. LiCl has a high capacity to absorb and to hold moisture. Lithium chloride is widely used in rotary wheel dehumidif ers. Molecular Sieves. These are actually synthetic zeolites, a solid desiccant and an adsorbent in the form of crystalline aluminosilicates produced by a thermal process. Molecular sie ves show physical stability and high moisture-releasing capacity at high re generating temperatures of 248 to 428 °F (120 to 220°C), and they are recommended in direct gas-f red applications.

FIGURE 29.5 Sorption isotherms of some desiccants. ( Reprinted with permission.)

Source: ASHRAE Handbook 1989,

Fundamentals.

AC SYSTEMS: PACKAGED AND DESICCANT-BASED SYSTEMS

29.27

As def ned in Sec. 3.4, a sorption isotherm is a constant-temperature curv e that indicates the relationship between the moisture content of the desiccant Xdes, or moisture absorbed as a percentage of its dry mass, and the relative humidity of the surrounding air sur. Figure 29.5 shows the sorption isotherms of silica gels, lithium chloride, and molecular sieves. LiCl has a much higher w ater-holding capacity than silica gels and molecular sieves.

Rotary Desiccant Dehumidifiers A rotary desiccant dehumidif er is a rotary wheel that removes moisture from the airstream f owing through it. A rotary desiccant dehumidi f er has similar construction and operating characteristics to the rotary total heat e xchanger, as discussed in Sec. 12.5. Both are rotary wheels. The main differences between a rotary desiccant dehumid f er and a rotary total heat e xchanger or rotary heat e xchanger are as follows: ●

●

●

The purpose of using a rotary desiccant dehumidi f er is to remo ve moisture from an airstream, while the purpose of using a rotary total heat e xchanger is to transfer heat ener gy from one airstream to another airstream. In a rotary desiccant dehumidi f er, there is a process airstream and a re generation airstream. The regeneration airstream has a very high temperature, whereas in a rotary total heat exchanger, there is an outdoor airstream and an exhaust airstream at indoor temperature. A rotary desiccant dehumidif er rotates at a speed of 6 and 10 r / h (revolutions per hour), whereas a rotary heat exchanger rotates at a speed of 10 to 25 r / min.

29.10 CASE STUDY: A DESICCANT-BASED AIR CONDITIONING SYSTEM FOR A SUPERMARKET Loads in Supermarkets A desiccant-based air conditioning system is operated at an open c ycle when the process air is entirely outdoor air. If the process air at the inlet to the rotary desiccant dehumidi f er is a mixture of recirculating air and outdoor air, the desiccant-based system is said to be operated at a closed c ycle. For supermarkets with many frozen food refrigerators, especially single-deck well types and single-deck island food refrigerators with considerable area exposed to indoor air, there are three types of loads: ●

●

●

Refrigeration load. The refrigeration load of the frozen food refrigerator depends mainly on the temperature of the foods in it, the design ambient conditions (the de w point temperature of the conditioned space air), and the frozen surf ace area exposed to the space air . The higher the space temperature and space relati ve humidity, and therefore the de w point, the greater the amount of moisture that may condense on the frozen surface, the larger the accompanying latent heat of condensation, and the greater the refrigeration load. Space cooling load . A considerable portion of the sensible space cooling load is remo ved by the cooler surfaces of the frozen food refrigerator and the e xf ltrated cold airstreams from the frozen refrigerators through their openings, which results in a space cooling load with a heavy latent load that may vary from 50 to 65 percent. Coil load. The coil load or refrigeration load of the desiccant-based air conditioning system in a packaged unit or sometimes an AHU to maintain a required space condition in the sales area is the sum of the space cooling load, outdoor ventilation air load, and system heat gain. During summer, the outdoor v entilation load is again mainly latent load, and it is af fected by the space temperature and relative humidity.

29.28

CHAPTER TWENTY-NINE

In many supermarkets, 50 percent of the electric ener gy used annually is consumed by frozen food refrigerators, and another 15 percent is used for the air conditioning system to maintain the required space condition.

System Description Consider a supermarket with an area of 30,000 ft 2 (2788 m2) and an outdoor ventilation air requirement of 3000 cfm (1416 L / s). The condition of outdoor air is dry-bulb temperature of 95°F, relative humidity of 40 percent, and humidity ratio of 0.0142 lb / lb (100 gr/ lb or 0.0142 kg / kg). The space cooling load is about 83,000 Btu / h (24,320 W). If a desiccant-based air conditioning system with an impregnated lithium chloride dehumidi f er is used instead of a con ventional vapor compression refrigeration system, it is possible to maintain an indoor space temperature of 75 °F (23.9°C) and a relative humidity of 45 percent (0.0085 lb / lb or kg / kg, 60 gs / lb). Meanwhile, the supply air v olume f ow rate can be reduced from the 1 cfm / ft2 (5 L / s m2) of a con ventional system to 0.5 cfm / ft2 (2.5 L / sm2) for a desiccant-based system. Thus, there is a signif cant reduction in refrigeration load of the frozen food refrigerators because of lo wer space relati ve humidity as well as the fan power consumption in the air conditioning system. Figure 29.4a is schematic diagram of a desiccant-based air conditioning system for a 30,000 ft 2 (2788 m 2) supermarket. The process airstream of the desiccant-based air conditioning c ycle is indicated on the psychrometric chart by the solid line in Fig. 29.4b. The state points of the process air at the exit of various components are as follows: Outdoor air Filter Rotary desiccant dehumidif er Heat-pipe heat exchanger Mixing with recirculating air Indirect evaporative cooler or refrigeration Supply air after supply fan Space air

o o dl pl m s s r

The state points at the e xits of the various components of the re generation airstream are as follows: Outdoor air Filter Heat-pipe heat exchanger Gas heater Rotary desiccant dehumidif er Exhaust fan to atmosphere

o o xl rg ro ro

Space Conditioning Line For a supermarket of 30,000 ft 2 (2788 m2), if 50 percent of the total space cooling load Qrc is latent load Qrl, then Qrl  0.5  83,000  41,500 Btu / h (12,160 W) If the supply air is at a temperature of 72.5 °F (22.5°C), a relative humidity of 48 percent, and a humidity ratio of 0.0079 lb / lb (55 gr / lb or 0.0079 kg / kg), from Eq. (20.72), the supply air v olume

AC SYSTEMS: PACKAGED AND DESICCANT-BASED SYSTEMS

29.29

f ow rate of the mixture of the process air and the recirculating air is V˙s  

Q rl 60s(wr  ws)h fg 32 41,500 60  0.075(0.0085  0.0079)(1061)

 14,487 cfm (6836 L /s)

It is approximately equal to 0.5  30,000  15,000 cfm (7078 L / s). Similarly, from Eq. (20.69), for a sensible cooling load of about 0.5  83,000  41,500 Btu / h, the supply volume f ow rate to maintain a space temperature of 75°F (23.9°C) is V˙s  

Q rs 60scpa(Tr  Ts) 41,500 60  0.075  (0.243)(75  72.5)

 15,180 cfm (7163 L/s)

The larger of these two calculated values should be the design supply volume f ow rate. Operating Parameters in Rotary Desiccant Dehumidifier Assuming that the humidity ratio of the process air at the e xit of the rotary desiccant dehumidi f er after dehumidif cation is 0.006 lb / lb (0.006 kg / kg or 42 gr / lb), the difference during the dehumidif cation process in the rotary dehumidif er is therefore wo  wdl  0.0142  0.006  0.0082 lb / lb (kg / kg or 57.4 gr / lb) For the process airstream, the outdoor air (point o), enters the rotary dehumidi f er at 95 °F (35°C) and a humidity ratio of 0.0142 lb / lb (kg / kg or 100 gr / lb). On the psychrometric chart, draw a line o-dl from point o with an acute angle of 5° between o-dl and the thermodynamic wet-bulb temperature line. This line intersects the 0.006 lb / lb (kg / kg or 57.4 gr / lb) humidity ratio line at point dl, which is the state point of process air leaving the rotary desiccant dehumidif er. At point dl, temperature Tdl  139°F (59.4°C). For process air at a volume f ow rate of 3000 cfm (1416 L / s), the dehumidifying capacity of the rotary dehumidif er is m˙deh  3000  0.075  0.0082  60  110.7 lb / h (50.2 kg / h) If the volume f ow rate of the re generation air is also 3000 cfm (1416 L / s) and the temperature of the regeneration air entering the rotary desiccant dehumidi f er is Trg  185°F (85.0°C) with a humidity ratio of 0.0142 lb / lb (kg / kg or 100 gr / lb), regeneration air leaves the rotary dehumidif er at Tro  130°F (54.4°C), and a humidity ratio of 0.023 lb / lb (kg / kg or 159 gr / lb). Heat-Pipe Heat Exchanger Assume that the effectiveness of the heat-pipe heat e xchanger is 0.65. If the dif ference between the specif c heat of the process air and re generation air is ignored, and if the mean density of the process air f owing through the heat-pipe heat e xchanger is hp  0.0685 lb / ft3 (1.096 kg / m3) and for the regeneration stream mean density rg  0.070 lb / ft3(1.12 kg / m3), then, from Eq. (15.35), 

p(Tdl  Tpl) rg(Tdl  To)

 0.65

29.30

CHAPTER TWENTY-NINE

The temperature of the process air after the heat-pipe heat exchanger is Tpl  139  0.65 

0.070(139  95)  109.8F(43.2C) 0.0685

The temperature of regeneration air leaving the heat-pipe heat exchanger Txl can then be evaluated as Txl  95 

0.0685(139  109.8)  123.5F (50.9C) 0.070

Mixing of Process Air and Recirculating Air Process air at a volume f ow rate of 3000 cfm (1416 L / s) is mixed with recirculating air with a volume f ow rate of 12,000 cfm (5663 L / s). The condition of the mixture point m can be determined from the psychrometric chart by dra wing a line that connects the space air and the process air after the heat-pipe heat exchanger, r-pl, so that 3000 rm   0.2 r-pl 15,000 From the psychrometric chart, the temperature of the mixture Tm is 81.7°F (27.6°C), and its humidity ratio wm is 0.0078 lb / lb (kg / kg or 55 gr / lb) with a dew point of 50°F (10°C). Indirect Evaporative Cooler or Refrigeration If the desiccant-based air conditioning system is installed in a location in which summer 1 percent cumulative frequency of occurrence of summer outdoor design wet-b ulb temperature is lo wer than 67°F (19.4°C), then an indirect e vaporative cooler is recommended. Otherwise, a scroll or reciprocating v apor compressor refrigeration system should be used. F or a supply v olume f ow rate of 15,180 cfm (7163 L / s), the sensible cooling capacity can be calculated as Qcs  60 V˙sscps(Tm  Ts)  60  15,180  0.075  0.243(81.7  72.5)  169,313 Btu / h (49,609 W) The refrigeration system is mainly used for sensible cooling. This is only possible when the e vaporating temperature Tev in the DX coil is higher than 50 °F (10°C). If Tev is lower than 50°F, a certain degree of dehumidif cation exists in sensible cooling process ms. The required refrigeration load is therefore greater than Qcs. Gas Heater The temperature of re generation air to reacti vate the desiccant LiCl is 185 °F (85°C). A gas heater is used to heat the air from Txl  123.5 to 185°F (50.9 to 85°C). The heating capacity of the gas heater is Qh  60 V˙ rgrgcpa(Trg  Txl)  60  3000  0.243  0.0685(185  123.5)  184,266 Btu / h (53,990 W) Operating Parameters of the Desiccant-Based Air Conditioning Cycle From the above calculations, the operating parameters of the desiccant-based air conditioning c ycle are as follows:

AC SYSTEMS: PACKAGED AND DESICCANT-BASED SYSTEMS

Point o Outdoor air dl After dehumidif er pl After heat-pipe heat exchanger m Mixture s Supply air r Space air Regeneration air xl After heat-pipe heat exchanger rg After gas heater ro After dehumidif er

Temperature, °F

29.31

Humidity ratio, lb / lb and kg / kg (gr / lb)

95 139 109.8 81.7 72.5 75, 45% RH

0.0142 (100) 0.006 (42) 0.006 (42) 0.0078 (55) 0.0078 (55) 0.0083 (57)

123.5 185 130

0.0142 (100) 0.0142 (100) 0.023 (161)

Part-Load Operation and Controls When the humidity ratio of the outdoor air drops or the space latent load f alls during part-load operation, there are three methods of maintaining the space humidity ratio wr and the space relati ve humidity r if the space temperature remains constant: ●

●

●

Modulation of the gas heating capacity and the temperature of the re generation air varies the dehumidifying capacity of the rotary desiccant dehumidif er. In bypass control, a portion of outdoor air bypasses the rotary desiccant dehumidi f er, so that the supply air has a higher humidity ratio. The rotational speed of the rotary desiccant dehumidi f er can be modulated by using a v ariablespeed drive. Lower rotating speed of the dehumidi f er means a smaller dehumidifying capacity of the rotary desiccant dehumidif er.

The space temperature is controlled by the v apor compression refrigeration system. If the sensible cooling load drops at part-load operation, a cylinder unloader or on /off control in multicompressor system can be used to maintain the space temperature within predetermined limits.

29.11 CASE STUDY: A DESICCANT-BASED AIR CONDITIONING SYSTEM FOR RETAIL STORES System Description Spears and Judge (1997) reported a f eld study of a large retail store, a new 188,000 ft 2 (17,465 m2) Wal-Mart supercenter in Norfolk, Nebraska. The outdoor v entilation is pro vided by an outdoor -air unit in a gas- f red desiccant-based air conditioning system, as shown in Fig. 29.6. The sales area is served by tw o desiccant-based outdoor air units (O AUs) and the grocery area by one O AU. These OAUs supply the required amount of outdoor air to maintain the preset space relati ve humidity and CO2 level. There are also rooftop packaged units (RPUs) to pro vide most of the sensible cooling capacity to offset sensible cooling load in the sales area. In the O AU, a DX coil also provides additional sensible cooling capacity to the space when needed. Each O AU supplies a constant 10,000 cfm (4719 L / s) to the sales area. Operating Characteristics During base ventilation operation, 60 percent of the supply air from the O AU, or 6000 cfm (2825 L / s) supplied from each of the O AUs is outdoor air . When the CO 2 sensor senses that the CO 2 level in the sales area e xceeds 1000 ppm, the OAU shifts to high v entilation-air operation; all 10,000 cfm (4719 L / s) supplied from the OAU is outdoor air until the CO2 level is brought back to an acceptable level.

CHAPTER TWENTY-NINE

EA

OA

Gas heater

OA

DX-coil

To store

Gas

Heat pipe

RA

Rotary desiccant dehumidifier

29.32

EA

OA Dampers

RA OA

Outdoor air

RA

Recirculating air

FIGURE 29.6 An outdoor air unit of a desiccant-based air conditioning system for a lar Nebraska.

EA

Exhaust air

ge retail store in Norfolk,

When the relative humidity sensor senses that the relati ve humidity of a representati ve zone exceeds a preset limit and calls for dehumidif cation, the entire 10,000 cfm (4719 L / s) of supply air in the OAU is dehumidif ed in the desiccant dehumidif er. When relative humidity drops to or below a preset limit and the dehumidif cation is not required, the entire 10,000 cfm (4719 L / s) bypasses the rotary desiccant wheel.

Performance At the Norfolk store, the desiccant-based outdoor v entilation air system pro vided excellent control of indoor relati ve humidity and comfort, generally maintaining relati ve humidity within 5 percent of the 45 percent set point. The space temperature w as maintained at 78 °F (25.6°C). When the two OAUs were operated in base v entilation mode, the CO 2 level in the store ne ver went abo ve 1000 ppm. During high v entilation air mode, the desiccant wheel remo ves 0.0065 lb / lb (kg / kg or 42 gr / lb) from the outdoor air . Despite the f act that the v entilation rate had been raised from 0.15 to 0.30 cfm / ft2 (2.73 to 5.46 m 3 /h m2) in the Norfolk store, the energy cost increased only 2.6 percent more.

29.12 CASE STUDY: A DESICCANT-BASED AIR CONDITIONING SYSTEM FOR OPERATING ROOMS Indoor Environment of Operating Rooms Because the surgical staff are gowning more heavily to avoid the hazard of infectious diseases such as AIDS and at the same time there is a higher lighting and electronic equipment load, today surgeons need 65°F (18.3°C), 45 percent relative humidity, a humidity ratio of 0.0058 lb / lb (kg / kg or

AC SYSTEMS: PACKAGED AND DESICCANT-BASED SYSTEMS

29.33

40 gr / lb), and a dew point of 43°F (6.7°C). Using conventional dew point, there is a 22°F (12.2°C) reheat and a dedicated refrigeration system.

System Description Busby (1996) reported a desiccant-based air conditioning system for a six-room operating suite at Straith Hospital, Southf eld, Michigan, as shown in Fig. 29.7. The process airstream is an outdoor airstream of 7600 cfm (3586 L / s) at an inlet with a temperature of 95 °F (35°C) with a humidity ratio of 0.0142 lb / lb (kg / kg or 100 gr / lb). It f ows through a pref lter of medium eff ciency and a water sensible cooling coil; the outdoor air is sensibly cooled to 80 °F (26.7 °C) and 0.0142 lb / lb (kg / kg or 100 gr / lb). Before entering the rotary desiccant dehumidi f er, the process air is di vided into two airstreams: a dehumidif ed airstream and a bypass airstream. At summer design conditions, the bypass air f ows through the bypass dampers at 400 cfm (189 L / s). The desiccant used in the rotary dehumidif er is titanium silica gel. After dehumidif cation in the rotary dehumidif er, the temperature of the process air is raised to 138 °F (58.9°C) and the humidity ratio reduced to 0.005 lb / lb (kg / kg or 36 gr / lb). Both the dehumidi f ed and bypass airstreams are e xtracted by the supply f an, at the outlet of the supply f an, the process air temperature is increased to 140 °F (60°C), and its humidity ratio is raised to 0.0057 lb / lb (kg / kg or 40 gr / lb). The process air then f ows through the evaporator end of the heat pipe, and its temperature drops to 95°F (35°C). When the process air f ows through a water-cooling coil, its temperature is sensibly

EA

OA

95 F 40 gr/lb

7200 cfm 53 F 40 gr/lb Heater Humidifier

140 F Supply 40 gr/lb fan 138 F 36 gr/lb

Rotary desiccant dehumidifier

Filter 95% eff.

Water cooling coil

Heat-pipe heat exchanger

Gas heater

Reactivated fan

Chilled water Filter 80 F coil 100 gr/lb

DDC controller

T H

Operating room 65 F f  50%, 46 gr/lb

FIGURE 29.7 A desiccant-based air conditioning system for a six-room operating suite at Straith Hospital, Southf eld, Michigan.

OA

95 F 7600 cfm 100 gr/lb

29.34

CHAPTER TWENTY-NINE

cooled further from 95 to 53 °F (35 to 11.7 °C) at a v olume f ow rate of 7200 cfm (3398 L / s). It is again f owing through a high-ef f ciency f lter of MERV15. At the supply inlet, the supply air temperature is 53°F (11.7°C) with a humidity ratio of 0.0057 lb / lb (kg / kg or 40 gr / lb). The process air is then supplied to the operating room through an electric heater and a humidi f er, so that the space temperature can be modulated by a DDC controller to maintain a preset indoor temperature and relative humidity according to sensed values. There is an e xhaust airstream of 7000 cfm (3303 L / s) to absorb the heat rejected from the condenser end of the heat pipe. There is also reactivating outdoor airstream of 250°F (121°C) heated by a gas heater. After it f ows through the rotary dehumidi f er, this reactivating airstream is e xhausted to outdoors.

29.13 APPLICATIONS OF DESICCANT-BASED AIR CONDITIONING SYSTEMS Comparision between Conventional Vapor Compression Refrigeration System and Desiccant-Based Air Conditioning System Compared with air conditioning systems using v apor compression refrigeration systems, the desiccant-based air conditioning systems have the following benef ts and savings: ●

●

●

Thermal energy costs are lo wer. Desiccant-based systems use lo wer-cost evaporative cooling to replace more e xpensive electric po wer cost for v apor compression refrigeration, and use lo wercost gas heating to dry the desiccant instead of more e xpensive electric power cost to remo ve latent load by vapor compression refrigeration. Desiccant air conditioning system results in drier ductw ork to prevent mold and bacterial growth. The all outdoor air process airstream pro vides a benef cial setup for demand-controlled minimum ventilation control. In a desiccant-based air conditioning system, it is more con venient to ha ve a built-in heat recovery if the supply and return ducts are located close together.

The disadvantages of a desiccant-based air conditioning system are that it is more e more complicated and needs more maintenance.

xpensive and

Conditions to Apply Desiccant-Based Air Conditioning Systems Desiccant-based air conditioning systems are widely used in supermark ets, outdoor air mak eup units, surgical operating rooms, and ice rinks. According to f eld experience, the following are the conditions under which a desiccant-based air conditioned system should be used: ●

●

●

●

The dew point temperature of the indoor space temperature is equal to 40°F (4.4°C) or lower. There is a need of dehumidif cation of large amount of outdoor air. There is a need of low space relative humidity at lower space temperature. There is a low sensible heat ratio SHR c in cooling and dehumidif cation process, such as SHR 0.7.

ASHRAE / IESNA Standard 90.1-1999 speci f es that where humidistatic controls, including dehumidif cation and humidi f cation, are pro vided, such controls shall pre vent reheating, mixing of heated and cold airstreams, or other simultaneous heating and cooling of the same airstream. Exceptions include: ●

Systems serving spaces where humidity levels are required to satisfy process requirements, such as computer rooms, museums, surgical suites, supermarkets, refrigerated warehouses, and ice arenas.

AC SYSTEMS: PACKAGED AND DESICCANT-BASED SYSTEMS ●

●

●

●

29.35

The system can reduce supply air v olume to 50 percent or less of the design air f ow rate or the minimal rate speci f ed in v entilation requirements in ASHRAE Standard 62 before simultaneous heating and cooling takes place. The individual fan-coil cooling (refrigeration) unit has a design cooling capacity of 40,000 Btu / h (11.7 kW) or less, or a design cooling capacity of 80,000 Btu/h (23.4 kW) or less and can unload to 50 percent capacity before simultaneous heating and cooling takes place. 75 percent or more of the ener gy of reheating or for pro viding warm air in the mixing process is provided from site recovered heat or from a site solar energy source. Where the heat added to the airstream is the result of the use of a desiccant system and 75 percent of the heat added by the desiccant system is remo ved by a heat e xchanger, either before or after the desiccant system, with heat recovery.

System Characteristics System characteristics of a desiccant-based air conditioning system (DB 29.2.

ACS) are listed in Table

REFERENCES Acker, W., Industrial Dehumidif cation: Water Vapor Load Calculations and System Descriptions, HPAC, no. 3, 1999, pp. 49 – 59. ASHRAE, ASHRAE Handbook 1996, HVAC Systems and Equipment, ASHRAE Inc., Atlanta, GA, 1996. ASHRAE, ASHRAE Handbook 1997, Fundamentals, Atlanta, GA, 1997. Banks, N. J., Desiccant Dehumidif ers in Ice Arena, ASHRAE Transactions, 1990, Part I, pp. 1269 – 1272. Burns, P. R., J. W. Mitchell, and W. A. Beckman, Hybrid Desiccant Cooling Systems in Supermarket Applications, ASHRAE Transactions, 1985, Part I B, pp. 457 – 468. Busby, R. L., Relieving the Headache of Humidity Control, Engineered Systems, no. 9, 1996, pp. 45 – 50. Carrier Corporation, Products and Systems 1992 / 1993 Master Catalog, Carrier Corporation, Syracuse, NY. Desmone, C. L., and P. L. Frank, Air Conditioning for Precision Manufacturing, Heating / Piping / Air Conditioning, no. 2, 1992, pp. 34 – 36. Haessig, D. L., A Solution for DX VAV Air Handlers, Heating / Piping / Air Conditioning, no. 5, 1995, pp. 83 – 86. Harriman, L. G., The Basics of Commercial Desiccant Systems, Heating / Piping / Air Conditioning, no. 7, 1994, pp. 77 – 85. Jones, R. S., Rooftop HVAC Equipment on Building Roofs, ASHRAE Transactions, 1991, Part I, pp. 442 – 444. Jordan, C. H., DX Refrigeration vs. Chilled Water, Heating /Piping /Air Conditioning, no. 10, 1991, pp. 94 – 98. Kegel, R. A., Unitary HVAC Enables School-to-Courthouse Conversion, HPAC, no. 3, 1996, pp. 55 – 60. Kovak, B., P. R. Heimann, and J. Hammel, The Sanitizing Effects of Desiccant-Based Cooling, ASHRAE Journal, no. 4, 1997, pp. 60 – 64. Manley, D. L., K. L. Bowlen, and B. M. Cohen, Evaluation of Gas-Fired Desiccant-Based Space Conditioning for Supermarkets, ASHRAE Transactions, 1985, Part I B, pp. 447 – 456. Marciniak, T. J., R. N. Koopman, and D. R. Kosar, Gas-Fired Desiccant Dehumidif cation System in a QuickService Restaurant, ASHRAE Transactions, 1991, Part I, pp. 657 – 666. McGahey, K., New Commercial Applications for Desiccant-Based Cooling, ASHRAE Journal, no. 7, 1998, pp. 41 – 45. Meckler, G., Use of Desiccant to Produce Cold Air in Gas-Energized Cold Air HVAC System, ASHRAE Transactions, 1990, Part I, pp. 1257 – 1261. Meckler, G., Comparative Energy Analysis of Gas-Energized Desiccant Cold-Air Unit, ASHRAE Transactions, 1991, Part I, pp. 637 – 640.

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CHAPTER TWENTY-NINE

Parsons, B. K., A. A. Pesaran, D. Bharathan, and B. Shelpuk, Improving Gas-Fired Heat Pump Capacity and Performance by Adding a Desiccant Dehumidif cation Subsystem, ASHRAE Transactions, 1989, Part I, pp. 835 – 844. Scof eld, M., and G. Fields, Joining VAV and Direct Refrigeration, Heating / Piping / Air Conditioning, no. 9, 1989, pp. 137 – 152. Shaffer, R., Comparison of Air and Water Cooled Reciprocating Chiller Systems, Heating /Piping /Air Conditioning, Aug. 1987, pp. 71 – 87. Shirey, D. B., Fan Cycling Strategies and Heat Pipe Heat Exchangers Provide Energy Eff cient Dehumidif cation, ASHRAE Journal, no. 3, 1995, pp. 31 – 33. Spears, J. W., and J. Judge, Gas-Fired Desiccant System for Retail Super Center, ASHRAE Journal, no. 10, 1997, pp. 65 – 69. Tinsley, W. E., B. Swindler, and D. R. Huggins, Rooftop HVAC System Offers Optimum Energy Eff ciency, ASHRAE Journal, no. 3, 1992, pp. 24 – 28. The Trane Company, Packaged Rooftop Air Conditioners, The Trane Company, Clarksville, TN, 1997.

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AIR CONDITIONING SYSTEMS: CENTRAL SYSTEMS AND CLEANROOM SYSTEMS

H1

30.1 CENTRAL SYSTEMS 30.2 30.2 FLOOR-BY-FLOOR CENTRAL SYSTEMS VERSUS CENTRAL SYSTEMS USING AIR SYSTEMS SERVING MANY FLOORS 30.2 Size of Air System 30.2 Separate Air Systems 30.2 Floor-by-Floor Air System versus Air System Serving Many Floors 30.3 30.3 CONTROLS AND OPERATING CHARACTERISTICS OF CENTRAL SYSTEMS 30.4 Controls at Part Load for Central Systems 30.4 Controls of Water, Heating, and Refrigeration Systems 30.4 Air and Water Temperature Differentials 30.5 Influence of Inlet Vanes on Small Centrifugal Fans 30.6 30.4 TYPES OF VAV CENTRAL SYSTEMS 30.7 30.5 SINGLE-ZONE VAV CENTRAL SYSTEMS 30.7 System Description 30.7 Supply Volume Flow Rate, Coil Load, and Zone Temperature Controls 30.7 System Characteristics 30.9 30.6 VAV COOLING CENTRAL SYSTEMS, VAV REHEAT CENTRAL SYSTEMS, AND PERIMETER-HEATING VAV CENTRAL SYSTEMS 30.9 System Description 30.9 Supply Volume Flow Rate and Coil Load 30.9 Zone Temperature Controls 30.10 System Characteristics 30.10

30.7 DUAL-DUCT VAV CENTRAL SYSTEMS 30.10 System Description 30.10 System Characteristics 30.11 30.8 FAN-POWERED VAV CENTRAL SYSTEMS 30.11 System Description 30.11 Zone Supply Volume Flow Rate and Coil Load 30.11 Case Study: A Fan-Powered VAV Central System 30.12 System Characteristics 30.13 30.9 CLEAN-ROOM SYSTEMS 30.14 System Description 30.14 Airflow 30.14 Pressurization 30.16 Temperature and Relative Humidities 30.16 System Characteristics 30.16 30.10 CASE STUDY: CLEAN-ROOM SYSTEMS FOR SEMICONDUCTOR INTEGRATED-CIRCUIT FABRICATION 30.16 Indoor Requirements 30.16 Energy Use of Components 30.17 System Description 30.17 Operating Characteristics 30.18 Summer Mode Operation 30.19 Part-Load Operation and Controls 30.19 Winter Mode Operation and Controls 30.20 System Pressure 30.21 Effect of Filter Final-Initial Pressure Drop Difference on System Performance 30.23 Design Considerations 30.24 REFERENCES 30.24

__SH __ST __LG 30.1

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A central air conditioning system consists of: a central plant in which a boiler and chillers are located, a water system to transport hot and chilled w ater from the central plant to the AHUs, and an air system often with AHUs to condition the mixture of outdoor and recirculating air and distrib ute the conditioned supply air to the conditioned space, as discussed in Sec. 1.4. Also, a return air system for energy saving, smoke control systems for multistory b uildings according to fire codes and mechanical exhaust systems may be required. A central air conditioning system, or simply a central system, is either a single-zone or a multizone air conditioning system. In a central system, air is heated or cooled by the hot or chilled w ater in coils in air -handling units (AHUs). Conditioned air is then distrib uted to v arious control zones through ducts, terminals, and diffusers. Hot w ater and chilled w ater are heated by the boilers and cooled by chillers in the central plant. In a central system, heating and cooling capacities can be accurately modulated by the w ater fl w; air contaminants can be effectively removed by high-efficien y, even HEPA and ULPA, filters Fan rooms and central plants can be located remotely from sensiti ve areas. Also there is signifi cantly less field and space H AC&R maintenance required; therefore, central systems are al ways used in the quietest space, the cleanest space, the most precision-oriented space, and the most demanding space.

30.2 FLOOR-BY-FLOOR CENTRAL SYSTEMS VERSUS CENTRAL SYSTEMS USING AIR SYSTEMS SERVING MANY FLOORS Size of Air System For a central system, the size of the air systems depends mainly on the following: ●

●

●

●

Building occupancy and layout The smaller the air system, the less the f an system total pressure with lo wer ener gy use, the smaller the diameter of the main duct, and the greater the number of the AHUs required, resulting in a higher initial cost Advantages to the operation and maintenance of the air system Advantages to the compartmentation of fire protectio

If the air-handling units are installed indoors, their optimum sizes are usually 15,000 to 25,000 cfm (7079 to 11,798 L / s). Above 25,000 cfm (11,798 L / s), the headroom available in the fan room to install an AHU and ducts is usually not adequate. AHUs having a supply volume fl w rate below 10,000 cfm (4719 L / s) are often more expensive per cfm (L / s) volume fl w than larger sizes. To increase the net rentable floor area weatherproof rooftop AHUs are often used instead of indoor units mounted in the f an room for buildings of only a few stories. In such instances, the size of rooftop AHUs is limited by tw o things: (1) the products currently a vailable, such as the f act that the largest rooftop AHU has a v olume fl w rate of 63,000 cfm (29,730 L / s), and (2) the f act that too large an air system always results in a higher system total pressure loss and, therefore, a greater energy cost.

Separate Air Systems SH__ ST__ LG__ DF

In a control zone with special requirements and a floor area greater than 1000 f 2 (472 m 2), it may be energy-efficient and cost-e fective to use a separate air system. Special requirements include the following:

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●

●

●

30.3

Special process temperature and humidity requirements Clean rooms or clean space Special health care requirements Special operating characteristics, such as after-hours operation

__RH TX

Floor-by-Floor Air System versus Air System Serving Many Floors In a floo -by-floor air system AHUs are installed at least one for each floor in an rooms or mechanical rooms. For an air system serving se veral or many floors conditioned air is supplied or returned through v ertical risers from and to the rooftop units, or from the AHUs installed in the f an rooms at the middle-level mechanical floor of a high-rise uilding, as shown in Figs. 1.1 and 1.2, or from the basement. A high-rise building is a multistory b uilding of four or more floors. A low-rise building has three floors or less Let us compare a floo -by-floor air system and an air system serving ma y floors both employing conventional air distribution. ●

●

●

●

●

●

●

●

●

●

Normally, a floo -by-floor air system needs a smaller duct system and less an energy and is more decentralized than an air system serving many floors which is more centralized. Because the peak loads for all the control zones in a system do not occur simultaneously , a centralized system always has the benefit of a smaller load d versity factor than a decentralized system. Therefore, its total capacity is smaller. An air system serving man y floors needs supply and return risers which reduces the amount of available rental space. The air v elocity in the return riser is al ways lower than that in the supply riser. Linford and Taylor (1989) recommended that rule-of-thumb estimate for the riser’ s area required of 1.3 ft2 (0.12 m2) per 1000 ft2 (472 m2) of conditioned area served. For a VAV system, it is far simpler to use a floo -by-floor air system as the suppl , and return air is more easily balanced in each floor when both supply and return olume fl w rates are varied at part load. Each floor in a flo -by-floor air system is a separate fire compartmen which avoids supply and return duct penetration between floors for fire safe . If local codes require smok e control and purging, the supply and return duct risers for a more centralized system can be easily used as part of the smoke control system during a building fire The fans or compressors in a more centralized system are often located f arther away from the occupied space than those in a floor -by-floor air system. The duct-borne noise is often the primary source to be attenuated by sound traps and the longer distances of inner -lined ductwork. A floo -by-floor air system has better redundan y, which confines a y malfunction and shutdown to the individual floo . A floo -by-floor air system has greater f xibility than a more centralized system for future de velopment, after-hours access for o vertime workers, easier and more accurate tenant metering, and staged completion and rental of building space. A more centralized system al ways has larger fans and compressors and, therefore, more efficien equipment. It is often able to use high-ef ficien y airfoil f ans and adjustable-frequenc y variablespeed drives. Operating and maintenance of a more centralized system are easier. The overall building cost of a more centralized system is often less than that for a floo -by-floo air system for lo w-rise buildings of tw o or three floors. or high-rise b uildings, a detailed lifecycle cost analysis should be undertaken to determine the optimum choice.

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30.3 CONTROLS AND OPERATING CHARACTERISTICS OF CENTRAL SYSTEMS Controls at Part Load for Central Systems For a multizone VAV central system, the zone thermal control actions in the air system, water system, and refrigeration system during summer cooling mode part-load operation are as follows: ●

●

●

●

●

●

In any of the control zones, when a drop in zone temperature Trn is detected by the temperature sensor, the DDC controller adjusts the position of the damper in the VAV box so that the v olume fl w rate of zone supply air is modulated to match the reduced sensible cooling load in that control zone, to maintain a preset zone temperature. As the dampers in the VAV boxes served by the same AHU are closed to smaller openings, the duct static pressure sensor detects the rise in static pressure. The DDC controller then closes the inlet vanes, or varies the speed of the supply f an, until the reduced supply volume fl w rate of the AHU is matched by the reduced sensible cooling load of that area served by that AHU. When the supply volume fl w rate of the AHU is reduced at part-load operation, the discharge air temperature Tdis tends to drop to a lo wer value. As this signal is sensed by the dischar ge temperature sensor, the DDC controller modulates the tw o-way valve of the w ater-cooling coil and reduces the fl w rate of the chilled water so as to maintain a nearly constant preset temperature, or a reset Tdis at part load. As the drop in the coil load, represented by the product of the chilled water fl w rate and the temperature difference between the chilled water entering and leaving the coil Twl  Twe, is sensed by a Btu meter, the DDC controller modulates the inlet v anes of the centrifugal compressor or v aries the compressor speed to reduce the refrigeration capacity to match the f all in the coil load at the cooling coils, to maintain a nearly constant or reset chilled w ater temperature lea ving the chiller. The capacities of the air , water, and refrigeration systems of a central system should be equal or nearly equal to each other so that an equilibrium between the space temperature, discharge air temperature, and the chilled w ater temperature leaving the chiller can be maintained at a specifi part load. Another important characteristic of the capacity control of central systems is that the control actions in all air , water, and refrigeration systems, except multiple constant-speed pumps used in building loops, are usually stepless, continuous modulation controls.

Controls in Water, Heating, and Refrigeration Systems For most VAV central systems discussed in this chapter , the controls for w ater, heating, and refrigeration systems are similar. ●

●

●

SH__ ST__ LG__ DF

Water system . The most widely used w ater system is the plant-b uilding loop system. There are four controls for the plant-b uilding loop system: (1) coil dischar ge air temperature control, (2) chilled water temperature leaving chiller control, (3) staging control, and (4) pressure differential control. Water system controls are discussed in Sec. 7.11. Heating system . Lo w-temperature hot w ater heating system with tw o-pipe indi vidual loop is widely used in central systems. Its controls include multiple-boilers staging control, hot w ater leaving boiler temperature reset, zone hot w ater temperature control, and safety controls. These are discussed in Secs. 8.3 and 8.7. Refrigeration system. Centrifugal and scre w chillers are widely used in central systems. Lea ving chilled water temperature control and reset, optimizing operation of multiple chillers, and safety controls are discussed in Secs. 11.11, 13.6, and 13.7.

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__RH TX

Air and Water Temperature Differentials Both air and water temperature differentials exist in a central air conditioning system. F or a central system in cooling mode operation, the supply air temperature dif ferential indicates the dif ference between the space air temperature Tr and the supply air temperature Ts. The discharge air temperature Tdis from the AHU after the dra w-through supply f an is about 3°F (1.7°C) lo wer than Ts because of the supply duct heat gain. The temperature of the conditioned air lea ving the cooling coil Tcc is about 5°F (2.8°C) lower than Ts as a result of the supply fan heat gain and duct heat gain. The implications of a greater supply air temperature dif ferential Tr  Ts, with respect to maintaining a specified space temperature and relat ve humidity, include the following: ●

●

●

●

●

A lower supply volume fl w rate V˙s and, therefore, less investment for ducts, VAV boxes, fl xible ducts, and diffusers A lower conditioned air off-coil temperature Tcc A greater risk of surface condensation due to a lower Ts A lower fan energy use because of the lower V˙s A higher compressor power input because of the lower Tcc.

The fan energy saving is often greater than the increase in compressor energy input when Tr  Ts is increased. In current practice, the supply air temperature differential is divided into two categories: Conventional Air Distribution. ●

●

●

●

This category has the following operating parameters:

A supply air temperature differential Tr  Ts of 15 to 24°F (8.3 to 13.3°C) A supply air temperature of 52 to 58°F (11.1 to 14.4°C), typically 55°F (12.8°C) A lowest temperature of chilled w ater leaving the e vaporator as lo w as 37°F (2.8°C) because of the improvements in freeze protection control in evaporators A space temperature of 75 to 78°F (23.9 to 25.6°C) and a space relati ve humidity between 45 and 50 percent

Neither ice storage systems nor glycol is used in this category. Cold Air Distribution. ●

●

●

●

This category includes the following operating features:

A supply air temperature differential Tr  Ts of 30 to 36°F (16.7 to 20.0°C) A supply air temperature of 42 to 47°F (5.6 to 8.3°C), typically 44°F (6.7°C) A supply of water to the water coil at a temperature of 34 to 38°F (1.1 to 3.3°C) A space temperature of 75 to 78°F (23.9 to 25.5°C) and a relati ve humidity between 35 and 45 percent

Cold air distribution is al ways used with an ice storage system as the brine melts the ice and is thus cooled to 34°F (1.1°C). Fan-powered units are often used to blend the cold primary air at 44°F (6.7°C) with the plenum air to produce a supply temperature of 55°F (12.8°C). Ethylene glycol or propylene glycol mix ed with w ater, or brine, is needed for freeze protection. Adequate insulation must be provided for ducts, terminals, and diffusers to prevent surface condensation. As described in Sec. 7.1, under currently accepted procedures, the chilled water temperature differential, i.e., the difference between the chilled water temperatures entering and leaving the evaporator Tee  Tel, is between 10 and 18°F (5.6 and 10.0°C) for w ater systems of the plant-b uilding loop. A lar ge w ater temperature dif ferential sa ves pump po wer and reduces the pipe size b ut requires a lo wer Tel. For a v ariable-fl w building water loop, although the chilled w ater temperature return from the coils Twl may be different from Tee at part load, Tee is nearly equal to Twl at design

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load. If glycol is not blended with the chilled water, the lowest chilled water temperature should not be lower than 37°F (2.8°C) to protect against freezing.

Influence of Inlet Vanes on Small Centrifugal Fans Inlet vanes mounted at the inlet of the centrifugal f an block a certain percentage of the air passage. For a small centrifugal f an, these vanes have a considerable ef fect on f an performance compared with similar types and sizes of fans without inlet vanes, even when the inlet vanes are fully open. ●

●

●

The ratio of the inlet diameter D1 to the impeller diameter D2, D1 /D2, is different for forw ardcurved and backward-curved centrifugal fans. For forward-curved fans, D1 /D2 varies from 0.8 to 0.9, and for backward fans, it varies from 0.65 to 0.8. The percentage of block ed air passage for small centrifugal f ans may be between 15 and 25 percent. Although the blocked area may be only 15 percent of the total air passage, the eddies and turb ulences after the inlet vanes, even if they are fully opened, result in a greater energy loss.

The follo wing is a comparison of f an performance between backw ard-curved and forw ardcurved centrifugal f ans with inlet v anes when the y are fully opened and f ans without inlet v anes. These data are taken from the manufacturer’s catalog of vertical modular AHUs (1990). Without inlet vanes Blade D2, in. V˙ , cfm vf, fpm ps, in WC pt, in. WC bhp, hp rpm pt,i / pt,o

BC 20 10,400 2063 3.75 4.01 10.93 1841 0.63

BC 22.25 13,000 2063 3.75 4.01 12.66 1645 0.86

With inlet vanes FC 15 6000 2143 3.75 4.04 5.79 1296 0.94

BC 20 10,400 2063 2.25 2.51 10.38 1839

BC 22.25 13,000 2063 3.17 3.43 14.45 1646

FC 15 6000 2143 3.50 3.79 5.95 1291

BC  backward-inclined or backward-curved centrifugal fan FC  forward-curved centrifugal fan V˙  volume f ow rate, in cfm (L / s) vf  air velocity at fan outlet, fpm (m / s) ps, pt  fan static pressure and fan total pressure, respectively, in. WC (Pa) bhp  brake horsepower, hp (kW) rpm  revolutions per minute pt,i, pt,o  fan total pressure with and without inlet vanes, respectively, in. WC (Pa)

If the ratio D1 /D2 for a backward-curved centrifugal fan of impeller diameter D2  20 in. (508 mm) is taken as 0.75 and the percentage of block ed area at the inlet is 20 percent, the velocity at the fan inlet when the inlet vanes are fully opened is vinlet 

10,400 V˙  Afree 0.8 [(20  0.75)/(12  2)]2  10,590 fpm (53 m / s)

SH__ ST__ LG__ DF

For such an e xtremely high velocity, the ratio of f an total pressure of this BC centrifugal f an of 20-in. (508-mm) impeller diameter when the inlet v anes are fully opened is pt,i /pt,o  0.63. More than one-third of the fan energy output is lost as a result of the installation of inlet vanes. Such a decrease in fan total pressure due to an e xtremely high velocity vinlet was verif ed in a f eld installation

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of a VAV system using inlet v ane modulation. A fan total pressure of less than 3 in. WC (750 P a) was available. After the inlet v anes had been dismantled, the fan total pressure increased to more than 4 in. WC (1000 Pa). For a BC fan with a 22.25-in.- (565-mm-) diameter impeller, the ratio pt,i /pt,o  0.86; and for an FC fan of 15-in.- (381-mm-) diameter impeller, the ratio pt,i /pt,o  0.94. During a cost comparison between inlet v anes and adjustable-frequenc y variable-speed drives, the total pressure loss due to the inlet v anes when they are fully opened must be tak en into consideration. Therefore, for centrifugal f ans in either AHUs or PUs, installation of inlet v anes is not recommended in backw ard-curved centrifugal f ans with impeller diameters smaller than 25 in. (635 mm).

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30.4 TYPES OF VAV CENTRAL SYSTEMS Similar to packaged systems, VAV central systems can be subdi vided into the follo wing air conditioning systems according to their conf guration and operating characteristics: ●

●

●

●

●

●

Single-zone VAV central system (SZVAVCS) VAV cooling central system (VAVCCS) VAV reheat central system (VAVRCS) Perimeter-heating VAV central system (PHVAVCS) Dual-duct VAV central system (DDVAVCS) Fan-powered VAV central system (FPVAVCS)

The dual-duct VAV central system is the only system with a dual-duct VAV air system. Re garding the operating characteristics of the air system, the difference between a VAV central system and a VAV packaged system is mainly due to the f act that water heating and cooling coils in an AHU are used to heat and to cool the supply air in the central system instead of gas or electric heater and DX coils in a PU to heat and cool the supply air in packaged systems. Their differences are discussed in detail in Sec. 29.1.

30.5 SINGLE-ZONE VAV CENTRAL SYSTEMS System Description A single-zone variable-air-volume central system is an air conditioning system that has a central plant and water systems to supply hot and chilled w ater to the w ater heating and cooling coils in AHUs to heat and cool the supply air , and distribution to the conditioned space through ducts and dif fusers, as shown in Fig. 21.1. A relief or return fan is employed to extract the recirculating air to the AHU, or it is exhausted outdoors during the air economizer c ycle. The zone air is maintained at a preset temperature or relative humidity by modulating the w ater f ow rate to the coils by means of a sensor and a DDC controller. Hot w ater is heated in the boiler , and the chilled w ater is cooled in the chiller in a central plant and transported to the AHU through hot and chilled w ater systems. Single-zone VAV systems are widely used in arenas, indoor stadiums, airport terminals, and many industrial applications.

Supply Volume Flow Rate, Coil Load, and Zone Temperature Controls The supply v olume f ow rate V˙s, in cfm [m / (60s)], of a single-zone VAV central system based on the zone sensible cooling load at summer design conditions can be calculated from Eq. (29.1); and

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TX

the supply volume f ow rate based on the zone heating load can be calculated from Eq. (29.2). Usually, the supply v olume f ow rate required to of fset the zone sensible cooling load is the greater of the two and is the adopted supply volume f ow rate for the single-zone VAV central system. The cooling coil load Qcc, in Btu / h (W), can be calculated from Eq. (29.3), and the heating coil load Qch, in Btu / h (W), can be calculated from Eq. (29.4). Zone temperature controls for a single-zone VAV central system are similar to those discussed in Sec. 21.2. TABLE 30.1 System Characteristics of VAV Central Systems SZVAVCS Single-zone DDC, electric PI, PID Automatic Optional NC 25 – 40

Multizone DDC PI, PID Automatic Optional NC 25 – 40

Multizone DDC PI, PID Automatic Optional NC 25 – 40

Indoor air quality Minimum ventilation control Filter eff ciency

DCV or MPC Medium or high

DCV or MPC Medium or high

DCV or MPC Medium or high

Air system Types

VAV, air mixing

VAV, air mixing

Forward, airfoil, centrifugal 4– 5 in. WC 50% Centrifugal or axial Relief 0.6 in. WC, return 0.5 – 1 in. WC 35% Fixed or differential drybulb or enthalpy

Forward, airfoil, centrifugal 4.5 – 5.5 in. WC 55% Centrifugal or axial Relief 0.6 in. WC, return 0.5 – 1 in. WC 35 – 40% Fixed or differential drybulb or enthalpy

VAV, air mixing, dual supply duct Single- or dual-fan, forward, airfoil, centrifugal 4.5 – 6 in. WC 55% Centrifugal or axial Relief 0.6 in. WC, return 0.5 – 1.5 in. WC 35 – 40% Fixed or differential dry-bulb or enthalpy

Centrifugal or screw HCFC-123, HCFC-22, HFC-134a Flooded — liquid cooler Water or evaporatively cooled 0.5 – 0.7 kW / ton

Centrifugal or screw HCFC-123, HCFC-22, HFC-134a Flooded — liquid cooler Water or evaporatively cooled 0.5 – 0.7 kW / ton

Centrifugal or screw HCFC-123, HCFC-22 HFC-134a Flooded — liquid cooler Water or evaporatively cooled 0.5 – 0.7 kW / ton

Hot water coil, electric heater

Hot water coil, electric heater

Conventional 78%, condensing 93%

Hot water coil, electric heater PHVAVCS lowtemperature hot water baseboard heater Conventional 78% condensing 93%

Conventional 78% condensing 93%

AHU, chiller

AHU, chiller

AHU, chiller

Supply fan total pressure Combined fan-motor-drive eff ciency Relief / return fan Relief / return fan total pressure Combined fan-motor-drive eff ciency Air economizer Cooling systems Refrigeration compressor Refrigerants Evaporator Condenser Energy performance, compressor Heating systems Type

Boiler AFUE

DF

DDVAVCS

Zone thermal and sound control Control zone Control methods Control modes Heating-cooling modes changeover Humidity control Sound control

Supply fan (SF)

SH__ ST__ LG__

VAVCCS / VAVRCS, PHVAVCS

Maintenance Fault detection and diagnostics

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System Characterisics System characteristics of a single-zone VAV central system (SZVAVCS) are listed in Table 30.1.

30.6 VAV COOLING CENTRAL SYSTEMS, VAV REHEAT CENTRAL SYSTEMS, AND PERIMETER-HEATING VAV CENTRAL SYSTEMS System Description A VAV cooling central system is a multizone system that has a central plant and chilled w ater systems to supply chilled w ater to the AHUs and w ater-cooling coils to cool and dehumidify the supply air in the AHUs and distrib ute it to v arious control zones through ducts, VAV boxes, diffusers, and controls in the interior zone of a b uilding. It is similar to that sho wn in Fig. 29.1 e xcept the packaged unit should be replaced by an AHU. A VAV reheat central system is a multizone system that has a central plant and w ater systems to supply hot w ater to the reheating VAV boxes, or an electric heater to heat the air in the reheating coil, and chilled w ater to the w ater-cooling coils to cool the supply air in AHUs; it distrib utes the conditioned air to various control zones in the perimeter zone of a building through ducts, reheating VAV boxes, diffusers, and controls, and to v arious control zones in the interior zone of a b uilding through ducts, VAV boxes, diffusers, and controls. A VAV reheat central system is similar to that shown in Fig. 29.2 except the rooftop packaged unit should be replaced by an AHU. A perimeter-heating VAV central system is a multizone system that has a central plant and w ater systems to supply hot water to the perimeter hot water heating system and chilled water to the cooling coils in the AHUs to condition the supply air and distrib ute it to v arious control zones in the perimeter and interior zones through ducts, VAV box es, diffusers, and controls, as shown in Fig. 21.6. In the perimeter zone, winter heating is pro vided by a lo w-temperature hot w ater heating system, such as basebord f nned-tube heaters. During winter heating, the VAV boxes supply cold primary air at minimum setting to various control zones in the perimeter zone for the required minimum outdoor ventilation air.

Supply Volume Flow Rate and Coil Load For VAV cooling central systems, VAV reheat central systems, and perimeter-heating VAV central systems, the supply volume f ow rate V˙sn, in cfm [m3 /(60 s)], of any control zone in the perimeter or interior zone at summer design conditions can be calculated from Eq. (29.12) as V˙sn 

Q rsn 60scpa(Tr  Tsn)

(30.1)

The cooling coil load of the AHU that has a supply v olume f ow rate of V˙s can be calculated from Eq. (29.15) as Q cc  V˙ss(h m  h cc)

(30.2)

In a VAV reheat central system, for any control zone in the perimeter zone, the reheating coil load of a reheating VAV box can be calculated from Eq. (29.16) as Qchn  Qrhxn  Qvenxn  Qrhxn  60V˙ sxnscpa(Trxn  Tenn)

(30.3)

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Zone Temperature Controls

Ch_30 Final Pass pg 30.10

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In a VAV cooling central system, as the temperature sensor detects that the zone temperature Trn is falling below a preset v alue (set point) in the interior zone during part-load operation, a DDC controller closes the damper opening in the VAV box until the minimum setting, such as 30 percent, is reached. When the VAV box damper is at minimum setting, if the zone temperature Trn drops 2 or 3°F (1.1 or 1.7 °C) lower than the set point, the DDC controller resets the dischar ge air temperature of the AHU Tdis by 2 or 3°F (1.1 to 1.7°C) higher in order to maintain a higher zone temperature Trn. In a VAV reheat central system, VAV boxes are installed in each of the control zones in the interior zone. ●

●

●

Since in the interior zone cold air is often still required to of fset the zone cooling load in the winter season, zone temperature control is the same as that in a VAV cooling central system. For each control zone in the perimeter zone, a reheating VAV box is installed. As discussed in Sec. 21.3, when the zone temperature drops within the range 72.0 Trxn

75.0°F (22.2 Trxn 23.9°C), zone temperature control is in deadband mode. The damper in the reheating VAV box is at 30 percent minimum setting, and the reheating coil is deenergized. When the zone temperature Trxn 72°F (22.2°C), zone temperature control is in heating mode. The DDC controller opens the two-way valve of the reheating coil and modulates its entering water mass f ow rate in a reverse-acting mode to maintain a preset zone temperature. The cold primary air is still at 30 percent minimum setting to provide outdoor ventilation air to the control zone.

In a perimeter-heating VAV central system, the zone temperature controls of the control zones in the interior zone during the cooling mode design load and part-load operations are similar to those in the VAV reheat central system e xcept that a VAV box is installed in each of the control zones in the perimeter zone. The operating mode of the zone temperature control in control zones of the perimeter zone is often automatically changed o ver from cooling mode to heating mode and vice v ersa. There is also a deadband mode between the cooling and heating modes. The width of the deadband mode is often limited to 1 to 3°F (0.6 to 1.7°C). When the zone temperature of a control zone in the perimeter zone Trxn 72°F (22.2°C), zone temperature control of the perimeter -heating VAV central system is in heating mode operation. In heating mode: ●

●

●

●

A DDC controller actuates the reverse action relay. The cold primary air supplied from the VAV box in the perimeter zone is still at 30 percent minimum setting. The hot water temperature leaving the boiler is reset according to the outdoor air temperature. The controller opens and modulates the mass f ow rate of hot water entering the f nned-tube baseboard heater in the perimeter zone to maintain a preset temperature of 72°F (22.2°C).

System Characteristics System characteristics for VAV cooling central systems, VAV reheat central systems, and perimeterheating VAV central systems are listed in Table 30.1.

30.7 DUAL-DUCT VAV CENTRAL SYSTEMS SH__ ST__ LG__ DF

System Description A dual-duct VAV central system is a multizone system that has a central plant and water systems to supply hot and chilled w ater to the w ater heating and cooling coils in the AHUs to condition

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the supply air; it distrib utes to various control zones in the perimeter zone by using a w arm air duct and a cold air duct through mixing VAV box es, diffusers, and controls, and to v arious control zones in the interior zone using only a cold duct through VAV box es, diffusers, and controls. Only central systems have field-built dual-duct VAV air systems. The number of supply fans, operating characteristics of mixing VAV box es, zone temperature control and sequence of operations, discharge air temperature control, and zone supply v olume flow rate of a dual-duct VAV central system are the same as those discussed in Sec. 21.4. The dual-f an dual-duct (cold and warm duct has its own supply fan) VAV central system allows the use of different ratio of outdoor air to supply air, simplifies its control and is more ener gy-efficient than a single-supply-fan dual-duct VAV central system. A dual-f an dual-duct VAV central system is sho wn in Fig. 21.9. The supply volume f ow rate, cooling coil load, and the heating coil load can be calculated in the same way as discussed in Section 21.4.

__RH TX

System Characteristics System characteristics of a dual-duct VAV central system are listed in Table 30.1.

30.8 FAN-POWERED VAV CENTRAL SYSTEMS System Description A f an-powered VAV central system is a multizone system that has a central plant and w ater systems to supply hot and chilled w ater to the AHUs and VAV box es, water-cooling coils in the AHUs, and reheating coils in the f an-powered VAV box es to cool and heat the supply air and distribute it to v arious control zones in the perimeter zone through ducts, fan-powered VAV boxes, diffusers, and controls, and to v arious control zones in the interior zones through ducts, VAV boxes, diffusers, and controls. There are two types of fan-powered VAV box: parallel fan-powered VAV box and series f an-powered VAV box, as discussed in Sec. 21.5. P arallel fanpowered VAV box es are more widely used because the y are more ener gy-efficient. A f anpowered VAV central system with parallel f an-powered VAV boxes is similar to that sho wn in Fig. 21.13. The fan energy use, fan characteristics, zone temperature control, and design considerations of a fan-powered VAV central system are similar to those discussed in Sec. 21.5.

Zone Supply Volume Flow Rate and Coil Load Zone supply volume f ow rates of a fan-powered VAV central system are the same as those in a f anpowered VAV packaged system, as discussed in Secs. 21.5 and 29.7. For conventional air distribution in the perimeter zone, zone peak supply volume f ow rate V˙sxn,d can be calculated from Eq. (29.17). F or conventional air distrib ution in interior zone, zone peak supply v olume f ow rate V˙sin,d can be calculated from Eq. (29.18). F or cold air distrib ution in perimeter zone, zone supply volume f ow rate of cold primary air V˙sxn,c can be calculated from Eq. (29.19) For cold air distrib ution, the volume f ow rate of the recirculating plenum air can be calculated from Eq. (29.20) as V˙sxn,r  V˙sxn  V˙sxn,c The reheating coil load in the f an-powered VAV box Qchxn can be calculated from Eq. (29.16). The cooling coil load of the AHU at summer design conditions Qcc can be calculated from Eq. (29.15).

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Case Study: A Fan-Powered VAV Central System

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A fan-powered VAV central system was designed and constructed for the 34-story off ce building of the Taipei World Trade Center, Taipei, Taiwan. It has a gross f oor space of 1.05 million ft 2 (97,580 m2) and a total air conditioned f oor space of 880,310 ft2 (81,810 m2). The peak refrigeration load of the building is about 2530 tons (8890 kW). The local utility rate structure strongly f avors the use of the refrigeration system during of f-peak hours. This project w on f rst place in the 1991 ASHRAE Technology Awards for HV AC&R system designs for commercial b uildings. The entrants were Hsing-Chung Yu and Carl E. Claus. Refrigeration System. The ice storage system w as a partial storage one (ice storage systems are discussed in the ne xt chapter). A separate f ooded liquid cooler w as connected to the scre w compressors to pro vide direct cooling during on-peak hours as well as during after -hours operation, to offset the signif cant nighttime load. During direct cooling, the compressors were controlled to operate at a higher suction temperature of 33°F (0.56°C) to conserve energy. A liquid o verfeed system w as used instead of direct e xpansion during the ice-making period because of its higher heat transfer . About 20 percent of the refrigerant coil surf ace in the ice builders could be saved. During ice melting, chilled water was supplied directly through a closedloop water circuit to the storage tanks, and the tanks were pressurized. Such an arrangement obviates the use of a heat e xchanger and thus a voids the corresponding chilled w ater temperature rise. Condensing heat was eff ciently rejected through evaporatively cooled condensers. Air Systems. Because of the lo wer chilled w ater temperature from the storage tank during ice melting, cold primary air at 45 °F (7.2 °C) was introduced at the parallel f an-powered unit. It w as mixed with the induced plenum air to produce a supply air of 56 °F (13.3°C) at the diffuser. Such a lower primary air temperature had the following consequences: ●

●

Summer space relati ve humidity dropped to 40 to 45 percent. The thermal comfort of the occupants was greatly improved. Cold primary air volume f ow rate was reduced to 40 percent compared with conventional air distribution.

Two built-up centralized AHUs were installed in the basement. Each w as used to supply 200,000 cfm (94,380 L / s) of primary air from the bottom to the top of the entire b uilding. Compared with the originally proposed con ventional system with additional equipment f oors at mid-le vel, this air distribution system saved considerable rental space. The centralized air system, having f oor-by-f oor automatic shutof f dampers, was closely matched with the roof-mounted smok e control system. In case of a b uilding f re, the return f an would be stopped and the rooftop e xhaust fan would be started to pur ge the smok e from the f re f oor through the opening and closing of dampers connected to the smok e exhaust system. At the same time, air would be supplied to the tw o f oors above the f re f oor as well as one f oor below to pressurize these f oors. The air system also did the following: ●

●

●

●

SH__ ST__ LG__ DF

Provided 15 cfm (7.1 L / s) of outdoor air for each occupant Included cartridge-type f lters and upstream pref lters with a dust spot eff ciency of 85 percent Used controllable-pitch axial fans for supply and return air Included electric heaters for winter heating

A microprocessor -base DDC system w as used for zone temperature; AHU operating parameters; ice-charging, discharging, and direct cooling; and emergency smoke controls.

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Cost and Electric Demands. The actual bidding price of the HVAC&R system for this project w as approximately $9 million, or $8.57 per ft 2 ($92.2 per m 2) gross area. Compared with the originally proposed conventional air conditioning system using centrifugal chillers, the air system po wer input was reduced from 1679 to 1136 kW , and the electricity demand dropped from 5960 to 4330 kW . The estimated total annual ener gy consumption per ft 2 gross area w as 33,190 Btu / ft2 yr (105 kWh / m2  yr).

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System Characteristics System charateristics of fan-powered VAV central systems are listed in Table 30.2. TABLE 30.2 System Characteristics for Fan-Powered VAV Central Systems (FPVAVCS) and Clean-Room Systems FPVAVCS Zone thermal and sound control Control zone Control methods Control modes Heating-cooling mode changeover Sound control Indoor air quality (IAQ) Minimum ventilation control Filters Humidity control Humidif ers Air system Types Supply fan Supply fan total pressure Combined fan-motor-drive eff ciency Relief / return fan Relief / return fan total pressure Combined fan-motor-drive eff ciency Fan-powered VAV box fan Box fan total pressure Combined fan-motor-drive eff ciency Air economizer Cooling systems Refrigeration compressor Refrigerants Evaporator Condenser Energy performance (compressor) Heating system

Boiler AFUE Maintenance Fault detection and diagnostics

Multizone DDC PI or PID Automatic 35 – 45 dBA

CRS Single or multizone DDC, or electric PI or PID Cooling mode year-round, summer-winter mode changeover, automatic NC 50 – 65

DCV or MPC Medium to high eff ciency Optional

Constant-volume HEPA or ULPA with pref lters Dew point control Steam, or heating element

VAV air mixing Forward, or airfoil centrifugal 4.5 – 5.5 in. WC 55% Centrifugal or axial Relief 0.6 in. WC, return 0.5 – 1 in. WC 35 – 40% Centrifugal, forward 0.5 in. WC 25% Fixed or differential dry-bulb or enthalpy

MAU and RAU conditioning and mixing Centrifugal or vane-axial MAU 4– 4.5 in. WC, RAU 2– 2.5 in. WC MAU 55%, RAU 55% Centrifugal or axial Return 0.5 – 1 in. WC 35 – 40% N/A

Centrifugal or screw HCFC-123, HCFC-22, HFC-134a Flooded liquid cooler Water cooled or evaporatively cooled 0.5 – 0.7 kW / ton Hot water coil, or electric heater

Conventional 78%, condensing 93%

Centrifugal or screw HCFC-123, HCFC-22, HFC-134a Flooded liquid cooler Water cooled or evaporatively cooled 0.5 – 0.7 kW / ton Space, year-round cooling, preheating — outdoor air, reheating — part load; hot water, electric heating Conventional 78%, condensing 93%

AHU, chiller

AHU, chiller, f lters

N/A

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30.9 CLEAN-ROOM SYSTEMS

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System Description A clean-room or clean space central air conditioning system, or simply a clean-room system (CRS) or clean space system, has a central plant and w ater systems to supply hot and chilled w ater to the AHUs, uses HEPA and ULPA f lters with pref lters and water coils in the AHU to remove air contaminants, and conditions the air . The conditioned air is supplied to the conditioned space through ducts, terminals, and air distribution devices in order to provide an indoor environment strictly controlled with the required cleanliness, temperature, relative humidity, airf ow pattern, pressurization, and noise. A clean-room system is shown in Fig. 30.1 and has the following characteristics: ●

●

●

●

Because of the complexity of the system con f guration and the higher requirements in the control of indoor environments, a clean-room system is a custom-built central system with AHUs and water-cooling and heating coils to condition the supply air. A clean-room system is required to provide airf ow of specif c velocity to reduce lateral air contamination. Therefore, a clean-room system is a constant-volume system. A clean-room system can be either a single-zone system or a multizone system. In a multizone clean-room system, the zone reheating coil is used to compensate for the v ariation in zone sensible load to maintain a nearly constant preset zone temperature. A clean-room system al ways has a separate mak eup air unit (MA U) to condition the outdoor air , and a recirculating air unit (RA U) recirculates the space air, f lters it, cools it, and pressurizes the mixture of outdoor and recirculating air. Such an arrangement minimizes the cross-contamination of airstreams as well as consolidates the f ltration of the outdoor air.

Clean rooms and clean spaces are widely used in semiconductor, pharmaceutical, aerospace, and health care industries and facilities. Airflow The volume f ow rate of the cleaned and conditioned air supplied to the clean room depends on the desirable air velocity that must be pro vided in the w orking area of the clean room. As discussed in Sec. 20.17, the supply volume f ow rate V˙s, in cfm (L / s), can be calculated from Eq. (20.73) as V˙s  Arvr

SH__ ST__ LG__ DF

(30.4)

According to ASHRAE Handbook 1999, HVAC Applications, U.S. Federal Standard 209E does not specify v elocity requirements. The 90 fpm (0.45 m / s) f gure is still widely used in clean rooms. Current research suggests lo wer velocity may be possible if the required cleanliness le vels can be maintained. Proper airf ow pattern is essential to predict the paths of the airstreams as well as to prevent contaminants from being deposited on critical surf aces in the w orking area. In clean rooms, there are two types of air f ow pattern: unidirectional airf ow, as shown in Fig. 30.1 a, and nonunidirectional airf ow. In a unidirectional air f ow pattern, airstreams f ow through the w orking area of the clean room in a single-pass, single direction of parallel airstreams. The unidirectional airf ow can be subdivided into vertical unidirectional airf ow and horizontal unidirectional airf ow. When the ceiling of a clean room is fully co vered by HEPA or ULPA f lters, the downward airf ow produces a unidirectional f ow of ultraclean air that covers the working area of the clean room. Baylie and Schultz (1994) reported for clean rooms with ceiling partly co vered by HEPA or ULPA f lters, a porous membrane that is added beneath the HEP A or ULPA f lters forms a small plenum which equalizes the pressure across the f ace of the ceiling and minimizes the turb ulence created by the larger grid required to support the HEPA f lters and lighting f xtures. When a membrane ceiling

~

(b) FIGURE

30.1

A clean-room

system for a class 10 clean room: (a) schematic diagram; (b) air conditioning

cycle.

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is added and air is returned from the bottom inlets of the sidewall, the downward airf ow underneath the membrane ceiling is in the form of unidirectional airf ow. For a clean room with a ceiling that is partly co vered by HEPA or ULPA f lters and without any porous membrane underneath the f lters, the downward airf ow in the clean room is in the form of nonunidirectional airf ow.

Pressurization Clean rooms and clean spaces al ways maintain a higher pressure than the surrounding less clean space to minimize the inf ltration of air contaminants. The following pressure differentials are often used: Pressure diffferential, in. WC (Pa) Between clean rooms or clean space and nonclean space Between clean rooms and less clean rooms

0.05 (12.5) 0.02 – 0.03 (5– 7.5)

Pressure control precision is typically between 0.01 and 0.03 in. WC ( 2.5 and 7.5 Pa). For a door or opening between tw o clean rooms of dif ferent cleanliness requirements, a minimum air velocity in the range of 15 to 50 fpm (0.07 to 0.25 m / s) should be maintained at the door (when it is opened) or the other openings.

Temperature and Relative Humidities Because staff wear heavy gowns in clean rooms with stringent indoor en vironmental requirements, a temperature between 66 and 68 °F (18.9 and 20 °C) is to be maintained. In class 10,000 or class 100,000 clean rooms with less restrictive garments, 70 to 72°F (21.1 to 22.2°C) may be satisfactory. A tolerance of 2°F (1.1°C) is adequate for comfort purposes. In clean rooms, space relative humidity is usually controlled at 45 5 percent. System Characteristics System characteristics of a clean-room system are listed in Table 30.2.

30.10 CASE STUDY: CLEAN-ROOM SYSTEMS FOR SEMICONDUCTOR INTEGRATED-CIRCUIT FABRICATION Indoor Requirements

SH__ ST__ LG__ DF

The fabrication of semiconductor inte grated circuits requires a highly sophisticated combination of advanced technologies. Clean-room central system with vertical unidirectional airf ow is one of these technologies in semiconductor w afer fabrication. It must meet stringent air quality requirements in air cleanliness, air temperature, humidity, airf ow pattern, pressurization, lighting, noise, and vibration for the sak e of successful manuf acturing. Contaminated semiconductors result in inferior products. As discussed in Sec. 4.11, clean-room air cleanliness requirements for semiconductor f abrication include classes 1, 10, 100, 1000, and 10,000. Vertical unidirectional air f ow with a verage air velocities from 60 to 90 fpm (0.3 to 0.45 m / s), typically at 90 fpm (0.45 m / s), is widely used.

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Manufacturing an integrated circuit involves photolithography, etching, and diffusion processes. These processes may take place over many hours, even days. A stable temperature is extremely important. Because clean-room production personnel wear smocks that fully co ver them, clean-room temperatures are controlled from 68 to 72 °F (20 to 22.2 °C) with tolerances of 0.1°F (0.056°C), 0.2°F(0.11°C), 0.5°F (0.28 °C), and 1.0 °F (0.56 °C). A closer tolerance is often maintained within the manuf acturing process itself; e.g., wafer reticle writing by electron beam technology needs 0.1°F (0.056°C). whereas 1.0°F (0.56°C) is often used for the open-bay area. In clean rooms, if the relative humidity is too lo w, static electricity is created and causes defective products. If the relati ve humidity is too high, some chemicals may e xpand and may cause equipment failure. For etching and dif fusing areas, the humidity should be maintained at 40 to 45 percent with a tolerance of 5 percent, whereas in the photolithography area, a relative humidity of 35 to 40 percent with a tolerance of 2 percent is usually maintained. The manufacture of metal oxide semiconductors requires large quantities of conditioned outdoor air as mak eup air, to replace processing e xhaust air and to maintain clean-room pressurization. A clean room is always maintained at a positive pressure to prevent the inf ltration of contaminated air from surrounding spaces. F or some clean rooms, the process exhaust airf ow may be as high as 10 cfm / ft2 (182 m 3 /h m2) of f oor area. The average process exhaust airf ow for semiconductor f abrication may be between 2 and 3 cfm / ft2 (36 to 54 m3 /h m2).

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Energy Use of Components In Naughton (1990a, 1990b), the breakdown of ener gy use in a typical clean-room central system with vertical unidirectional f ow is approximately as follows: Manufacturing equipment (75 W / ft2 or 807 W / m2) Air conditioning (HVAC&R) Electric lights Building envelope and others

50 percent 40 percent 6 percent 4 percent

The chiller and f ans each consume 45 percent of the total HV AC&R energy use. Both plant and building pumps use the remaining 10 percent. Because of extremely high space sensible cooling load and very small space latent load, the sensible heat ratio of the space conditioning line SHRs can often be taken as 0.99. The operating cost of HVAC&R is only 5 to 20 percent of the total cost needed to produce an integrated circuit (semiconductor w afer). The air cleanliness, temperature, relative humidity, airf ow, and pressurization required for successful f abrication are still e xtremely important goals of an HVAC&R system design. Because of high utility rates, however, a reduction in the operating cost of the clean-room system also becomes a v ery inf uential factor in clean-room design and operation.

System Description Semiconductor clean rooms are of tw o types: open-bay design or clean tunnel. The clean tunnel consists of narrow modular clean rooms which may be isolated from one another . Open-bay design includes lar ge open-construction clean rooms. A typical clean-room air conditioning system, or simply clean-room system, for a class 10 clean room of the open-bay con f guration is shown in Fig. 30.1. This system consists of the following components: Recirculating Air Unit (RAU). The function of a recirculating air unit is to recirculate the space air; to f lter it, to cool it, to pressurize the mixture of recirculating and conditioned mak eup air, and to force the mixture to f ow through the ULP A f lters and the clean-room w orking area. An RAU

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comprises the following components: ●

●

●

●

●

●

A mixing box to mix recirculating and conditioned makeup air A pref lter with a dust spot eff ciency of 30 percent A chilled water sensible cooling coil A recirculating fan Makeup air and recirculating air dampers Two sound attenuators, one located immediately before the f an inlet and the other after the f an outlet

Usually an axial f an is used as the recirculating f an because of its higher f an total ef f ciency (between 75 and 82 percent) and its better operating characteristics. An unhoused centrifugal fan, often called a cabinet fan, with a fan total eff ciency of 58 to 63 percent is sometimes used because of its lower sound power level. The chilled water entering the sensible cooling coil in the RAU is often at a temperature of 50°F (10°C). Makeup Air Unit (MAU). The function of an MAU is to supply outdoor air to the clean room for process exhaust and pressurization, to condition it, and to control the humidity of the clean room by cooling and dehumidifying, or heating and humidifying, the makeup air. The system components in an MAU include the following: ●

●

●

●

●

●

●

●

An outdoor air damper A pref lter with a MERV 8 (dust spot eff ciency of 30 percent) A preheating coil A chilled water cooling coil A makeup air centrifugal fan A HEPA f lter of 99.97 percent DOP eff ciency A humidif er, most often a steam humidif er MAU shutoff damper

Chilled water entering the cooling coil in the MA U for cooling and dehumidifying is at a temperature of 40 °F (4.4°C). It is more ener gy-eff cient to ha ve a separate chiller to pro vide chilled w ater for an MAU. ULPA Filters and the Unidirectional Airfl w Clean Room. The function of ULP A f lters is to provide ultraclean air for the clean room. The pressurized plenum or ducted ULP A f lter modules are often used for an even distribution of unidirectional downward air f ow. For class 1, 10, and 100 clean rooms, ULPA f lters with a DOP eff ciency of 99.9997 percent of 0.12- m particles and unidirectional air f ow are used. For class 1000 through class 100,000, HEPA f lters with a DOP ef f ciency of 99.97 percent of 0.3- m particles and nonunidirectional air f ow may provide satisfactory contamination control. Unidirectional downward airf ow produces a uniform air sho wer of ultraclean air . Internally generated contaminants will not mo ve laterally against the 90 fpm (0.45 m / s) airf ow and will be carried away by predictable parallel airstreams. Recirculating air enters either the bottom side return inlets directly or the perforated raised f oor panels. It is then returned to the RAU to mix with makeup air again.

Operating Characteristics SH__ ST__ LG__ DF

The following temperature and relative humidity are to be maintained in a class 10 clean room:

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Room temperature, year-round Space humidity, year-round

30.19

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69 1.0°F (20.6 0.56°C) 42.5 2.5 percent

For a class 10 clean room with an area of 1000 ft 2 (472 m 2), a supply v olume f ow rate of 90,000 cfm (42,470 m 2) is required to pro vide an air v elocity of 90 fpm (0.45 m / s) in order to produce a unidirectional f ow in this clean room. The outdoor air intake for process exhaust, space pressurization, and occupants is typically 6000 cfm (2830 L / s). Summer Mode Operation In Mandelbaum (1991), the operating modes of clean rooms of classes 1 through 1000 are di vided into summer and winter modes. When the de w-point temperature of the outdoor air To is 46 °F (7.8°C) and above, the clean-room system is in summer mode operation. Let us consider a hot summer day . Outdoor air at a summer design temperature of 100 °F (37.8°C) and a wet-bulb temperature of 78°F (25.6°C) enters the MAU, as shown in Fig. 30.1a and b. It is cooled and dehumidi f ed at the cooling coil to a lea ving coil condition of air temperature Tcc  46°F (7.8°C), relative humidity r  99 percent, and dew-point temperature  46°F (7.8°C). After the conditioned outdoor air absorbs the f an heat of the MAU, the air enters the RAU at a discharge temperature Tdis of 47.5°F (8.6°C) and a relative humidity of 92 percent. In the RA U, conditioned air from the MA U is mix ed with the recirculating air from the clean room at a temperature of 69 °F (20.6°C), a relative humidity of 42.5 percent, and a dew-point temperature of 46 °F (7.8°C). The ratio of v olume f ow of recirculating air to mak eup air is 12:1. The mixture m enters the sensible cooling coil at temperature Tm  67°F (19.4°C) and a de w point of 46°F (7.8°C). It is then sensibly cooled to a temperature Tsc. If the maximum space sensible cooling load is 563,000 Btu / h (265,680 L / s) and if the density of supply air s  0.078 lb / ft3 (1.248 kg / m3), the temperature of supply air Ts can be calculated as Ts  Tr   69 

Q rs 60V˙sscpa 563,000  63.5F (17.5C) 60  90,000  0.078  0.243

(30.5)

Usually the temperature rise due to f an heat in the RAU is 1°F (0.56°C). The temperature of the air leaving the sensible cooling coil is then Tsc  63.5  1  62.5°F (16.9°C). Its dew-point temperature is still 46 °F (7.8°C). Because of the short supply duct, large volume f ow, and the fact that the surrounding space is conditioned, the duct heat gain is usually ignored. Because the SHRs of the space conditioning line is 0.99, after the air has absorbed the space sensible cooling load and a v ery small amount of latent load, the space temperature is then maintained at 69°F (20.6°C). The space relative humidity is 42.5 percent, and the dew point is 46°F (7.8°C). To cool and dehumidify the makeup air at the cooling coil to a leaving dew-point temperature of 46°F (7.8°C), the chilled water entering the cooling coil should be pro vided at a temperature Twe  40°F (4.4°C) without using glycol for freeze protection. The chilled water leaving the chiller in the plant loop is usually 1°F (0.56°C) lower, or 39°F (3.8°C). Part-Load Operation and Controls Two controls maintain the required space temperature at 69 °F (20.6 °C) and the relati ve humidity within acceptable limits during summer mode part-load operation: the dischar ge air temperature control incorporating de w point control of the MA U and the space (zone) air temperature control. When the dry- and wet-b ulb temperatures of the outdoor air To and To drop but the de w-point

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temperature is still at 46°F (7.8°C) or above, the discharge air temperature sensor T2 detects the fall in Tdis and the DDC controller modulates the tw o-way valve of the chilled w ater cooling coil in the MAU to reduce its w ater f ow until a 47.5 °F (8.6°C) discharge air temperature and a dischar ge air dew point of 46°F (7.8°C) are maintained. If the space sensible cooling load f alls below the design load, the space temperature drops accordingly. As the space temperature sensor T1 detects such a f all in temperature, the DDC controller modulates the two-way valve of the sensible cooling coil in the RA U, reduces its water f ow, and tends to maintain a constant space temperature of Tr  69°F (20.6°C). In addition to the MAU discharge air temperature and space temperature controls, there are controls for space humidity, space pressurization, f lter pressure drop monitoring, and smoke detection. Because the discharge air dew point from the MA U is 46 °F (7.8°C), the space dew point is also 46°F (7.8°C), and the mixture of recirculating and mak eup air still has a de w point of 46 °F (7.8°C). After sensible cooling in the RAU, the dew point of supply air, point s, remains at 46°F (7.8°C). Because the sensible heat ratio of the space conditioning line SHR s  0.99, the space relative humidity will be always around 42.5 percent if the space temperature Tr is maintained at 69°F (20.6°C). For a space lik e a clean room ha ving only a ne gligible latent load, supply air dew-point control can always maintain both space temperature and relati ve humidity within required limits by means of sensible cooling and reheating when the space sensible cooling load is varied. In Fig. 30.1a, there is a pressure sensor P1 in the clean room. This sensor measures the pressure differential pr between the clean room and the surrounding area. This pressure differential is generally maintained at 0.05 in. WC (12.5 Pa) to prevent inf ltration of contaminated air. According to the signal from P1, the DDC controller modulates the opening of the interlock ed makeup air and recirculating dampers. When there is an increase in mak eup air and a decrease in recirculating air in the mixing box of RA U, the space pressure differential tends to rise. A decrease in makeup air and an increase in recirculating air cause a drop in space pressure dif ferential. The supply of mak eup air is balanced by the process e xhaust air and the e xf ltrated air from door gaps due to the pressure differential. Pressure sensors are also used for MER V 8 pre f lters, HEPA f lters, and ULPA f lters. If the pressure drop e xceeds a predetermined limit, an indicating lamp f ashes or an alarm is ener gized. Both signals call for replacement. When the smok e detector is ener gized, a signal is sent to the b uilding f re control system. The f re alarm and the smoke control system are energized accordingly. Winter Mode Operation and Controls

SH__ ST__ LG__ DF

When the de w point of the outdoor air To is 44 °F (6.7°C) or belo w, the clean-room system is in winter mode operation. If To 44°F (6.7 °C) and To 53.5°F (11.9 °C), the outdoor air is preheated at the preheating coil to a temperature Tph  53.5°F (11.9°C). After absorbing the f an heat of the MA U, it enters the RA U at a dischar ge temperature Tdis  55°F (12.8 °C). If To 44°F (6.7°C) and To  53.5°F (11.9°C), the outdoor air is sensibly cooled to 53.5°F (11.9°C). Again, it is discharged to the RAU at Tdis  55°F (12.8°C) after absorbing fan heat. If makeup air at a temperature of 55 °F (12.8°C) and a de w point of 44 °F (6.7°C) is mixed with winter mode recirculating air of temperature 69°F (20.6°C) and 40 percent relative humidity (a dew point of 44°F, or 6.7°C), the resulting mixture will be at a temperature of 67.5 °F (19.7°C). The dew point will still be 44 °F (6.7 °C). If the mixture is cooled at the sensible cooling coil to 62.5 °F (16.9°C) according to the signal of the temperature sensor T1, and if it is supplied to the clean room at a temperature of 63.5°F (17.5°C) and a dew point of 44°F (6.7°C) after absorbing the fan heat of RAU, then it will maintain a space temperature Tr  69°F (20.6 °C), r  40 percent, and a de w point of 44°F (6.7°C) at a maximum space sensible cooling load of 563,000 Btu / h (164,960 W). A lower space sensible cooling load results in a decrease in the space temperature Tr. As the sensible cooling capacity at the cooling coil is modulated and reduced by the DDC controller, the space temperature will still be maintained at 69°F (20.6°C). If To 44°F (6.7°C), space relative humidity r may be less than 40 percent. When the space

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30.21

humidity sensor detects such a shortage, the DDC controller ener gizes the steam humidi f er in the MAU to humidify the mak eup air until r  40 percent. In winter mode operation, r will always be maintained around 40 percent. During winter mode operation, the chilled w ater temperature entering the cooling coil in the MAU is reset from 40 to 45 °F (4.4 to 7.2 °C) for a more ener gy-eff cient operation. Also, the preheating coil and the cooling coil in the MA U are sequentially controlled; the y are not ener gized simultaneously. In winter operation, outdoor air is preheated in the MA U, and the mixture of outdoor and recirculating air is again sensibly cooled in the RAU. Such simultaneous heating and cooling are mainly a result of the requirement for steam humidi f cation. It is not appropriate to humidify the air to near saturation. Uneven distribution may cause surface condensation. During winter mode operation, if the de w point of the outdoor air To is between 44 and 46 °F (6.7 and 7.8 °C), the operation mode remains in winter mode. When To  46°F (7.8 °C), it will change o ver to summer mode operation. Similarly , if 44 To 46°F (6.7 To 7.8°C), the operating mode remains in summer mode until To 44°F (6.7°C). Then it changes o ver to winter mode operation. Such a control strate gy prevents hunting between summer and winter mode operation during intermediate seasons.

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System Pressure According to Naughton (1990a, 1990b) and Hunt et al. (1990), the system pressure loss of an MAU is about 3.8 to 4 in. WC (950 to 1000 Pa). For an RAU, the system pressure loss is usually between 2.0 and 2.50 in. WC (500 and 625 Pa). A typical breakdown of the fan total pressure of the recirculating air fan in an RAU is as follows: System components RAU Pref lter Cooling coil Inlet sound attenuator Discharge sound attenuator Discharge elbow External pressure drop HEPA f lters, f nal Distribution duct work Perforated raised f oor panel Return path Total

Pressure drop, in. WC (Pa) 0.25 (62) 0.35 (87) 0.10 (25) 0.15 (38) 0.10 (25) 0.80 (200) 0.20 (50) 0.10 (25) 0.15 (38) 2.20 (550)

HEPA f lters of lo wer pressure drop, such as a f nal loaded pressure drop of 0.50 in. WC (125 Pa), are also a vailable. However, they are e xpensive and require more space. This f act should be carefully analyzed and considered by the production engineer and architect. To reduce the pressure drop across coils, the most effective method is to reduce their face velocities to 300 to 400 fpm (1.5 to 2 m / s). The result is a larger MAU and RAU. Again, the compromise is between initial and operating costs. All duct f ttings along the air f ow must be carefully designed to reduce pressure losses. Square elbows should al ways be installed with splitter v anes. High-eff ciency f ans and pumps should be used, along with high-COP and -EER chillers. According to Naughton (1990a, 1990b), the energy use of fans, chillers, and pumps in a clean-room system for a class 1 clean room may be around 50 W / ft2 (538 W/ m2) of f oor area.

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FIGURE 30.2 Difference in f nal and initial pressure drops of f lters affects system performance. ( a) Airfoil blade centrifugal f an; (b) forward-curved centrifugal fan; (c) vane-axial fan.

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FIGURE 30.2 (Continued)

Effect of Filter Final-Initial Pressure Drop Difference on System Performance During the calculation of the system pressure loss of an air system for the purpose of selecting f ans and AHUs, one should use the f nal pressure drop of the f lter pf.f, in in. WC (Pa), to provide the required volume f ow rate whene ver f lters are either clean or loaded with speci f ed capacity. For purposes of ener gy estimation, the average pressure drop of the f lters during the w orking period pf,m, in in. WG, should be employed instead of f nal or initial pressure drops. For most pref lter and ultrahigh-eff ciency air f lter assemblies used in AHUs, the difference between the f nal pressure drop and the initial pressure drop of the f lter assembly is 0.4 to 0.8 in. WC (100 to 200 P a). Let us consider an air system whose design v olume f ow rate V˙ is 30,000 cfm (14,160 L / s), and whose system pressure loss when f lters are loaded psy is 2.2 in. WC (550 Pa). At V˙  30,000 cfm (14,157 L / s), the difference between the f nal and initial pressure drops is 0.5 in. WC (125 Pa). The effect of the dif ference in f nal and initial pressure drops of a f lter on system performance is illustrated in Fig. 30.2. In Fig. 30.2 a, an airfoil-blade centrifugal f an is used. The fan and AHU are selected according to these criteria: V˙  30,000 cfm (14,160 L / s) and psy  2.2 in. WC (550 Pa). If the f lters are loaded, the operating point is P. If the f lters are clean, the system pressure drops to 1.7 in. WG (425 P a). The operating point mo ves along the selected f an curve to point Q, which has V˙  33,000 cfm (15,573 L / s) at p  1.7 in. WC (425 Pa). In Fig. 30.2b, a forward-curved centrifugal fan is used. A forward-curved fan has a slightly f atter fan curve at operating point Q, and the system volume f ow rate increases to 33,500 cfm (15,810 L / s). For Fig. 30.2 c, a vane-axial fan is used. A vane-axial fan has a v ery steep f an curve. Therefore, at point Q the system volume f ow rate is increased to only about 31,000 cfm (14,630 L / s). The air system in a clean-room system is al ways a constant-v olume system. The fan power inputs at point Q for airfoil blade and v ane-axial f ans are approximately the same as at point P, whereas for forward-curved fans, their fan power input is comparatively greater.

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During the selection of AHUs, if the fan performance of the AHU is expressed in volume f ow V˙ versus fan total pressure pt, the designer should ascertain that the pressure drops of the loaded f lters are included in pt, not those of the clean ones. Design Considerations The successful construction of a class 1, 10, or 100 clean room depends on the combined ef fort of architects, mechanical engineers, production engineers, owners, contractors, HVAC&R equipment manufacturers, and many others. To maintain a temperature tolerance of 0.1°F (0.056°C) for a manufacturing process, it is important to stabilize all internal loads during the processing period. The process should be surrounded by conditioned space for which the tolerance is 0.5°F (0.3°C). Ev en mo vement of production personnel to ward the processing area may cause radiant heat turbulence and temperature f uctuations. Energy use per ft 2 (m2) f oor area of a clean-room system for class 1 through class 100 clean rooms is 5 to 10 times higher than that of an air conditioning system for a commercial b uilding. Careful analyses and impro vements based on e xperience with similar projects will pro vide a satisfactory indoor environment for the manufacturing process and also will reduce the energy use of the clean-room system. Semiconductor products that are being produced no w may become obsolete within a few years. As with many other HVAC&R systems, air conditioning system design for clean rooms must incorporate f exibility for change and future development.

REFERENCES

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ASHRAE, ASHRAE Handbook 1996, HVAC Systems and Equipment, ASHRAE Inc., Atlanta, GA, 1996. ASHRAE, ASHRAE Handbook 1999, HVAC Applications, Atlanta, GA, 1999. ASHRAE, ASHRAE / IESNA Standard 90.1-1999, Energy Standard for Buildings Except New Low-Rise Residential Buildings, Atlanta, GA, 1999. Atkinson, G. V., and G. R. Martino, Control of Semiconductor Manufacturing Cleanrooms, ASHRAE Transactions, 1989, Part I, pp. 477 – 482. Austin, S. B., HVAC System Trend Analysis, ASHRAE Journal, no. 2, 1997, pp. 44 – 50. Baylie, C. L., and S. H. Schultz, Manage Change: Planning for the Validation of HVAC Systems for a Clinical Trials Production Facility, ASHRAE Transactions, 1994, Part I, pp. 1660 – 1668. Garr, H. B., Clean Room Humidity Control, Heating / Piping / Air Conditioning, no. 3, 1992, pp. 67 – 70. Hunt, E., D. E. Benson, and L. G. Hopkins, Fan Eff ciency vs. Unit Eff ciency for Cleanroom Application, ASHRAE Transactions, 1990, Part II, pp. 616 – 619. Linford, R. G., and S. T. Taylor, HVAC Systems: Central vs. Floor-by-Floor, Heating / Piping / Air Conditioning, no. 7, 1989, pp. 43 – 58. Mandelbaum, I., HVAC Modif cations for Semiconductor Fabrication, Heating / Piping / Air Conditioning, no. 3, 1991, pp. 29 – 32. Naughton, P., HVAC Systems for Semiconductor Cleanrooms — Part 1: System Components, ASHRAE Transactions, 1990a, Part II, pp. 620 – 625. Naughton, P., HVAC Systems for Semiconductor Cleanrooms — Part 2: Total System Dynamics, ASHRAE Transactions, 1990b, Part II, pp. 626 – 633. Ottmer, J. H., Central Plant Cools and Heats World’s Largest Airport, Heating / Piping / Air Conditioning, no. 5, 1994, pp. 43 – 47. Rose, T. H., Noise and Vibration in Semiconductor Clean Rooms, ASHRAE Transactions, 1986, Part I B, pp. 289 – 298. Schneider, R. K., How to Design a Cleanroom, Engineered Systems, no. 3, 1996, pp. 58 – 62.

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Schneider, R. K., Healthy Cleanroom Design, Engineered Systems, no. 7, 1997, pp. 46 – 54. Schuler, M., Dual Fan, Dual-Duct System Meets Air Quality, Energy-Eff ciency Needs, ASHRAE Journal, no. 3, 1996, pp. 39 – 41. Stokes, R., The System Choice, Heating / Piping / Air Conditioning, July 1983, pp. 87 – 89. Tao, W., and R. R. Janis, Modern Cooling Plant Design, Heating / Piping / Air Conditioning, May 1985, pp. 57 – 81. Yu, H. C., and C. E. Claus, Thermal Storage System Cools Off ce Building, ASHRAE Journal, March 1991, pp. 15 – 17.

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CHAPTER 31

AIR CONDITIONING SYSTEMS: THERMAL STORAGE SYSTEMS 31.1 THERMAL STORAGE SYSTEMS 31.1 System Description 31.1 Building Energy Consumption and Thermal Storage Systems 31.2 Electric Deregulation and the Impact on Thermal Storage Systems 31.2 Benefits and Drawbacks of Thermal Storage Systems 31.2 Full Storage and Partial Storage 31.3 Ice Storage and Chilled Water Storage 31.5 31.2 ICE-ON-COIL, INTERNAL-MELT ICE STORAGE SYSTEMS 31.6 System Description 31.6 Brine and Glycol Solution 31.6 Ice Storage Tank 31.7 Case Study: Operating Modes of Ice-on-Coil Ice Storage System 31.7 System Characteristics 31.9 31.3 ICE-ON-COIL, EXTERNAL-MELT ICE STORAGE SYSTEMS 31.10 System Description 31.10 Ice Builders 31.11 Refrigerant Feed 31.11 Ice-Charging Control 31.11 System Characteristics 31.11 Case-Study: An Ice-on-Coil External-Melt Ice Storage System 31.13 31.4 ENCAPSULATED ICE STORAGE SYSTEMS 31.13 System Description 31.13 Location of Chiller and Storage Tank 31.14 Controls 31.14

Charging and Discharging 31.15 System Characteristics 31.15 31.5 ICE-HARVESTING ICE STORAGE SYSTEMS 31.15 System Description 31.15 Ice Making or Charging 31.16 Chiller Operation 31.17 System Characteristics 31.17 31.6 COMPARISON OF VARIOUS ICE STORAGE SYSTEMS 31.17 31.7 STRATIFIED CHILLED WATER STORAGE SYSTEMS 31.18 System Description 31.18 Basics for Chilled Water Storage 31.18 Storage Tanks 31.19 Stratified Tanks 31.19 Temperature Gradient and Thermocline 31.20 Diffusers 31.20 Charging and Discharging Temperature versus Tank Volume 31.22 System Characteristics 31.23 31.8 CASE STUDY: A STRATIFIED CHILLED WATER STORAGE SYSTEM 31.23 Chilled Water Storage System 31.23 Concentric Double-Octagon Diffusers 31.24 Charging Process 31.26 Discharging Process 31.26 Part-Load Operation 31.27 System Performance 31.28 REFERENCES 31.28

31.1 THERMAL STORAGE SYSTEMS System Description A thermal energy storage air conditioning system, or simply a thermal storage system, consists of a central plant, a chilled water or brine system incorporated with a thermal storage system, a hot water system, an air system including AHUs , terminals, return air system, smoke control systems, and 31.1

31.2

CHAPTER THIRTY-ONE

mechanical exhaust systems. In addition, the electric-driven refrigeration compressors in the central plant are operated at off-peak or at off-peak and on-peak hours. Stored ice or chilled w ater in tanks is used to pro vide cooling in b uildings during on-peak hours when there is a higher electric rate. Hot water from the boilers in the central plant is used to pro vide heating in winter . Air is conditioned in the AHUs and terminals. The conditioned air is then distrib uted to v arious control zones through ducts, terminals, diffusers, and controls. A thermal storage system is ●

●

●

Always a central system that uses stored chilled water or brine from the central plant to cool the air Often a multizone VAV system Often a cooling storage system

Building Energy Consumption and Thermal Storage Systems Lorsch (1993) analyzed the ener gy consumption for a 264,000 ft 2 (24,536 m 2) building using the Ontario Hydro electric rate structure. The chiller equipment needed 31 percent of the on-peak demand but consumed only 8 percent of the annual kWh. The annual load f actor for this b uilding is only 37 percent. Commercial b uildings have poor load profiles of electric p wer demand. Electric utilities are a capital-intensive industry. To improve their load f actor and to sell more kWh, electric utilities tend to shift daytime, more ener gy-inefficient diesel and gas-turbine plant operation t nighttime base-load highly efficient coal and nuclear plant operation. Electric utilities o fer a higher energy rate in daytime on-peak hours and a lo wer energy rate in of f-peak hours. A thermal storage system shifts a part of the electric po wer demand of the HV AC&R system from the daytime onpeak hours, usually from noon to 8 p.m., to nighttime off-peak hours. A thermal storage system often consumes approximately the same amount of electricity as a conventional air conditioning system; i.e., a thermal storage system does not necessarily sa ve energy. However, a thermal storage system significantly reduces ene gy cost. Electric Deregulation and the Impact on Thermal Storage Systems Because of the electric deregulation and the use of real-time pricing (RTP) and other electricity rate structures instead of the time-of-use (T OU) rate structure, as discussed in Sec. 25.7, open-market competition creates pressure on utilities to cut costs. According to Silvetti and MacCracken (1998) and the analyses from EIA, the following trend may dominate: ●

●

●

There is a conflict between the interests of the generationwning utilities in higher -generation prices and the ef fects of some demand side management (DSM) programs to reduce demand and possibly to help hold down competitive prices for generation. DSM programs are causing problems. Regarding energy-efficient equipment-related incent ve plans, something is wrong when a po wer producer gives money to a customer to purchase less po wer, particularly if the financial incent ve might eventually benefit a competito . Thermal storage systems benefit the p wer provider, the customer, the power user, and the industry setting the price for that energy.

Benefits and Drawbacks of Thermal Storage Systems Silvetti and MacCracken (1998) noted why a thermal storage system is attractive: ●

Thermal storage systems are one of the fe w le gitimate tools which shift the higher electric demand / electric rate for HV AC&R fully or partially from on-peak hours to a lo wer electric demand / electric rate in off-peak hours, and therefore lower operating costs.

AIR CONDITIONING SYSTEMS: THERMAL STORAGE SYSTEMS ●

●

●

●

●

31.3

Thermal storage systems reduce the equipment size and save initial cost. Thermal storage systems do not ne gatively impact the b uilding’s indoor en vironmental control operations, as the load shedding or some load control programs have done. As a thermal storage system is a central system, it uses chilled water from the central plant to cool the air in the AHUs. Thermal storage systems are easily adaptable to the water system of a central air conditioning system. As discussed in Sec. 1.5, central systems serv ed 24 percent of the floo space of commercial buildings in 1995. Real-time pricing and other rate structures are a vailable as the power providers maneuver to offer the most competitive structure in the electric deregulation environment. Thermal storage systems have longer operating hours of compressors and pumps, and chillers and cooling to wer at full-load operation, at lo wer outdoor temperatures at nighttime, and a backup source for cooling during emergencies.

Drawbacks of thermal storage systems include high initial cost and complicated operation, maintenance, and control.

Full Storage and Partial Storage The ton-hour, or ton  h (kWh), is the unit of stored refrigeration. One ton-hour is the refrigeration or heat absorption of 12,000 Btu (3.516 kWh) performed by a refrigeration system during a 1-h period. The aim of thermal storage strate gy is to incur the lo west possible ener gy cost and initial investment so that life-cycle costs are minimized. The economic benefit of a thermal storage syste can also be assessed by calculating the simple payback period of its cost. A simple payback or lifecycle cost analysis of the building load profile utility electric rate structure, and system characteristics is always necessary. Determination of the optimum size of a thermal storage system is based on the utility’ s electric rate structure (difference between on-peak and of f-peak unit charges) and the building refrigeration load profile Direct cooling denotes the process by which compressors produce chilled w ater to cool the building directly. When the cost of the direct cooling by a refrigeration system is lo wer than the cost of stored energy, the operation of the thermal storage system is said to be at chiller priority . On the other hand, if the cost of the direct cooling is higher than the cost of stored energy, the operation is said to be at storage priority . Construction cost and the utility’ s incenti ve payments should be considered. If the ener gy cost dif ference between on-peak and of f-peak hours is great, the full use of stored energy during on-peak hours may be most economical. There are tw o kinds of thermal storage: full storage and partial storage. Figure 31.1 sho ws the load-time diagrams for full storage and partial storage. F or a full-storage, or load shift thermal storage, system, all refrigeration compressors cease to operate during on-peak hours, and the b uilding refrigeration load during that period is entirely offset by the chilled water supplied from the thermal storage tank, as shown in Fig. 31.1a. Partial storage, or load leveling, can be either in load-leveling mode, in which refrigeration compressors are operated at full capacity during on-peak hours, as shown in Fig. 31.1 b, or in demandlimited mode, in which b uilding electric demand limits only part of the refrigeration compressors operated, as shown in Fig. 31.1c. A utility’s demand charge is the b uilding total demand char ge, which is the sum of the demand charges for HVAC&R systems and other uses including the electricity for lighting, escalators, computers, and electric appliances. The following is a comparison of dif ferent storage strate gies of a thermal storage system with heat reco very and cold air distrib ution for an of fice uilding in Dallas, Texas, as described by Tackett (1989).

CHAPTER THIRTY-ONE

Ice making

Refrigeration load Qrl, ton

On-peak

Off-peak Refrigeration load

1200

800

400

24 Midnight

6

12 Noon (a)

Off-peak Refrigeration load Qrl, ton

Direct cooling

Ice burning

Off-peak

18

On-peak

24 Midnight Off-peak

1200

800

400

24 Midnight

6

12 Noon (b)

Off-peak Refrigeration load Qrl, ton

31.4

18

On-peak

24 Midnight Off-peak

1200

800

400

24 Midnight

6

12 Noon (c)

18

24 Midnight

FIGURE 31.1 Full and partial storage: (a) full storage; ( b) partial storage, all compressors operating; (c) partial storage, 50 percent of compressors operating.

AIR CONDITIONING SYSTEMS: THERMAL STORAGE SYSTEMS

Full storage Utility incentive, $ Net incremental cost, $ Demand limit saving, kW Total saving, $ Simple payback, years

288,750 291,750 1,200 124,250 2.35

Partial storage 132,500 21,750 400 64,000 0.34

31.5

Demand limited 215,000 124,250 800 92,750 1.34

The larger the thermal storage system and the capacity of refrigeration compressors, the greater the savings. Actually, selection of the size of the storage system is better determined according to lifecycle cost, and system reliability should also be considered.

Ice Storage and Chilled Water Storage In the late 1970s, engineers began experimental thermal storage applications in buildings. Throughout the late 1990s, there are several thousand thermal storage systems operated in v arious commercial buildings including off ces, shopping centers, schools, and hospitals and industrial applications in the United States. Two thermal storage media are widely used for air conditioning systems: ice storage and chilled water storage. At a temperature dif ference of 18 °F (10°C), 1 lb (2.2 kg) of chilled w ater can store 18  1  18 Btu (19 kJ) of thermal ener gy, whereas 1 lb of ice can store 1  144  60  35  169 Btu (178 kJ). If the density of w ater is 62.3 lb / ft3 (997 kg / m3) and the density of ice is 57.5 lb / ft3 (920 kg / m3), for the same stored cooling capacity , the storage v olume for ice is only about 0.12 that of the chilled water. In addition, ice storage systems generally provide chilled water at a temperature of 34 to 35 °F (1.1 to 1.7 °C) to produce cold supply air between 42 and 49 °F (5.6 and 9.4 °C). Ice storage systems incorporating cold air distrib ution signi f cantly reduce the volume f ow rate of supply air , so air -side f an ener gy consumption and initial in vestment drop accordingly. Generally, ice storage systems ha ve a rather lo w incremental capital cost compared with conventional air conditioning systems without thermal storage. Ice storage systems can be easily incorporated with cold air distrib ution. A lower incremental capital cost than that of chilled w ater storage systems and the incorporation with cold air distrib ution are tw o main bene f ts of the ice storage system. In addition to the ice storage and chilled w ater storage systems, phase-change material storage systems are sometimes used. The most common phase-change material used for cool thermal storage is a mixture of inor ganic salts, water, and nucleating and stabilizing agents which melts and freezes at 47 °F (8.3°C). Phase-change materials ha ve high discharge temperatures and a high storage tank v olume of 6.0 ft 3 /ton h (0.048 m 3 /kWh) instead of 2.4 to 3.3 ft / ton h (0.019 to 0.027 m3 /kWh) for other ice storage systems; therefore, they ha ve limited applications in commercial buildings. Currently, ice storage and chilled water storage systems can be classif ed into the following categories: ●

●

●

●

●

Ice-on-coil, internal-melt ice storage system (IMISS) Ice-on-coil, external-melt ice storage system (EMISS) Encapsulated ice storage system (EISS) Ice-harvesting ice storage system (IHISS) Stratif ed chilled water storage system (SCWSS)

Chilled water storage systems are often lar ge-capacity storage systems. In ne wly installed chilled water storage systems, the stratif ed chilled water storage system is the most widely used.

31.6

CHAPTER THIRTY-ONE

31.2 ICE-ON-COIL, INTERNAL-MELT ICE STORAGE SYSTEMS System Description An ice-on-coil, internal-melt ice storage system uses brine f owing inside coils to make ice and to melt ice in the water that surrounds the coil. The central plant (cooling) and a chilled w ater or brine-incorporated ice storage system of an ice-on-coil, internal-melt ice storage system consists of the following components: chillers, ice storage tanks, chiller pumps, building pumps, controls, piping, and f ttings as well as AHUs, terminals, return air system, smoke control systems, and mechanical exhaust systems. Centrifugal, screw, and reciprocating chillers are usually used in ice-on-coil internal-melt ice storage systems depending on the size of the plant and types of condenser (w ater-cooled, aircooled, or evaporatively cooled) used. In locations where the outdoor air temperature during nighttime off-peak hours drops 20 °F (11.3°C) lower than the daytime maximum temperature, air-cooled chillers may sometimes be more eff cient than water-cooled chillers. Figure 31.2 is a schematic diagram of a typical ice-on-coil, internal-melt ice storage system for an off ce building near Dallas, Texas. Brine and Glycol Solution Brine is a salt solution or an aqueous glycol solution used as a heat-transfer medium. Its freezing point is lo wer than that of w ater, and depends on the concentration of salt or glycol in solution.

FIGURE 31.2 Schematic diagram of ice-on-coil, internal-melt ice storage system.

AIR CONDITIONING SYSTEMS: THERMAL STORAGE SYSTEMS

31.7

Brine is also used as a liquid coolant to absorb or to store heat ener gy in refrigeration and thermal storage systems. Ethylene glycol and prop ylene glycol are brines. They are colorless, nearly odorless liquids. They are often mixed with water at various concentrations and used as freezing point depressants to lower the freezing point of w ater. Inhibitors must be added to ethylene and prop ylene glycols to prevent metal corrosion. The freezing point of an aqueous ethylene glycol solution with a concentration of 25 percent by mass drops to 10°F (12.2°C), and its rate of heat transfer is about 5 percent less than that of water. The freezing point of a prop ylene glycol solution with a concentration of 25 percent by mass drops to 15°F (9.4°C). The physical properties of aqueous ethylene glycol solution are more appropriate for thermal storage systems than those of aqueous prop ylene glycol solution. In certain applications, toxicity considerations may be dictated by federal EPA requirements and local codes and regulations.

Ice Storage Tank In an ice storage system, ice making or char ging is the process in which compressors are used to produce ice. Ice burning (ice melting or discharging) means that ice in the storage system is melted in order to produce chilled water to offset the required refrigeration load. In an ice-on-coil, internal-melt ice storage system, ice is produced, or char ged, in multiple storage tanks where closely spaced multicircuited polyethylene or plastic tubes are surrounded by water, as shown in Fig. 31.3. Brine, an aqueous ethylene glycol solution with 25 to 30 percent ethylene glycol and 70 to 75 percent w ater, circulates inside the tubes at about 24 °F ( 4.4°C). The water surrounding the tubes freezes into ice up to a thickness of about 0.5 in. (12.7 mm). Tubes containing glycol solution entering and leaving the tank are arranged side by side alternately to provide more uniform heat transfer. Brine typically leaves the storage tank at 30 °F ( 1.1°C). Plastic tubes occup y about one-tenth of the tank volume, and another one-tenth is left empty to accommodate the expansion of ice during ice making. Multiple ice storage tanks are al ways connected in parallel. During ice b urning or ice melting, brine returns from the cooling coils in the air -handling units at a temperature of 46 °F (7.8°C) or higher. It melts the ice on the outer surf ace of the tubes and is thus cooled to 34 to 36 °F (1.1 to 2.2°C). The brine is then pumped to the air-handling units to cool the air again. In the storage tank, ice is stored and the high-pressure brine inside the tubes is separated from the water, usually at atmospheric pressure, surrounding the tubes in the storage tank.

Case Study: Operating Modes of Ice-on-Coil Ice Storage System In a typical ice storage system using ice-on-coil, internal-melt storage tanks in a 550,000 ft2 (51,115 m2) of f ce b uilding near Dallas, Texas, as described by Tackett (1989), there are tw o centrifugal chillers. Ethylene glycol is used as the coolant. Each chiller has a refrigeration capacity of 568 tons (1997 kW) when it produces 34 °F (1.1 °C) brine at a po wer consumption of 0.77 kW / ton (COP 4.56). If the brine lea ves the chiller at 24 °F ( 4.4°C), the refrigeration capacity then drops to 425 tons (1494 kW) with a power consumption of 0.83 kW / ton (COP 4.24). A demand-limited partial-storage strate gy is used; i.e., one chiller is operated during on-peak hours, as shown in Fig. 31.1c. Meanwhile, ice is also burned during on-peak hours to reduce the demand charge. Ice is char ged during off-peak hours to reduce ener gy costs. The system uses 90 iceon-coil, internal-melt storage tanks. For summer cooling, the daily 24-h operating c ycle can be di vided into three periods: off-peak, direct cooling, and on-peak. Off-Peak. This is the period from 8 p.m. until the air -handling units start the ne xt morning. During this period, the primary operating mode is ice making, at a maximum capacity of 650 tons

31.8

CHAPTER THIRTY-ONE

FIGURE 31.3 Ice-on-coil, internal-melt ice storage tank.

(2285 kW). The chillers also provide direct cooling at a capacity less than 200 tons for refrigeration loads that operate 24 h. In this operating mode, the ice-on-coil, internal-melt storage tanks are char ged. At the same time, a 34 °F (1.1 °C) ethylene glycol solution is supplied to the air -handling units for nighttime cooling. The DDC controller controls the ice storage system in the following operating sequence, as shown in Fig. 31.2: 1. Open control valves CV-1, CV-2, CV-5, CV-6, CV-8, and CV-11; and close control valves CV-3, CV-4, CV-6, CV-7, CV-9, and CV-10. 2. Reset the temperature of the glycol solution leaving the chiller Tel to 22°F ( 5.6°C). 3. Reset the load limit of both chillers to 100 percent. 4. Start the chiller and condenser pumps. Chiller pumps operate at high speeds during ice making to provide a higher f ow rate and a high rate of heat transfer in the storage tanks as well as a greater head to o vercome the pressure drop for both the e vaporator and the coils in the ice storage tanks. Chiller pumps operate at low speeds in direct cooling mode. 5. Start chillers 1 and 2 following the lead / lag sequence. 6. After chillers are started, open control valves CV-3 and CV-7. 7. Start the building chilled water circulating pumps in sequence.

AIR CONDITIONING SYSTEMS: THERMAL STORAGE SYSTEMS

31.9

8. Modulate control v alves 4 and 5, and maintain a 34 °F (1.1°C) glycol solution supply temperature to the air-handling units. If the ice-on-coil, internal-melt ice storage tanks are all 100 percent char ged, the ice storage capacity is 7500 ton h (26,370 kWh). When the sensors detect that the ice storage tanks are 100 percent charged, the ice-making mode is terminated. If nighttime after-hours cooling is not needed, the ice storage system shuts down. If the ice inventory (the amount of stored ice in the tanks) falls below 90 percent, ice making starts again. There are two additional operating modes during this period: ice making without direct cooling for after-hours use and ice burning for after-hours use with all chillers shut off. Direct Cooling. Direct-cooling operation lasts from the start of the air -handling units until noon on weekdays. This period has two operating modes: ●

●

Direct cooling mode. In this operating mode, chillers are operating and are reset to 34 °F (1.1°C). Direct cooling with ice-b urning or ice-melting mode . In this mode, both chillers are turned on. When the required refrigeration load exceeds both chillers’ capacity, some ice storage will be discharged to supplement the chillers.

On-Peak. On-peak hours are from noon until 8 p.m. weekdays. Ice-b urning mode, with or without chiller operation, is used in this period. In ice-b urning or ice-melting mode, one chiller is operated at the demand limit. This is the primary operating mode during summer cooling. The operating sequence is as follows: 1. Open control valves CV-3, CV-6, CV-7, and CV-10; and close CV-8, CV-9, and CV-11. 2. Open control v alve CV-1 and close CV -2 if chiller 1 is required to operate. Open CV -2 and close CV-1 if chiller 2 is required to operate. 3. Modulate control valves 4 and 5 at normal open positions. 4. Reset chilled water temperature leaving the chiller to 32°F (0°C). 5. Set the load limit of the operating chiller to 400 kW. 6. Start one condenser pump. 7. Start chiller pumps 1 and 2 at low speed. Both pumps will operate during ice burning. 8. Start one chiller according to the lead / lag sequence. 9. Modulate control valves 4, 5, 6, and 7 to maintain a 34°F (1.1°C) chilled water supply temperature to the air-handling units. 10. Start the brine circulating pumps in sequence. During on-peak hours, the brine circulating pump needs a greater head to o vercome the pressure drop of the coil in the AHU as well as the pressure drop of the coil in the ice storage tanks.

System Characteristics An ice-on-coil, internal-melt ice storage system is a modular system consisting of man y closely packed storage tanks connected in parallel. Such an ice storage system is more f exible during the installation of storage tanks, especially for retrof t projects. In ice-on-coil, internal-melt ice storage systems, off-peak cooling of the b uilding can be provided by direct cooling from the chillers. During ice b urning, melted water separates the tube and ice. Water has a much lower thermal conductivity (0.35 Btu / h ft °F, or 0.61 W / m °C) than that of ice (1.3 Btu / h  ft °F, or 2.25 W / m°C), so the capacity of the ice-on-coil, internal-melt ice storage system is dominated by the rate of ice burning or melting. Other system characteristics of ice-on-coil, internal-melt ice storage systems are listed in Table 31.1.

31.10

CHAPTER THIRTY-ONE

TABLE 31.1 System Characteristics of Ice-on-Coil, Internal-Melt Ice Storage Systems IMISS / EMISS, EISS

IHISS

SCWSS

Zone thermal and sound control Control zone Conrol methods Control modes Sound control

Single or multizone DDC or electric PI, PID, or on / off NC 25 – 40

Single or multizone DDC or electric PI, PID, or on / off NC 25 – 40

Indoor air quality Minimum ventilation control Filter eff ciency Humidity control

DCV or MPC Medium- to highOptional

DCV or MPC Medium- to highOptional

VAV air mixing or CV Centrifugal 4.5 – 5.5 in. WC 55% Fixed or differential dry bulb or enthalpy Centrifugal or axial Relief 0.6 in. WC, return 0.5 – 1 in. WC

VAV air mixing or CV Centrifugal 4.5 – 5.5 in. WC 55% Fixed or differential dry bulb or enthalpy Centrifugal or axial Relief 0.6 in. WC, return 0.5 – 1 in. WC

35 – 40%

35 – 40%

Centrifugal or screw HCFC-123, HCFC-22, HFC-134a Flooded liquid cooler

Centrifugal or screw HCFC-123, HCFC-22, HFC-134a IHISS: harvester; SCWISS: f ooded liquid cooler Water or evaporatively cooled

Air system Types Supply fan Supply fan total pressure Combined fan-motor-drive eff ciency Air economizer Return / relief fan Relief / return fan total pressure Combined fan-motor-drive eff ciency Cooling systems Refrigeration compressor Refrigerants Evaporator Condenser Energy performance Water pumps Combined pump-motor eff ciency Heating systems Boiler AFUE Maintenance Fault detection and diagnostics

Water or evaporatively cooled IMISS: 0.85 – 1.2 kW / ton, EMISS: 0.85 – 1.4 kW / ton, EISS: 0.85 – 1.2 kW / ton Centrifugal 65% Hot water coil or electric heater Conventional 78%, condensing 93%

Conventional 78%, condensing 93%

AHU, chiller

AHU, chiller

IHISS: 0.95 – 1.3 kW / ton, SCWSS: FOM 0.85 to 0.92 Centrifugal 65% Hot water coil or electric heater

FOM indicates f gure of merit. Refer to Sec. 31.7 for details.

31.3 ICE-ON-COIL, EXTERNAL-MELT ICE STORAGE SYSTEMS System Description In an ice-on-coil, external-melt ice storage system, ice b uilds up on the outer surf ace of coils or tube banks, which are submer ged in w ater in a storage tank. The refrigerant f ows and e vaporates inside the tubes. When the ice melts, it cools the w ater at a temperature between 34 and 38 °F (1.1 and 3.3°C) for cooling in the AHUs. The central plant (cooling) and chilled w ater incorporated ice storage system of an ice-on-coil, external-melt system consists of chillers, evaporating coils, stor-

AIR CONDITIONING SYSTEMS: THERMAL STORAGE SYSTEMS

31.11

age tanks, condenser, heat exchanger, refrigerant pumps, chilled water pumps, air system controls, piping, and f ttings. A typical ice-on-coil, external-melt ice storage system is shown in Fig. 31.4. Screw and centrifugal compressors are often used because of their higher ef f ciency. For an iceon-coil, external-melt ice storage system with a capacity less than 2400 ton  h (8438 kWh), a reciprocating compressor may also be used. Ev aporatively cooled condensers ha ve a higher system energy eff ciency ratio (EER) and are often used in many new projects. Ice Builders Ice builders are large, well-insulated steel tanks containing man y serpentine coils, usually made of steel pipes of 1- to 114 -in. (25- to 31-mm) diameter. HCFC-22 is often used as the refrigerant. The refrigerant-f lled serpentine coils are submerged in water in the ice builder and function as evaporators. The ice build up on the coil is between 1 and 2.5 in. (25 and 64 mm) thick. When ice builds up on the coil, the suction temperature of the compressor falls to 22 to 24°F ( 5.6 to  4.5°C). Ice is melted by the w ater circulating o ver it. The steel tubes of the serpentine coil should be spaced so that the b uilt-up ice c ylinders do not bridge each other . If the c ylinders are bridged, the paths of w ater circulation are block ed. Baff e plates are sometimes added to guide the w ater f ow and provide a secondary heat-transfer surface between the refrigerant and water. The storage tanks containing refrigerant coils are usually located at a lo wer level or on a grade because of their weight. Because the chilled w ater system in a multistory b uilding is always under a static head at lower levels, a heat exchanger, usually plate-and-frame type, is used to isolate the storage tank brine system from the chilled water system connected to the AHUs. An alternative is to supply chilled water directly to the storage tanks and pressurize the tanks. This arrangement obviates the use of a heat exchanger and a corresponding increase in brine temperature of about 3°F (1.7°C). Refrigerant Feed Two kinds of refrigerant feed are widely used in ice-on-coil, external-melt ice storage systems: direct expansion and liquid overfeed. Direct expansion (DX) uses the pressure difference between the receiver at the high-pressure side and the suction pressure to force the refrigerant to f ow through the ice b uilder. A suction-liquid heat e xchanger can be used to cool the liquid refrigerant from the condenser for better ef f ciency. Direct e xpansion is simple, and no refrigeration pump is required. Its main drawback is that 15 to 20 percent of the coil surf ace is used for superheat and is not a vailable for ice buildup. As described in Secs. 10.1 and 10.4, liquid overfeed uses a refrigerant pump to feed ice-b uilder coils about 3 times the evaporation rate they need. Because the liquid refrigerant wets the inner surface of the ice-builder coils, it has a higher heat-transfer coeff cient than direct expansion. Ice-Charging Control The thicker the ice built up on the coils, the greater the amount of ice stored in the tank. The thickness of ice on the coil should be measured to meet ice-b urning requirements during on-peak operating hours or in the direct cooling period. Because ice has a higher v olume than w ater, as the ice builder is char ged (i.e., as ice b uilds up on the coil), the w ater level rises. An electric probe can sense the water level in the tank and thereby determine the amount of ice stored in the tank. System Characteristics The ice-on-coil, external-melt ice storage system is the oldest type of ice storage system. It may be costly and complex. In the storage tank, stored ice occupies only about one-half the v olume of the tank, so the ice builder must be larger and heavier.

CHAPTER THIRTY-ONE

31.12

Water

Refrigerant Water level

CWS CWR Condenser

Ice builder

Storage tank

Suction-liquid heat exchanger

Plate-and-frame heat exchanger (a)

Compressor

460 Ice burning

Refrigeration load, tons

400

300

200

Ice burning and direct cooling

Ice making

137.5 100

0

4

8

12 Noon

Off-peak

On-peak

16

20 Off-peak

(b) FIGURE 31.4 A typical ice-on-coil, external-melt ice storage system: (a) schematic diagram; (b) operating diagram.

24 Hours

AIR CONDITIONING SYSTEMS: THERMAL STORAGE SYSTEMS

31.13

Case Study: An Ice-on-Coil, External-Melt Ice Storage System Gilbertson and Jandu (1984) presented an ice-on-coil, external-melt ice storage system for a 24story off ce building with an area of 265,000 ft 2 (24,628 m2) in San Francisco, California. This project was completed in the early 1980s. A schematic diagram of this system is sho wn in Fig. 31.4 a. This off ce tower had a peak refrigeration load of 460 tons (1617 kW), and an ice storage capacity of 3300 ton h (11,603 kWh) of refrigeration over 24 h. It required a 137.5-ton (4843-kW) refrigeration capacity for the ice storage system for a hot summer day at summer outdoor design conditions. For redundancy, two identical refrigeration systems using HCFC-22 as refrigerant were installed. Each had a 70-nominal-ton (246-kW) compressor, an ice builder of 960 ton h (3375 kWh), a watercooled condenser, and other accessories. (Nominal ton is the refrigeration capacity , in tons or kW , at rating conditions.) Direct-e xpansion refrigerant coils with a suction-liquid heat e xchanger were used. When the thickness of b uilt-up ice was about 2 in. (50 mm), the ice storage capacity reached 960 tonh (3375 kWh). A plate-and-frame heat e xchanger w as used to isolate the storage tank from the to wer chilled water system. The 34°F (1.1°C) water from the ice b uilder entering the heat e xchanger with an 8°F (4.4°C) temperature increase cooled the chilled w ater from 54 to 38 °F (12.2 to 3.3 °C). This was a partial-storage system. The operating modes during a 24-h c ycle for this 24-story of f ce to wer, shown in Fig. 31.4b, are as follows: Time period

Operating mode

6.00 p.m. to 8.00 a.m. 8.00 a.m. to 9.00 a.m. 9.00 a.m. to 6.00 p.m.

Off-peak, ice making On-peak, ice making and direct cooling On-peak, ice melting and direct cooling

The HVAC&R and plumbing system, including ice storage, for this of f ce building cost $2.4 million, which was slightly less expensive than a conventional central system using centrifugal chillers. System characteristics of ice-on-coil, external-melt ice storage systems are listed in Table 31.1.

31.4 ENCAPSULATED ICE STORAGE SYSTEMS System Description In an encapsulated ice storage system, plastic containers, f lled with deionized water and ice-nucleating agent, are immersed in a secondary coolant ethylene glycol solution in a steel or concrete tank. Ice is charged and stored when the secondary coolant is at a temperature between 22 and 26 °F ( 5.6 and  3.3°C) circulated through the tank. Ice is melted when the w arm coolant returned from the AHUs is circulated through the tank. Chillers can also pro vide direct cooling at a coolant temperature from 36 to 42°F (2.2 to 5.6°C). The central plant (cooling) and the chilled water incorporated ice storage system of an encapsulated ice storage system consist of the follo wing components: chillers, steel tank, encapsulated containers, pumps, air system controls, piping, and accessories. Two types of encapsulated ice containers are currently a vailable in the United States: dimpled spheres of 4-in. (100-mm) diameter and rectangular containers approximately 138 by 12 by 30 in. (35 by 300 by 750 mm). The containers are made of high-density polyethylene and are designed to withstand the pressure due to the expansion during freezing. When containers are put or stack ed inside the storage tank, they allow free circulation of f uid and do not provide unwanted short-circuit f uid f ow which causes degradation of performance. The storage tank can be an open, nonpressurized type or pressurized type. An open storage tank needs a barrier to keep the frozen containers submerged into the coolant.

31.14

CHAPTER THIRTY-ONE

Location of Chiller and Storage Tank In encapsulated ice storage systems and ice-on-coil, internal-melt ice storage systems, the chillers and storage tanks are usually connected in series, as shown in Fig. 31.5. When partial storage is used, their relative location can be either in chiller upstream or chiller downstream arrangements. In a chiller upstream arrangement, as shown in Fig. 31.5 a, the chilled w ater returned from AHUs at 46°F (7.8°C) is often f rst cooled in the chiller to 40 °F (4.4°C), and then it enters the storage tank and is cooled do wn to 34 °F (1.1 °C). In the chiller upstream arrangement, since the chilled w ater cooled at the chiller is at a higher temperature, this results in a higher COP at the chiller . However, the usable portion of the total storage capacity will be reduced because of the lo wer storage tank discharge temperature. In a chiller downstream arrangement, as shown in Fig. 31.5b, the chilled water returned from the AHUs at 46°F (7.8°C) is often f rst cooled in the storage tank to 40°F (4.4°C), and then it enters the chiller and is cooled do wn to 34 °F (1.1°C). In a chiller do wnstream arrangement, the COP of the chiller is lo wer, and the usable portion of the total storage capacity of the ice storage tanks is increased. Usually, during partial storage, the chiller upstream arrangement is often used for higher ef f ciency in the chiller . When ice storage capacity becomes a problem, an analysis should be undertaken to make an optimum choice.

Controls According to Dor gan and Elleson (1993), ice-charging inventory in the storage tank is measured based on the displacement of w ater in the tank when the ice is formed inside the encapsulated

Chiller

AHU

Tank

AHU

Tank

Tank

Tank

Tank

Tank Chiller (a)

(b)

FIGURE 31.5 Relative location of chiller and storage tank: (a) chiller upstream; (b) chiller downstream.

AIR CONDITIONING SYSTEMS: THERMAL STORAGE SYSTEMS

31.15

containers. For open tanks, a static pressure transducer is often used to detect the w ater level in the storage tank. In pressurized tanks, the e xpansion of the frozen containers forces the secondary coolant overf owing into a separate inventory tank, and its water level is measured. A pump is used to pump the o verf owing f uid in the in ventory tank back into the storage tank after dischar ging. With nondimpled spherical containers that e xpand very little as the encapsulated ice freezes, storage inventory can be monitored based on the integrated f ow and temperature measurements. Encapsulated ice storage systems use a storage tank bypass three-w ay modulating valve to control the chilled water leaving temperature. Encapsulated ice storage systems with a chiller upstream arrangement are well suited to chiller priority control. When the system refrigeration load is less than the chiller capacity , the chilled water bypasses the storage tanks completely . As soon as the system refrigeration load e xceeds the chiller capacity and the chiller leaving temperature increases above the leaving set point, the control system diverts part of the chilled w ater f ow through the storage tanks to maintain the required supply temperature to the AHUs. Storage priority is more complicated to achie ve. A required refrigeration load prediction algorithm to forecast the chiller cooling is needed each day . Chiller capacity is then limited by increasing the chilled water leaving setpoint, and most of or all the refrigeration load is then met by the ice storage. Generally, chillers should be controlled at full load during char ging to pre vent the reduction of system eff ciency and incomplete char ging of ice storage. The chiller lea ving temperature setpoint should be set at or belo w the minimum required char ging temperature so that the chiller is fully loaded throughout the charging cycle.

Charging and Discharging For encapsulated ice storage systems, the charging temperature decreases during the charge cycle as the thickness of ice through which heat is transferred increases. Arnold (1991) reported that encapsulated containers are subject to supercooling, i.e., cooling of the liquid w ater inside the container below its freezing point prior to the ice formation. Supercooling occurs only in fully dischar ged containers and results in a reduced rate of heat transfer at the beginning of the charging process. Supercooling can be signif cantly reduced by the addition of nucleating agents. For an entering chilled water temperature at the beginning of charging of 32°F (0°C) and a chilled water temperature at the end of charging of 20 to 26°F ( 6.7 to  3.3°C), a typical range of charging temperatures is between 4 and 12 °F (2.2 and 6.7 °C) corresponding to a char ging period of 8 to 16 h. Encapsulated ice storage systems have a steadily falling discharge rate when the discharge temperature is k ept constant, or a steadily rising dischar ge temperature when the dischar ge rate is constant. This is due to the decreasing area of ice in contact with the container as the ice melts. The encapsulated ice storage discharge temperature typically begins at 32°F (0°C) and ends at a discharge temperature of 38 to 45°F (3.3 to 7.2°C) with a discharge temperature range of 6 to 13°F (3.3 to 7.2°C).

System Characteristics System characteristics of encapsulated ice storage systems are listed in Table 31.1.

31.5 ICE-HARVESTING ICE STORAGE SYSTEMS System Description In an ice-harv esting ice storage system, ice is produced in a harv ester, which is separate from the storage tank where ice is stored. The e vaporator of the chiller is a v ertical plate heat e xchanger mounted abo ve a w ater / ice storage tank. Lo w-pressure liquid refrigerant is forced through the

31.16

CHAPTER THIRTY-ONE

FIGURE 31.6 Schematic diagram of a typical ice-harvesting ice storage system.

hollow inner part of the plate heat e xchanger, in which liquid refrigerant is vaporized, and produces a refrigeration effect. The central plant (cooling) and brine incorporated ice storage system in an ice-harv esting ice storage system consists of the following equipment and main components: chillers, an ice harvester, storage tank, air system controls, piping, and accessories, as shown in Fig. 31.6. Ice Making or Charging During ice-making or -charging mode, a chilled aqueous ethylene glycol solution with a concentration of 25 to 30 percent is pumped from the storage tank and distrib uted over the outer surf ace of the evaporator plates at a temperature equal to or slightly above 32°F (0°C). It then f ows downward along the outer surface of the plate in a thin f lm. Water is cooled and then frozen into ice sheets approximately 0.2 to 0.3 in. (5 to 7.5 mm) thick. Periodically , hot gas is introduced into one-fourth of the evaporator plates by re versing the refrigerant f ow. Ice is harv ested, or released from the outer surface of the plates, in the form of f akes or chunks and f alls into the storage tank belo w. Ice is formed in 20 to 30 min and is harv ested within 20 to 40 s. During harv esting, this section of plate evaporator acts as a condenser. Ice accumulates in the storage tank to occup y slightly less than 60 percent of the v olume of the tank. Because the ice f akes are usually smaller than 6 in. by 6 in. by 0.25 in. (1500 mm by 1500 mm by 63 mm), there is a large contact area between the return brine from the cooling coils and the ice. The time required to melt the ice in the storage tank is less than one-tenth of the time needed in ice making or charging. For a reciprocating compressor using an e vaporative condenser, the power consumption of the chiller during ice making is about 0.95 to 1.1 kW / ton (COP 3.7 to 3.22). Because the e vaporator

AIR CONDITIONING SYSTEMS: THERMAL STORAGE SYSTEMS

31.17

plates must be located abo ve the storage tank, ice-harvesting systems need more headroom than other ice storage systems.

Chiller Operation During off-peak hours, an ice-harvesting system can also be used to lo wer the temperature of brine returning from the AHUs. Brine at a temperature of 34 °F (1.1 °C) is supplied to the air -handling units to cool the air to a supply temperature of 42 to 45 °F (5.6 to 7.2°C) during direct cooling. It is then returned to the ice harv ester at 50 to 60 °F (10 to 15.6 °C) and distrib uted over the e vaporator plates directly. After f alling from the e vaporator plates, brine is again cooled to a temperature of 34°F (1.1°C) before it leaves the storage tank. Because of the higher temperature of return brine distrib uted over the evaporator plates, the capacity of the ice harv ester increases, and its power consumption decreases during chiller operation. These changes in capacity and po wer consumption depend mainly on the temperature of return brine distributed over the e vaporator plates. During chiller operation, power consumption usually varies between 0.75 and 0.85 kW / ton (COP 4.69 and 4.14). Ice-harvesting ice storage systems are ef fective in ice-making and ice-b urning operations. The temperature of brine from the storage tank of the ice harv ester can be lo wered to 34 °F (1.1 °C), which is 2 °F (1.1°C) lower than in the ice-on-coil, internal-melt ice storage system. Ice-harv esting systems have been successfully used in load shifting and load le veling to reduce electric demand and energy cost. Ho wever, melting of the ice during the harv esting process not only decreases the amount of ice harv ested, but also adds an incremental refrigeration load to the system, which increases the power consumption during the ice-making process. An ice-harvesting ice storage system with brine system is an open system. More water treatment is required than in an ice-on-coil, internal-melt ice storage system whose brine system is a closed system.

System Characteristics System characteristics of an ice-harvesting ice storage systems are listed in Table 31.1.

31.6 COMPARISON OF VARIOUS ICE STORAGE SYSTEMS According to Dorgan and Elleson (1993), the discharge temperature, the energy consumption of the chillers, the roughly estimated installed cost of storage tanks and accessories, and the chiller cost without installation are as shown below:

Discharge temperature, °F Chiller energy use, kW / ton Storage tank volume, ft3 /ton  h Storage installed cost, $ / ton h Chiller cost, $ / ton

IMISS

EMISS

EISS

IHISS

34 – 38 0.85 – 1.2 2.4 – 2.8 50 – 70 200 – 500

34 – 36 0.85 – 1.4 2.8 50 – 70 200 – 500

34 – 38 0.85 – 1.2 2.4 – 2.8 50 – 70 200 – 500

34 – 36 0.95 – 1.3 3.0 – 3.3 20 – 30 1000 – 1500

According to Potter et al. (1995), a survey of 196 thermal storage systems w as undertaken. The users of the thermal storage systems were ask ed to assess the ice storage systems according to an index of 1 to 10. One indicated complete dissatisf action, whereas 10 meant complete satisf action. The ratings for the ice storage systems are as follows:

31.18

CHAPTER THIRTY-ONE

Ice-on-coil, internal-melt ice storage systems Ice-on-coil, external-melt ice storage systems Encapsulated ice storage systems Ice-harvesting ice storage systems

7.82 6.78 7.25 5.02

In addition to these systems, there is another ice storage system called an ice slurry ice storage system in which suspended ice crystals are formed. Ice slurry systems ha ve limited commercial application because of their higher costs.

31.7 STRATIFIED CHILLED WATER STORAGE SYSTEMS System Description A stratif ed chilled w ater storage system often uses a lar ge storage tank to store chilled w ater at a temperature between 40 and 45 °F (4.4 and 7.2 °C). The stored chilled w ater of fsets the b uilding refrigeration load during on-peak hours to shift the load to the of f-peak hours and reduces the energy cost. Chilled w ater in the storage tank is strati f ed into three re gions because of its gra vity: top warmer return w ater from the AHUs, middle region of steep temperature gradient, and bottom colder chilled water from the chillers. The central plant (cooling) and the chilled water incorporated storage system consist of chillers, a c ylindrical storage tank, pumps, piping, air system controls, and accessories.

Basics for Chilled Water Storage The stored cooling capacity of a chilled w ater storage system depends on the temperature dif ference between the w arm water return from the AHUs and the chilled w ater stored in the tank, and the amount of w ater stored. The larger the storage tank, the lower the capital cost per unit stored volume. According to Dor gan and Elleson (1993), a chilled w ater storage system is economical when its storage capacity e xceeds 2000 ton  h or 200,000 gal (7000 kWh or 760 m 3). Currently, chilled water systems achieve thermal separation between cold char ged water and warm return water by strati f cation, multiple tanks, membrane, diaphragm, and baf f es. The stratif ed tank is the simplest and most eff cient method. Chilled water storage systems need a storage tank volume of 11 3 to 21 ft 3 /ton h (0.089 to 0.169 m 3 /kWh) compared to 2.4 to 3.3 ft / ton  h (0.019 to 0.027 3 m /kWh) for ice storage systems. Charging and Discharging. Charging is the process of f lling the storage tank with chilled w ater from the chiller , usually at a temperature between 40 and 45 °F (4.4 and 7.2 °C). Meanwhile, the warmer return chilled w ater from the air -handling units or terminals, usually at a temperature between 55 and 60 °F (11.1 and 15.6 °C), is extracted from the storage tank and pumped to the chiller to be cooled. Discharging is the process of dischar ging the chilled w ater, at a temperature between 41 and 45°F (5.0 and 7.2 °C), from the storage tank to the air -handling units and terminals. At the same time, the warmer return chilled w ater from the coils f lls the tank by means of storage w ater pumps. Loss of Cooling Capacity during Storage. During the storage of chilled w ater, the follo wing processes result in losses in cooling capacity: ●

Stored chilled water is warmed by direct mixing of w armer return chilled water and stored colder chilled water.

AIR CONDITIONING SYSTEMS: THERMAL STORAGE SYSTEMS ●

●

31.19

Heat from prviously stored w armer return chilled w ater is transferred from the w armer tank wall to the stored chilled water. Heat is transferred through the tank wall from the warmer ambient air.

Figure of Merit. A more easily measured, enthalpy-based figu e of merit (FOM) is often used to indicate the loss of cooling capacity of the stored chilled w ater during the charging and discharging processes in a complete storage cycle. The FOM is def ned as FOM 

m˙wcpw(Trc  To) Q dis  Q ch m˙wcpw(Trc,m  Ti)

(31.1)

where Qdis  cooling capacity available during discharge process, Btu / h (W) Qch  theoretical cooling capacity available during charging process, Btu / h (W) m˙w, m˙w  mass f ow rate and summation of mass f ow rate of water, lb / h (kg / s) cpw  specif c heat of water, Btu / lb  °F (J / kg °C) Trc  warmer return chilled water temperature f lling storage tank during discharge process, °F (°C) To  outlet temperature of stored chilled water, °F (°C) Trc,m  mass-weighted average temperature of return chilled water at inlet during discharge process, °F (°C) Ti  inlet temperature of stored chilled water during charging process, °F (°C) The smaller the losses of cooling capacity during chilled w ater storage, the greater the value of FOM.

Storage Tanks Chilled w ater storage tanks are usually f at-bottomed v ertical c ylinders. A c ylindrical tank has a lower surf ace-to-volume ratio than a rectangular tank. Lar ge c ylindrical tanks typically ha ve a height-to-diameter ratio of 0.25 to 0.35. Steel is the commonly used material for abo ve-grade tanks, and concrete is widely used for under ground tanks. In certain projects, precast, prestressed, cylindrical concrete tanks with encla ved w atertight steel diaphragms are used for lar ge chilled w ater storage facilities with a v olume over 2.5 million gal (9463 m 3). All outdoor above-grade structures should have a 2-in.- (50-mm-) thick e xternal insulation layer spray-on polyurethane foam, a vapor barrier, and a highly ref ective top coating. Stratified Tanks Stratif ed tanks rely on the b uoyancy of w armer return chilled w ater, which is lighter than colder chilled water, to separate these tw o chilled w aters during char ging and dischar ging. Diffusers are used to lo wer entering and lea ving w ater v elocity to pre vent mixing. In a strati f ed tank, colder stored chilled water is always charged from the bottom dif fusers arranged concentrically, as shown in Fig. 31.7a. It is also discharged from the same bottom diffusers. The warmer return chilled w ater is introduced to and withdra wn from the tank through the top lateral diffusers. According to f eld measurements, stratif ed tanks have a f gure of merit between 0.85 and 0.92. Tran et al. (1989) showed that there is no signif cant difference in FOM between stratif ed tanks and membrane tanks or empty tanks. A membrane tank is a storage tank in which a membrane separates the colder stored chilled w ater and w armer return w ater. An empty tank is a storage tank in which walls are used to separate the colder and warmer chilled water. Compared with membrane tanks and empty tanks, stratif ed tanks ha ve the adv antages of simpler construction and control, greater storage capacity , and lower cost. Strati f ed tanks are widely used in chilled water storage installations.

31.20

CHAPTER THIRTY-ONE

FIGURE 31.7 Double-octagon diffuser for a c ylindrical stratif ed tank: (a) plan view of bottom dif fusers; (b) 12-in. diameter PVC pipe diffuser.

Temperature Gradient and Thermocline Vertical temperature prof les are formed during charging or discharging in stratif ed tanks at various time intervals. Temperature prof les may be illustrated on a height-temperature (H-T) diagram at the beginning, the middle, and near the end of the char ging process, as shown in Fig. 31.8. In the middle of the charging process along the vertical height of the storage tank, chilled water is divided into three re gions: bottom colder -and-heavier stored chilled w ater, thermocline, and top w armer-andlighter return chilled water. A thermocline is a strati f ed region in which there is a steep temperature gradient. The water temperature often v aries from 42 to 60 °F (5.6 to 15.6 °C). The thermocline separates the colder stored chilled water from the w armer return chilled w ater. The thinner the thermocline, the smaller the mixing loss.

Diffusers The layout and con f guration of diffusers in a strati f ed tank have a signif cant effect upon the mixing of the colder and warmer chilled water, as well as on the formation of the thermocline. The purpose of the dif fusers and their connecting piping is to distrib ute the incoming chilled w ater evenly, so that it f ows through the inlet openings with suff ciently low velocity (usually lower than 0.9 ft / s, or 0.27 m / s) to minimize mixing of colder and w armer chilled water. Wildin (1990) recommended that diffuser design and installation take into account the following: ●

●

Warmer return chilled w ater should be introduced as closely as possible to the top w ater level in the stratif ed tank. Colder stored chilled water should be introduced just above the bottom f oor of the tank. The inlet temperature of chilled w ater should be controlled within a narro w band (  2°F or  1.1°C) during charging to avoid additional mixing.

AIR CONDITIONING SYSTEMS: THERMAL STORAGE SYSTEMS

31.21

FIGURE 31.8 Temperature prof le and thermocline in a stratif ed tank.

●

●

●

●

●

●

Obstructions in the f ow crossing the tank, other than diffusers and the connecting piping, should be minimized. The primary function of diffusers is to reduce mixing. Mixing can occur at tw o points: at the start of the char ging and dischar ging processes during the formation and reformation of the thermocline, and at the inlet side of the thermocline after the thermocline has been formed. Mixing near the inlet diffuser can be minimized if the incoming chilled w ater initially forms a thin layer of gra vity current that tra vels across the tank because of the gra vity dif ference, rather than inertia. Gra vity current slo wly pushes the chilled w ater originally in the tank out of the way so that mixing only occurs at the front of the gra vity current when it f irst crosses the tank. There are tw o kinds of dif fusers: linear diffusers and radial disk dif fusers. Large stratif ed tanks usually incorporate linear diffusers. Inlet f ow from the top dif fusers should be upw ard or horizontal. Bottom dif fusers should f ow downward and have slots spreading at 120°. The cross-sectional inlet area of the branch pipe leading to the diffuser should be at least equal the total area of the diffuser openings in that branch. Mixing on the inlet side of thermocline depends on the inlet Re ynolds number Re i and Froude number Fri. The inlet Reynolds number is closely related to the inlet velocity and is def ned as Rei  

V˙w l dif w 0.00223V˙gal l dif w

where vV˙w  olume f ow rate of chilled water, ft3 /s

(31.2)

31.22

CHAPTER THIRTY-ONE

V˙gal  volume f ow rate of chilled water, gpm ldif  linear length of diffuser, ft w  kinematic viscosity of water, ft2 /s According to Wildin (1990), when Re i 850, loss due to mixing and loss of cooling capacity during discharge can be signif cantly reduced. If Re i is determined, the length of the linear diffuser can be calculated from Eq. (31.2). The inlet Froude number Fri is def ned as Fri 

V˙w l dif[gh 3i (i  a) / i]0.5

(31.3)

where g  acceleration of gravity, ft / s2 (m / s2) i  density of inlet water, lb / ft3 (kg / m3) a  density of ambient water stored in tank at diffuser level, lb / ft3 (kg / m3) In Eq. (31.3), hi indicates the inlet opening height, ft (m). It is the v ertical distance occupied by the incoming f ow when it lea ves the dif fuser and forms the gra vity current. F or the bottom dif fusers, inlet opening height hi indicates the vertical distance between the tank f oor and the top of the opening of the diffuser. Self-balancing. Self-balancing means that the water f ow introduced to or extracted from the tank should be self-balanced according to requirements, i.e., evenly distrib uted at all f ow conditions. This includes the following requirements: ●

●

●

●

●

The piping design should be symmetric. Branch pipes should be equal in length. Flow splitters should be added at the appropriate points. Pipe diameter reduction should be combined with the f ow splitter. Long-radius elbows should be used.

Charging and Discharging Temperature versus Tank Volume The FOM of a strati f ed tank is closely related to the temperature dif ference of the outlet temperature of stored chilled w ater during dischar ging To and the inlet temperature of stored chilled w ater during charging Ti. For a complete charging and discharging cycle, the average To during discharging is al ways greater than the a verage Ti because of the mixing loss and heat gains, provided that the water f ow rate is constant. The smaller To  Ti, the higher the FOM. Figure 31.9 sho ws curves of chilled w ater temperature v ersus tank v olume during the char ging and discharging processes of a complete chilled w ater storage c ycle in a lar ge stratif ed tank. Inlet and outlet temperatures are measured at the openings of the top and bottom dif fusers. During the charging process, return chilled w ater is e xtracted from the top dif fusers of the strati f ed tank, cooled in the chiller, and charged into the stratif ed tank again through the bottom diffusers. The inlet temperature of the stored chilled w ater Ti gradually decreases as the stored v olume increases. This is due to a comparati vely lower rate of heat transfer to the inlet w ater from the w armer ambient water, piping, and tank wall after the beginning of the charging process. During the dischar ging process, stored chilled w ater is e xtracted from the strati f ed tank and supplied to the cooling coils in the air -handling units and terminals. The return chilled w ater is introduced to the stratif ed tank through the top dif fusers. The outlet temperature of the stored chilled water To gradually increases as the stored volume decreases. Both the outlet temperature of return chilled w ater during charging Trc,o and inlet temperature of return chilled w ater during dischar ging Trc.i should be controlled between 55 and 60 °F (12.8 and 15.6°C) so that stratif cation can be maintained in the storage tank.

AIR CONDITIONING SYSTEMS: THERMAL STORAGE SYSTEMS

31.23

FIGURE 31.9 Chilled w ater temperature v ersus tank v olume curv es during char ging and dischar ging processes.

System Characteristics System characteristics of stratif ed chilled water storage systems are listed in Table 31.1.

31.8 CASE STUDY: A STRATIFIED CHILLED WATER STORAGE SYSTEM A full-storage strati f ed chilled w ater storage system w as completed in August 1990 to serv e a 1.142 million ft 2 (106,134 m 2) electronics manuf acturing f acility in Dallas, Texas. The following are the details of this project as described by Fiorino (1991).

Chilled Water Storage System The 2.68-million-gal (10,144-m 3) stratif ed tank is a precast, prestressed, cylindrical concrete water tank. The design parameters of this stratif ed chilled-water storage system are as follows:

31.24

CHAPTER THIRTY-ONE

Storage cooling capacity Maximum refrigeration load Charge process duration Discharge process duration Inlet temperature during charging Ti Limiting outlet temperature during discharging To Inlet temperature during discharging Tre,i Maximum volume f ow rate Tank diameter Tank height Tank volume Usable tank volume

24,500 ton h (861,420 kWh) 3200 tons (11,251 kW) 16 h 8h 40°F (4.4°C) 42°F (5.6°C) 56°F (13.3°C) 5120 gpm (323 L / s) 105.5 ft (32.2 m) 41 ft (12.5 m) 2.68 million gal (10,144 m3) 90%

There are two 1200-ton (4220-kW) and tw o 900-ton (3164-kW) centrifugal chillers to serv e zone 1 and zone 2, as shown in Fig. 31.10. There are also f ve chiller pumps, one of which is a standby . Two variable-speed building pumps are installed for each zone, one of which is a standby pump. Three storage pumps are used for chilled water storage, one of which is standby.

Concentric Double-Octagon Diffusers Concentric double-octagon dif fusers are used at both the top and bottom of the strati f ed tank, as shown in Fig. 31.7a. Each octagon introduces 50 percent of the total v olume f ow during the charging process. Ov er time, a double-octagon arrangement pro vides nearly twice the ef fective linear length for diffusers of a single-octagon arrangement. The total effective length of the eight diffusers in the outer octagon is 559 ft (170 m). For a total f ow of 0.5  5120  2560 gpm (162 L / s) through the outer octagon, the volume f ow rate per linear ft (0.30 m) of diffuser is V˙dif 2560  0.1337   0.0102 ft 2/s l dif 559  60 The kinematic viscosity of water at 42°F is 1.66  10  5 ft2 /s, so the inlet Reynolds number is calculated as Rei 

V˙dif l dif w



0.0102

 615

1.66  10 5

Similarly, the inlet Reynolds number for inner octagon is 1068. Both are close to the upper limit of 805. The acceleration of gravity g  32.2 ft / s2, and

i  a i



62.4263  62.3864 62.3864

 0.00064

For a common inlet opening height hi  5.64 in, or 0.47 ft (0.14 m), the inlet Froude number for diffusers in the outer octagon is calculated as Fri 

0.0102 [(32.2)(0.47)3(0.00064)]0.5

 0.22

FIGURE 31.10 Texas.

Schematic diagram of a strati f ed chilled w ater storage system for an electronic f acility in Dallas,

31.25

31.26

CHAPTER THIRTY-ONE

Similarly, the inlet Froude number for diffusers in the inner octagon is 0.38. At a lateral slot spacing of 0.5 ft (0.15 m) in the inner octagon dif fusers and 0.87 ft (0.26 m) in the outer octagon diffusers, each linear diffuser has 32 lateral slots. If the maximum inlet velocity is 0.9 ft / s (0.27 m / s), the opening area for each lateral slot is 5120  0.1337 60(0.9)(32)(8  8)

 0.025 ft 2 (0.023m2)

Because the lateral slot is spread at an angle of 120 ° downward, if the length of the lateral slot is about 1 ft (0.30 m), the width of the slot is 0.025 / 1  0.025 ft, or about 0.3 in. (7.6 mm). If the cross-sectional area of each linear diffuser is equal to the slot openings, the diameter of the diffuser is therefore D

 4A 

0.5



4  0.025  32 (0.5)  1.01 ft (305 mm)

To provide even distribution of incoming water f ow, f ow splitters are used to divide the water f ow evenly to the split mains and branches.

Charging Process During full-load operation in hot weather , charging is performed from 8 p.m. until noon the ne xt day. All four chillers and all w ater pumps except the standby pumps are operated simultaneously to provide direct cooling during of f-peak hours as well as the required stored cooling capacity during on-peak hours the next day. The direct-cooling refrigeration load during off-peak hours varied from 1980 to 2600 tons (6962 to 9142 kW), with a total of 34,800 ton  h (122,357 kWh). The required stored cooling capacity to meet the refrigeration load during on-peak hours w as 21,300 ton h (74,891 kWh) on July 17, 1989. For a char ging process of 16 h, the four chillers are operated at an a verage refrigeration load of about 3600 tons (12,658 kW). The set point of the chilled w ater temperature lea ving the chiller is 39.5°F (4.2°C), and the outlet return chilled water temperature is around 56°F (13.3°C). Before charging, the chillers and water pumps are started. During the char ging process, the control valves CV-1 and CV-2 and solenoid valves SV-3, SV-5, SV-8, and SV-9 are opened, as shown in Fig. 31.10; and solenoid valves SV-4, SV-6, and SV-7 are closed. This provides both the charging of the storage tank and the supply of chilled water to building pumps during off-peak hours. Return chilled water at around 56°F (13.3°C) is extracted from the stratif ed tank through the top diffusers and the control v alve CV-1 and solenoid v alve SV-3 by storage pumps SP-1 and SP-2. It f ows through solenoid valves SV-5 and SV-9 and the chiller pumps, cools in the chiller, and leaves the chiller at 39.5 °F (4.2°C). After that, the chilled w ater is di vided into tw o streams. One of the streams f ows through solenoid v alve SV-8 and control v alve CV-2 and is char ged to the strati f ed tank through the bottom dif fusers at 40 °F (4.4 °C). The other stream is e xtracted by the b uilding pump and supplied to the air-handling units and terminals for direct cooling.

Discharging Process Before the shutoff of the chillers and chiller pumps, the control valves and solenoid v alves should be switched over to the followings: control valves CV-1 and CV-2 and solenoid valves SV-1, SV-2, SV-4, SV-6, and SV-7 should be open, and solenoid valves SV-3, SV-5, SV-8, and SV-9 should be closed. Stored chilled water at 41 °F(5°C) is extracted by storage pumps SP-2 and SP-3 via bottom diffusers and control valve CV-2. It then f ows through SV-6, is extracted again by the building pumps, and is supplied to the cooling coils in the air -handling units and terminals. Return chilled w ater at a

AIR CONDITIONING SYSTEMS: THERMAL STORAGE SYSTEMS

Off-peak

On-peak

Off-peak Electric demand with storage

9

Electric demand, MW

31.27

8

Drop in electric demand in on-peak hours

7

Electric demand without storage

6 5

Refrigeration load, tons

3600 3200

Refrigeration load

Charging

Charging

2800 Estimated refrigeration load curve

2400 Discharging 2000

0 Midnight

Direct cooling 2

4

6

8

Direct cooling 10

12 Noon

14

16

18

20

22

24 Midnight

FIGURE 31.11 Electric demand curv es and the storage c ycle of the strati f ed chilled w ater storage system for an electronic manufacturing facility in Dallas, Texas.

temperature Trc  56°F (13.3°C) is then forced through solenoid valve SV-7 and control valve CV-1 and introduced to the stratif ed tank through the top diffusers. Temperature Trc is maintained by changing the chilled water temperature supplied to the air-handling units from 45 to 52°F (7.2 to 11.1°C) instead of 41°F (5°C). As soon as temperature sensor T2 senses a return chilled w ater temperature Trc drop below 56°F (13.3°C), the DDC controller closes the control v alve CV-z slightly. Less stored chilled w ater is e xtracted by the b uilding pumps. The required amount of return chilled water at 56°F (13.3°C) bypasses the crossover and mixes with the 41°F (5°C) stored chilled w ater, which results in a higher supply temperature to the cooling coils. During on-peak hours, the discharging process requires a refrigeration load from 2500 to 2800 tons (8790 to 9845 kW), with a total of 21,250 ton h (74,715 kWh) for the 8-h on-peak period.

Part-Load Operation This electronic manufacturing facility includes clean rooms, computer rooms, compressed air aftercoolers, and manuf acturing equipment, all of which need 24-h continuous cooling. Its daytime refrigeration load averages about 1240 tons (4360 kW).

31.28

CHAPTER THIRTY-ONE

When the refrigeration load is reduced or the entering temperature of condenser w ater drops because of a lo wer wet-b ulb temperature, the follo wing adjustments are made during part-load operation: ●

●

The temperature of inlet w ater entering the strati f ed tank during the char ging process Ti is raised from 39.5° to 42.5°F (4.2 to 5.8°C), which increases the chillers’ capacity and lowers their power consumption. All four chillers and their auxiliary equipment do not ha ve to operate simultaneously. Instead of 8 p.m. daily, the start of the char ging process can be delayed until nearly all the stored chilled water in the strati f ed tank is dischar ged. The discharging process might last 10 to 14 h during part-load operation instead of 8 h in full-load operation.

System Performance According to the operating cycle extended from August 24 to August 26, 1990, the electric demand dropped about 2.5 MW during on-peak hours as intended because four chillers and their corresponding auxiliary equipment were shut do wn during on-peak hours, as shown in Fig. 31.11. The maximum storage capacity of the strati f ed tanks w as 27,643 ton  h (97,193 kWh). The difference between the average outlet temperature during discharging and the average inlet temperature during charging To  Ti was 1.1°F (0.61°C), and the f gure of merit was 92.2 percent.

REFERENCES Arnold, D., Laboratory Performance of an Encapsulated Ice-Storage, ASHRAE Transactions, 1991, Part II, pp. 1170 – 1178. ASHRAE, ASHRAE Handbook 1996, HVAC Systems and Equipment, ASHRAE Inc., Atlanta, GA, 1996. ASHRAE, ASHRAE Handbook 1997, Fundamentals, Atlanta, GA, 1997. Denkmann, J. L., Performance Analysis of a Brine-Based Ice Storage System, ASHRAE Transactions, 1985, Part I B, pp. 876 – 891. Dorgan, C. E., and J. S. Elleson, Design of Cold Air Distribution Systems with Ice Storage, ASHRAE Transactions, 1989, Part I, pp. 1317 – 1322. Dorgan, C. E., and Elleson, J. S., Design Guide for Cool Thermal Storage, ASHRAE Inc., Atlanta, 1993. Fields, W. G., and D. E. Knebel, Cost Effective Thermal Energy Storage, Heating / Piping / Air Conditioning, July 1991, pp. 59 – 72. Fiorino, D. P., Case Study of a Large, Naturally Stratif ed, Chilled-Water Thermal Storage System, ASHRAE Transactions, 1991, Part II, pp. 1161 – 1169. Gatley, D. P., Successful Thermal Storage, ASHRAE Transactions, 1985, Part I B, pp. 843 – 855. Gilberston, T. A., Ice Cools Off ce-Hotel Complex, Heating / Piping / Air Conditioning, August 1989, pp. 47 – 52. Gilbertson, T. A., and R. S. Jandu, 24-Story Off ce Tower Air Conditioning System Employing Ice Storage — A Case History, ASHRAE Transactions, 1984, Part I B, pp. 387 – 398. Harmon, J. J., and H. C. Yu, Design Consideration for Low-Temperature Air Distribution Systems, ASHRAE Transactions, 1989, Part I, pp. 1295 – 1299. Hittle, D. C., and T. R. Smith, Control Strategies and Energy Consumption for Ice Storage Systems Using Heat Recovery and Cold Air Distribution, ASHRAE Transactions, 1994, Part II, pp. 1221 – 1229. Lorsch, H. G., Air-Conditioning System Design Manual, ASHRAE Inc., Atlanta, GA, 1993. Lumpkin, R. M., Thermal Storage: A Reversible Process, HPAC, no. 1, 1998, pp. 136 – 142. MacCracken, C. D., Off-Peak Air Conditioning: A Major Energy Saver, ASHRAE Journal, December 1991, pp. 12 – 22. Pearson, F. J., Ice Storage Can Reduce the Construction Cost of Off ce Buildings, ASHRAE Transactions, 1989, Part I, pp. 1308 – 1316.

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31.29

Potter, R. A., D. P. Weitzer, D. J. King, and D. D. Boettner, ASHRAE RP-766: Study of Operational Experience with Thermal Storage Systems, ASHRAE Transactions, 1995, Part II, pp. 549 – 557. Schiess, K., RTP  TES  ?, Engineered Systems, no. 10, 1998, pp. 102 – 110. Silvetti, B., and M. MacCracken, Thermal Storage and Deregulation, ASHRAE Journal, no. 4, 1998, pp. 55 – 59. Sohn, C. W., and J. J. Tomlinson, Diurnal Ice Storage Cooling Systems, ASHRAE Transactions, 1989, Part I, pp. 1079 – 1085. Spethmann, D. H., Optimal Control for Cool Storage, ASHRAE Transactions, 1989, Part I, pp. 1189 – 1193. Stamm, R. H., Thermal Storage Systems, Heating / Piping / Air Conditioning, January 1985, pp. 133 – 151. Tackett, R. K., Case Study: Off ce Building Uses Ice Storage, Heat Recovery, and Cold Air Distribution, ASHRAE Transactions, 1989, Part I, pp. 1113 – 1121. Townsend, S. B., and J. G. Asbury, Cooling with Off-Peak Energy: Design Implications of Different Rate Schedules, ASHRAE Transactions, 1984, Part I B, pp. 360 – 373. Tran, N., J. E. Kreider, and P. Brothers, Field Measurement of Chilled Water Storage Thermal Performance, ASHRAE Transactions, 1989, Part I, pp. 1106 – 1112. The Trane Company, Ice Storage Systems 1987, Applications Engineering Manual, American Standard, Inc., La Crosse, WI, 1987. Trueman, C. S., Operating Experience with a Large Thermally Stratif ed Chilled-Water Storage Tank, ASHRAE Transactions, 1987, Part I, pp. 697 – 707. Wildin, M. W., Diffuser Design for Naturally Stratif ed Thermal Storage, ASHRAE Transactions, 1990, Part I, pp. 1094 – 1101. Williams, C. D., Optimizing TES Chiller Management, ASHRAE Journal, no. 4, 1996, pp. 43 – 48.

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COMMISSIONING AND MAINTENANCE 32.1 HVAC&R COMMISSIONING 32.1 Commissioning and Design Intent 32.1 Necessity of HVAC&R Commissioning 32.1 Scope of HVAC&R Commissioning 32.2 Testing, Adjusting, and Balancing (TAB) 32.3 HVAC&R Commissioning Team 32.4 When to Perform HVAC&R Commissioning 32.4

Cost of HVAC&R Commissioning 32.5 32.2 HVAC&R MAINTENANCE 32.5 Basics 32.5 Maintenance Contractors and Maintenance Personnel 32.5 Monitoring and Fault Detection and Diagnostics Assisting Predictive Maintenance 32.6 REFERENCES 32.6

32.1 HVAC&R COMMISSIONING Commissioning and Design Intent According to Wilkinson (1999a, 1999b), commissioning is defined as a systematic documentated, and collaborative process that includes inspection, testing, and training conducted to confirm that building and its associated serving systems are capable of being operated and maintained in conformance with the design intent. Design intent is the occupants’ assumed intention of the design as well as their assumed operation of the b uilding and the servicing systems. Design intent sets the requirements for occupant satisf action and forms the foundation and basis of technical criteria and documents — design drawings and specifications. H AC&R is one of the building servicing systems. Wilkinson (1999a, 1999b) recommended that the design intent document include the follo wing: ●

●

●

●

●

●

●

●

General description of the building type and occupancy category Particular needs such as air cleanliness, and outdoor air v olume fl w rate, as well as processes that require special indoor environments Outdoor and indoor design conditions Space pressurization and relative pressurization of adjacent spaces Emergency operation during utility outages Applicable codes and fir / life safety requirements HVAC&R system selection if it is mandated by the owner Operations and maintenance manuals and staff training

Necessity of HVAC&R Commissioning Air conditioning or HVAC&R systems in buildings are commissioned based on a sequence of planning, design, bidding, and construction. Man y installed air conditioned systems do not w ork as

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expected by the owners. Because of the higher demands of indoor environmental control, indoor air quality, and energy-efficient operation as well as the ad ances in technology in recent years, especially the microprocessor-based controls, an air conditioning system becomes more and more complicated and is required to provide considerably more functions than several decades ago. Tseng (1998) reported on a survey of 60 commercial buildings; the results were presented in the 1994 National Conference on Building Commissioning sponsored by Portland Ener gy Conservation, Inc. This survey found the following: ●

●

●

More than one-half of the buildings suffered control problems. Of the buildings 40 percent had HVAC&R equipment problems. Of the buildings 25 percent had ener gy management control systems, economizers, and variablespeed drives that did not run properly.

There are man y of b uilding o wners and f acility managers who claim that the functional performance of their HVAC&R systems does not meet their expectations. ASHRAE / IESNA Standard 90.1-1999 mandates that HV AC control systems shall be tested to ensure that control elements are calibrated, adjusted, and in proper w orking condition. For projects larger than 50,000 ft 2 (4650 m2) conditioned space area, except warehouses and semiheated spaces, detailed instructions for commissioning HV AC systems shall be pro vided by the designer in plans and specifications. Portland Energy Conservation noted top deficiencies disc vered by HVAC&R commissioning: ●

●

●

●

●

●

●

●

●

Incorrect scheduling of HVAC&R Incorrect cooling and heating sequence of operation Incorrect calibration of sensors and instrumentation Lack of control strategies for optimum comfort and energy-efficient operation Improper air and water economizer operations Microprocessor-based DDC systems not fully utilized Short cycling of HVAC&R equipment Lack of design intent Lack of training of HVAC&R operators or service contractor for complex systems Tseng (1998) cites the benefits of H AC&R commissioning:

●

●

●

●

●

SH__ ST__ LG__ DF

HVAC&R commissioning is the quality control tool for b uilding owners, and they know that the savings from fewer costly change orders and lo wer operating and maintenance costs will accrue. For those jobs with commissioning requirements, contractors and subcontractors kno w that it is not advisable for them to cut corners, and they benefit from l wering their cash set-asides for warranty reserves and callbacks. The architects can e xpect a b uilding with f ar fe wer postconstruction headaches to handle after commissioning. The engineers kno w that the HV AC&R systems are virtually assured of operating as intended, thus eliminating postoccupancy troubleshooting visits. The b uilding occupants enjo y a higher -performing b uilding with a smoothly functioning HVAC&R system which provides a healthy and comfortable indoor environment with the benefit of increased productivity.

Scope of HVAC&R Commissioning Tseng (1998) and Ellis (1998) described the scope of HVAC&R commissioning as follows:

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●

●

●

●

●

32.3

Clarifly wner priorities and design intent. The commissioning outcomes include a healthy and comfortable indoor environment, acceptable indoor air quality, energy efficien y, optimized functional performance, maintainability, and constructibility. Constructibility is the ability of a project to be built as designed. Document and monitor all f acets of construction. Documentation includes preparation of the commissioning specification preparation of the v erification test procedures and re view of k ey equipment and instrumentation submittals. Monitoring includes preparation of system readiness checklists such as electricity to the f ans, pumps, and compressors; monitoring and v erifying equipment / system start-up and operation as well as the performance of temperature and EMCS controls; and testing, adjusting, and balancing (TAB) work. Verify TAB work. In HVAC&R commissioning, TAB is a tool and is a part of the commissioning. It is often provided by a TAB subcontractor. Extensively test all subsystems and their components and controls (acceptance commissioning). Verify and document functional performance testing of all systems including control systems, so that all systems comply with the contract documents. Establish an as-deli vered performance record, and verify the as-b uilt record for all systems. The documentation should also include the corrective action reports for every deficien y found, the follow correction and retest. For acceptance, the final commissioning report should be completed and submitted and thus, the recommendation of acceptance determined. All documents should be turned o ver to the owner. Provide detailed training of operating personnel for each major system and equipment. Training plans and attendance records should be supervised. It is important to conduct system training as well as its operation so that all the indi vidual components are functioning together as a system. System operators cannot be e xpected to operate the system properly unless the y know how it is intended to work and why. If possible, operation and maintenance staff are encouraged to participate in the verification tests as an xcellent training. Continue post-acceptance commissioning and ongoing monitoring. Post-acceptance commissioning and ongoing monitoring are a continuation of adjustment, optimization, and improvement of HVAC&R systems to meet a specified goal by monitoring specific system operating parameter Data trend and diagnostic capabilities of DDC control systems are v aluable during HV AC&R commissioning and post-acceptance commissioning.

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Testing, Adjusting, and Balancing (TAB) Testing means to determine quantitati ve performance of equipment. Adjusting denotes that flui fl w rate and air fl w pattern are adjusted as specified in design. Balancing indicates that f ws are adjusted in proportion within the distribution system based on design requirements. ASHRAE / IENSA Standard 90.1-1999 mandates that construction documents shall require that all HVAC systems be balanced in accordance with generally accepted engineering standards. Ducted air and water fl w rates shall be measured and adjusted within 10 percent of design fl w rates. Variable-speed, variable-volume fl w distribution systems need not be balanced upstream of a pressureindependent control de vice. A written balance report shall be pro vided to the o wner for HVAC&R systems serving zones with a total conditioned area exceeding 5000 ft2 (465 m2). Standard 90.1-1999 also mandates that air systems shall be balanced first to minimize throttling losses. Then, in systems with fans of power greater than 1 hp (0.75 kW), fan speed shall be adjusted to meet design fl w conditions, except variable-fl w distribution systems need not be balanced upstream of the controlling device (such as a calibrated VAV box). Standard 90.1-1999 also mandates that hydronic (w ater) systems shall be proportionally balanced first to minimize throttling losses and then the pump impeller shall be trimmed or pump speed shall be adjusted to meet design v olume fl w requirements. Each w ater system shall ha ve facility either to measure pressure increase across the pump or ha ve test ports on each side of each

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pump. Exceptions include pumps with motors of 10 hp (7.5 kW) or less or when throttling losses are no greater than 5 percent of the po wer draw or 3 hp (2.3 kW), whichever is greater, above that required after the impeller is trimmed.

HVAC&R Commissioning Team An HVAC&R commissioning team should be or ganized to perform commissioning. According to Tseng (1998), this team should include the following members and perform the duties listed belo w: Owner. The o wner should be acti vely involved in commissioning, such as by defining require ments and assigning commissioning responsibilities, defining the scope and tasks for the commis sioning authority, and developing enforceable contractual pro visions to ensure compliance by the contractor. Commissioning Authority or Comission Consultant (CC). The commissioning authority or commission consultant should have specific duties including r viewing the plans and specifications tha related to the commissioning processes, scheduling and conducting all pertinent commissioning activities, such as precommissioning activities, training, walk-through inspections, tests and balances, and documentation. The commission consultant plays a k ey role in an HV AC&R commissioning. For a lar ge HVAC&R project, an outside e xpert can be hired as an independent CC. F or a small HVAC&R project, the owner, the engineer, the general contractor, or the principal mechanical subcontractor may serve as the CC. Architect. The architect ensures that the commissioning authority re submittal.

views the shop dra wing

Engineer. The engineer must be acti vely included in the commissioning acti vities: design intent documentation, verification testing and training. General Contractor . The general contractor must include the cost for commissioning requirements in the bid price and ensure that such requirements in the mechanical and electrical subcontracts are complied with. Mechanical / Electrical Subcontractor . The mechanical / electrical subcontractor must also include commissioning requirements in the contract price, ensure participation of subcontractors, coordinate all testing with pertinent specialty subcontractors, conduct walk-through inspections and hands-on trainings, provide certification of system performance and functional performance of ma jor equipment to the CC, and turn o ver a complete set of as-b uilt drawings to the design engineer . Test, Adjusting, and Balancing Contractor . The TAB contractor conducts TAB w ork, demonstrating to the CC the tested performance and participating in training sessions. Energy Management and Control System (EMCS) Contractor . A properly installed and debugged EMCS system is essential for HVAC&R commissioning to succeed.

When to Perform HVAC&R Commissioning

SH__ ST__ LG__ DF

Ellis (1998) recommended three kinds of HVAC&R commissioning processes: First, design-through-occupancy commissioning is ideal for the HV AC&R commissioning process. The CC serv es as the o wner’s technical liaison throughout the design, construction, and start-up of the project. The CC reviews the design engineer’s plans for compliance with the design

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32.5

intent. The CC also makes sure that during verification testing the temperature, pressure, fluid f w, control system signals, and actions are commissionable and included in the design. Second, postconstruction commissioning is most pre valent recently. Postconstruction commissioning often requires at least 1 year after the completion of the construction so that both summer cooling and winter heating mode operations can be included. In this case, the commission consultant is most often an outsider who w as not involved in the problems and politics of the design and construction. Therefore, conciliatory and professional interpersonal skills are e xtremely important for a CC. The CC also will learn about, evaluate, test, and document the installed HVAC&R system and become involved in the operators’ training. Third, existing building recommissioning helps to reduce HV AC&R energy consumption, improves IAQ, and improves indoor environmental control.

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Cost of HVAC&R Commissioning Because the HVAC&R commissioning process is still being de veloped and standardized, a rule of thumb for HVAC&R commissioning cost is between 2 and 5 percent of the construction cost of the systems to be commissioned. The cost of HVAC&R commissioning can be set aside as an independent budget item or added to the architectural / engineering or contractor fees. The commissioning cost will be recovered by reductions in change orders and claims.

32.2 HVAC&R MAINTENANCE Basics HVAC&R maintenance is the w ork required to maintain or restore HV AC&R systems, including equipment, instrumentation, components, and materials, to condition such that the y can be ef fectively operated to meet specified requirements ASHRAE Handbook 1999, HVAC Applications, defines repair as to mak e good or restore to good or sound conditions, and defines service as what is necessary to effect a maintenance program short of repair . In the ASHRAE handbook, maintenance also is classified into the foll wing categories: ●

●

●

Run-to-failure is a kind of arrangement such that no mone y is spent on maintenance prior to equipment or system breakdown. In planned maintenance, all functions and resources in this cate gory must be planned, budgeted, and scheduled. Planned maintenance can be subdi vided into pre ventive and correcti ve maintenance. Pre ventive maintenance is a kind of scheduled maintenance for an HV AC&R system, equipment, or components, in order to maintain durability , reliability, eff ciency, and safety. In corrective maintenance, corrective action is often the remedial action performed before failure occurs. Corrective action taken during a shutdown in response to a failure is called a repair. Predictive maintenance is based on equipment and system monitoring, the operating conditions, and performance to discover faults and degradations and thus the remedy performed.

Most manuf acturers give detailed instructions in ho w to operate and maintain the specific equip ment and instruments. These instructions must be followed.

Maintenance Contractors and Maintenance Personnel In most small facilities with comparatively simple HVAC&R systems, outside maintenance contractors often provide maintenance service based on the specified maintenance program. The time inter-

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CHAPTER THIRTY-TWO

val between tw o maintenance programs depends on the operating hours of the HV AC&R system; dirt accumulation in the filte , coil, and distribution devices; and recalibration requirements of the instruments. F or small and complicated HV AC&R systems, whenever the operator cannot repair and service the HV AC&R system, the owner or the f acility manager should ensure that qualifie contractors are hired for maintenance. For more complicated medium-size multizone HVAC&R systems, operating personnel are often responsible for HV AC&R system maintenance. The maintenance programs should be detailed in the operation and maintenance manual and tailored to each specific uilding. For highly technical equipment and systems, an outside maintanance contractor with the specific xpertise required should be called for service. For large HVAC&R systems with central plants, a management organization with one operations person and one maintenance person is required. Computerized maintenance programs should be used to provide detailed timing of system maintenance procedures. Logged information and proper data management associated with predicti ve maintenance will reduce f ailure response requirements. An outside maintenance service contractor with specific xpertise may need to be called.

Monitoring and Fault Detection and Diagnostics Assisting Predictive Maintenance As discussed in Sec. 5.15, in EMCS and fault detection and diagnostics, many operating parameters of the HVAC&R system are monitored and measured. Addition of necessary monitoring parameters for predictive maintenance may be considered. Faults occur when the actually measured parameters de viate from the normal operating v alues. Performance degradation is an evolving fault accumulated during a certain period of time. By using rule-based e xpert systems, autoregressive with e xogenous inputs (ARX) models, and artificial neural net ork (ANN) models, performance de gradation and f aults can be determined, and the corrective maintenance action recommended. HVAC&R maintenance is already an important part of optimizing an ener gy-efficient H AC&R system operation, to improve IAQ, and to maintain a safe, healthy, and comfortable indoor environment. More and more computer software will be developed to combine optimizing, energy-efficien HVAC&R system operation and predictive maintenance in the future.

REFERENCES

SH__ ST__ LG__ DF

ASHRAE, ASHRAE Handbook 1999, HVAC Applications, ASHRAE Inc., Atlanta, GA, 1999. Austin, S. B., HVAC System Trend Analysis, ASHRAE Journal, no. 2, 1997, pp. 44 – 50. Bearg, D. W., The Use of Multipoint Monitoring as a Tool for Commissioning Buildings for IAQ, ASHRAE Transactions, 1999, Part I, pp. 1101 – 1108. Chamberlain, C. S., Statistical Validation Testing vs. Commissioning, Heating / Piping / Air Conditioning, no. 8, 1995, pp. 59 – 62. Ellis, R. T., Building Systems Commissioning, Engineered Systems, no. 1, 1998, pp. 108 – 116. Seem, J. E., J. M. House, and R. H. Monroe, ASHRAE Journal, no. 7, 1999, pp. 21 – 26. Taber, G., Preventive Maintenance, Engineered Systems, no. 6, 1994, pp. 68 – 73. Tseng, P. C., Building Commissioning: Benefits and Costs HPAC, no. 4, 1998, pp. 51 – 59. Tseng, P. C., D. R. Stanton-Hoyle, and W. M. Withers, Commissioning through Digital Controls and an Advanced Monitoring System – A Project Perspective, ASHRAE Transactions, 1994, Part I, pp. 1382 – 1392. Wilkinson, R. J., Integrated Commissioning Avoids IAQ Pitfalls through Quality Construction, HPAC, no. 3, 1999a, pp. 79 – 83. Wilkinson, R. J., The Commissioning Design Intent Narrative, ASHRAE Journal, no. 4, 1999b, pp. 31 – 35.

APPENDIX A

NOMENCLATURE AND ABBREVIATIONS

A.1 NOMENCLATURE A a Aa Ac AD Ae,l Af Ai Ak Ao Ap Ar Ar AS B Bi C Ccc Cd Cdi Ce Ccir Ciu Cn

area, ft2 turbulence factor face area, ft2 core area of outlet, ft2 Dubois surface area of naked body, ft2 effective leakage area, ft2 area of fins ft2 inner surface area, ft2 net or unobstructed area of grille, ft2 outer surface area, ft2 primary surface area, ft2 cross-sectional area of room perpendicular to airfl w, ft2 Archimedes’ number annual savings, dollars ratio of outer surface area to inner surface area Ao/Ai Biot number scale factor; concentration of air contaminant, mg/m3; local loss coefficient; cost dollars cloudy-cover factor discharge coefficient degradation coefficien duct installation cost, dollars energy cost, dollars circulating factor unit cost of duct installation, $/ft2 clearness number of sky

Co cp cpa cpd cpr Cpre cps cpw csat Cs,i Cs,o Ci Cto Cv Cw CR CDD COP COPc COPhp

local loss coefficien specific heat at constant pres sure, Btu/lb °F specific heat of moist air at con stant pressure, Btu/lb°F specific heat of dry air at con stant pressure, Btu/lb °F specific heat of liquid refrigeran at constant pressure, Btu/lb°F pressure loss coefficien specific heat of ater vapor at constant pressure, Btu/lb °F specific heat of ater at constant pressure, Btu/lb °F saturation-specific heat per de gree wet-bulb temperature, Btu/lb °F inlet system effect loss coefficien outlet system effect loss coeffi cient scale factor for temperature lines, °F/ft total cost, dollars fl w coefficient a fl w rate of 1 gpm at a pressure drop of 1 psi scale factor for humidity ratio lines, lb/lb ft convective and radiative heat loss, Btu/hft2 cooling degree-day, degree-day coefficient of performanc coefficient of performance o chiller coefficient of performance o heat pump

A.1

A.2

APPENDIX A

COPhr COPref CRF D Daw De Dh Dlv DT E e Edif Emax Er Ersw EER ET* F f Fa,t Fblock Fcl Fcyc Fp Frej, HRF Fs Ft Fo Fr G g gc

coefficient of performance o heat recovery coefficient of performance of re frigeration capital recovery factor diameter, in. or ft; depreciation, dollars per year mass diffusivity for water vapor through air, ft2/s equivalent diameter, in. or ft hydraulic diameter, in. or ft mass diffusivity of liquid and vapor, ft2 /h mass diffusivity due to temperature gradient, ft2 /h evaporative heat loss, Btu/h ft2; electric potential, Volt; efficien y energy, Btu/b evaporative heat loss due to direct diffusion, Btu/hft2 maximum level of evaporative heat loss, Btu/h  ft2 unit energy cost, $/kWh sweating due to thermoregulatory mechanism Btu/hft2 energy efficien y ratio, Btu/hW effective temperature, °F factor, shape factor, solar heat gain factor friction factor; frequency, Hz air transport factor blockage factor clothing efficien y cycling loss factor performance factor heat rejection factor coil core surface area parameter fin thickness in.; fan total pressure, in. WC Fourier number Froude number mass velocity, lb/hft2 gravitational acceleration, ft/s2 dimentional conversion factor, 32 lbmft/lbfs2

Gr H h hb hcon Hf hig hfg,r hfg,32 hi hm ho hr hru Hs hs h Hs,r hs,r Ht Hv hwet HC HDD I i ID Id IG

Grashof number head, ft of water column; height, ft; hour angle, deg enthalpy, Btu/lb; heat-transfer coefficient Btu/hft2 °F boiling heat-transfer coefficient Btu/h ft2°F condensing heat transfer coeffi cient, Btu/h ft2 °F frictional head loss, ft of water latent heat of vaporization, Btu/lb latent heat of vaporization of refrigerant, Btu/lb latent heat of vaporization at 32°F, Btu/lb inner surface heat-transfer coeffi cient, Btu/h ft2 °F convective mass-transfer coeffi cient, ft/h heat-transfer coefficient at oute surface, Btu/h ft2 °F; outdoor air enthalpy, Btu/lb enthalpy of space air enthalpy of recirculating air, Btu/lb static head, ft of water enthalpy of saturated air film Btu/lb enthalpy of supply air head difference between supply and return mains, ft WC enthalpy of saturated air film a evaporating temperature, Btu/lb total head, ft of water velocity head, ft of water heat-transfer coefficient of wet ted surface, Btu/h ft2 °F heating capacity, Btu/ft2 °F heating degree-day, degree-day electric current, amp interest rate, percent direct radiation, Btu/h ft2 diffuse radiation, Btu/h ft2 global radiation on a horizontal plane, Btu/h ft2

NOMENCLATURE AND ABBREVIATIONS

IDN im Ia Irad Iref It Itur Isc IC IL J j K k K’ Kcc Ki Kp Kp KV L Laf Le Lw Lp LpA Lpr Lw Lt Lw,b Lwr Le

solar radiation on a surface normal to sun rays, Btu/h ft2 moisture permeability of clothing extraterrestrial solar intensity, Btu/h  ft2 effective radiant field W/ft2 reflection of solar radiation Btu/h  ft2 total intensity of solar radiation, Btu/h  ft2 intensity of turbulence solar constant, 434.6 Btu/h ft2 initial cost, dollars insertion loss, dB Joule’s equivalent, 778 ftlbf /Btu cost escalation factor constant, factor, coefficient derivative gain thermal conductivity, Btu  in./h ft2 °F, Btu/h ft °F wet-bulb temperature constant cloudy reduction factor integral gain power constant proportional gain, pressure constant volume constant distance, thickness, ft; sound level, dB; latitude angle, deg airfl w noise, dB equivalent length, ft sound power level, dB re 1 pW sound pressure level, dB re 20 Pa sound pressure level in dBA room sound pressure level, dB vertical distance between state points, ft horizontal distance between state points, ft branch power division, dB room sound power level, dB Lewis number

LHG LR M m m m˙ m˙par m˙r ms N n

n Nr NTU Nu OC P p Pair pat Pcfm Pcom pcon pdis pdy pev Pf pf pfil pfi pf,s

A.3

latent heat gain, Btu/h Lewis relation molecular weight; metabolic rate, Btu/hft2 mass, lb slope of air enthalpy saturation curve mass fl w rate, lb/min, lb/h rate of air contaminants generated, mg/s mass fl w rate of refrigerant, lb/h, lb/min surface density, lb/ft2 number number of moles, mol; number of air changes, ach; circulation number; amortization period, year depreciation period, year number of rows number of transfer units Nusselt number operating cost, dollars power, hp, kW; perimeter, ft; penetration pressure, psi, psia, psig air power, hp atmospheric pressure, psia power per unit volume fl w, W/cfm compressor power, hp condensing pressure, psig discharge pressure of compressor, psig dynamic loss, in WC evaporating pressure, psig fan power, input, hp pressure drop due to frictional and dynamic losses, in. WC fill pressure abs psia fi ed part of system pressure loss, in. WC frictional loss per floor inside th pressurized stairwell, in. WC

A.4

APPENDIX A

pf,u PH Pin pod Pp pp-od pres ps ps,i ps,o ps,r pst psuc psy pt pt,ex PV pv pvar pvo pvw pw PMV Pr Q q Qc Qcc Qc,c Qc,r qc,wet Qev, Qref, Qrl qlg Qch

duct frictional per unit length, in. WC horizontal projection, ft power input, hp total pressure loss of damper when fully open, in. WC pump power, hp total pressure loss of airf ow path excluding damper, in. WC residual pressure, in. WC static pressure, in WG, psig inlet system effect pressure loss in. WC outlet system effect pressure loss, in. WC static regain, in. WC pressure due to stack effect, lbf /ft2 suction pressure of compressor, psig system pressure loss, in. WC total pressure, in. WC external total pressure, in. WC vertical projection, ft velocity pressure, in. WC variable part of system pressure loss, in. WC velocity pressure at outlet, in. WC wind velocity pressure, in. WC water vapor pressure, psia predicted mean vote Prandtl number rate of heat transfer, Btu/h rate of heat tansfer, Btu/h coil’s load, Btu/h cooling coil’s load, Btu/h corrected cooling capacity, Btu/h catalog-listed cooling capacity, Btu/h heat and mass transfer, Btu/h refrigeration load at evaporator, Btu/h heat input to the f rst-stage generator, Btu/h ton heating coil load, Btu/h

qint Qrc Qrej, THR Qrh qRCi Qrs Qrsp qrs,t qsen qtran qwi R R R* Rc Rcl Rcom Ren Rf Rfa Rg Rload, LR Ro RT,l Re ROR S s Sf SH SW

internal heat gain, Btu/h space cooling load, Btu/h total heat rejection, Btu/h space heating load, Btu/h inward heat f ow from inner surface of sunlit window, Btu/h space sensible cooling load, Btu/h space sensible cooling load at part load, Btu/h sensible cooling load at time t, Btu/h rate of sensible heat transfer, Btu/h tramission loss, Btu/h heat gain admitted into conditioned space, Btu/h gas constant, ftlbf /lbm °R; electric resistance,  R-value, h  ft2 °F/Btu thermal resistance, h °F/Btu; f ow resistance, in WC/(cfm)2; ratio radius of curvature, in. or ft thermal resistance of clothing, h ft2  °F/Btu compression ratio entrainment ratio fouling factor, h ft2 °F/Btu ratio of free area to gross area gas-side thermal resistance, h ft2 °F/Btu load ratio universal gas constant ft lbf /lbm °R ratio of temperature lift Reynolds number rate of return salvage value, dollars; heat storage, Btu/hft2 specif c entropy, Btu/lb °R; dimensionless distance f n spacing, f ns/in. shaded height, ft shaded width, ft

NOMENCLATURE AND ABBREVIATIONS

SC Sc SHG SHGF SHR SHRc SHRs SHRsp SP St T T T* T tan Tco Tdew, T  Tdis Tf Ten Tm Tm To T os To, ws Tp TR Tr Tra, Trad Trm Trp Trp Tru

shadding coeff cient Schmidt number sensible gain, Btu/h solar heat gain factor, Btu/h  ft2 sensible heat ratio sensible heat ratio of cooling and dehumidif cation process sensible heat ratio of space conditioning line sensible heat ratio of space conditioning line at part load simple payback, years Strouhal number temperature, °F wet-bulb temperature, °F thermodynamic wet-bulb temperature, °F bulky air temperature unaffected by surface, °F annual operating hours, h changeover temperature, °F dew point temperature, °F discharge temperature, °F fan temperature rise, °F mass temperature of building envelope, °F temperature of mixture, °F log-mean temperature difference, °F operative temperature, °F; outdoor temperature, °F summer mean coincident wetbulb temp, °F statistically determined winter design outdoor temperature, °F plenum air temperature, °F absolute temperature, °R space temperature, °F mean radiant temperature, °F average space temperature, °F space temperature at part load, °F space temperature in perimeter zone, °F temperature of recirculating air, °F

Ts Tsa Tt,r Tws TD TL TLin TLout U u Ui Uu UAC V V˙ v vc vcon V˙ conv V˙ ef vfc V˙ gal V˙lk,V˙L V˙o V˙o,dm V˙oif V˙osn

A.5

supply temperature, °F temperature difference between the surface and air, °F throttling range, °F chilled water supply temperature, °F temperature differential, °F transmission loss, dB break-in transmission loss, dB breakout transmission loss, dB overall heat transfer coeff cient, Btu/h ft2 °F internal energy, Btu/lb; peripheral velocity, fpm overall heat-transfer coeff cient based on inner surface area, Btu/h ft2 °F overall heat-transfer coeff cient based on outer surface area, Btu/h ft2 °F uniform annual cost, dollars volume, ft3 volume f ow rate, cfm velocity, fpm of ft/sv specif c volume or moist volume, ft3/lb centerline velocity, fpm air velocity in constricted part of damper or duct f ttings, fpm volume f ow rate of upward convective f ow, cfm volume f ow rate of exf ltrated air, cfm face velocity, fpm volume f ow rate of chilled water, gpm volume f ow rate of air leakage, cfm volume f ow rate of outdoor air, cfm minimum outdoor air supply volume f ow at design conditions, cfm volume f ow rate of outdoor and inf ltrated air, cfm zone outdoor air supply volume f ow rate, cfm

A.6

APPENDIX A

V˙p V˙s V˙sp W w Win Wisen Wrsw w*s X x xrl y Z z zstat

piston displacement, cfm supply volume f ow rate, cfm supply volume f ow rate at part load, cfm work, Btu/lb; mechanical work performed, Btu/hft2; sound power, dB; width, ft relative velocity, fpm; humidity ratio, lb/lb work input, Btu/lb isentropic work, Btu/lb wetted portion of human body due to sweating saturated humidity ratio at thermodynamic wet-bulb temperature, lb/lb moisture content, dimensionless or percent; mass fraction mole fraction; quality or dryness fraction; coordinate dimension quality or refrigerant leaving overfeed cooler vertical drop of air jet, ft compressibility factor elevation, ft stationary level, ft

Subscripts a ab ae, aen alv am at av b bg by c ca

air, ambient, absorber absorber, air at dry-wet boundary entering air leaving air ambient atmospheric average body bleed, branch, building material building bypass coil, cooling, cold, convective, condenser, compressor, common end, corrected cooling air, cooled air

cc c,d ce hc cl cn co com con corr cr cs d dam deh des dif dis dl dy e ee ef eff el en en,c ev ev,c ex, exh exf, ef f fc fx f fu g 1g

cooling coil closed door entering condenser heating coil clothing, cooling load, cooling coil, leaving condenser common changeover compressor condensing, condenser correction body core sensible cooling coil duct, design damper dehumidif er desiccant diffuser discharge process air leaving desiccant dehumidif er dynamic entering entering evaporator exf ltrated effective elevation, leaving evaporator, equivalent entering cooling water entering condenser evaporating, evaporator evaporative cooling exhaust exf ltration fan fan coil f xed part l oor furnace moisture gain, gas, globe, ground, generator f rst-stage generator

NOMENCLATURE AND ABBREVIATIONS

2g go h hg h,t hu i in inf int k l lc le liq lk lr lv m mat max min mo, mot n o o,d oi os o,s o,sys out p par pd pl

second-stage generator saturated water vapor at 0°F higher, heat exchanger, heating, hot heat gain heat transformer humidif er inlet, input, indoor, interior, intermediate, inner surface input, indirect, interior inf ltration intermediate conduction latent, liquid, lower, lights, leaving liquid at condenser water entering evaporator liquid leakage liquid refrigerant leaving mixture, mean, maximum, motor material maximum minimum motor number outdoor, output, outlet, outer surface, oversaturation open door inward f ow from outer surface outer surface summer outdoor system outdoor air outdoor air constant pressure, people, partload, pump, process air, primary air particulates in air dry air at constant pressure process air after sensible heat exchanger

ps pt r rc rd rec ref reff rel res ret, rt rf rg rh rl ro rp rs rs ru s s sa sa sat sb sc sc sd sen sf sg sh

A.7

water vapor at constant pressure, primary air supply plant space, room, return, refrigerant, radiative, regeneration air space cooling, refrigeration capacity return duct recirculating refrigeration free refrigeration release residual, respiration return return fan, relief fan, refrigeration effect regeneration air entering desiccant dehumidif er space heating space latent regeneration air leaving desiccant dehumidif er return plenum space sensible return system recirculating air entering the AHU or PU supply, steam, saturated state, surface, sunlit, straight-through end, summer saturated at thermodynamic wetbulb temperature solution at absorber saturated air f lm saturation surface at dry-wet boundary subcooled cold air supply supply duct sensible supply fan solution at generator shaded

A.8

APPENDIX A

sh si sil sk sn sn,d sn,p sol ss st suc sun sur sx T t un var ve w wb we wet wl w,o ws xl

warm air supply supply air for interior zone silencer skin supply air for zone n zone supply air at design condition supply air for zone n at part load solar supply system supply temperature suction sunlit surroundings supply air for perimeter zone total, overall, temperature time unconditioned variable part saturated water vapor leaving evaporator water, condensate, winter water at dry-wet boundary, boiler hot water water entering wet air water leaving winter outdoor water vapor at saturated state regeneration air leaving the sensible heat exchanger, exit of heat exchanger

Greek Letter Symbols





mean temperature coeff cient; angle between the air conditioning process and horizontal line on psychrometric chart, deg; thermal diffusivity, ft2/s; absorptance; spreading angle of air jet, deg; damper characteristic ratio solar altitude angle, deg; blade angle, deg

 

ex

sat

wet  cb com cp dr f fu isen mec mo, mot ov p s sat sy,h t v ww     suc   

ratio of specif c heat at constant pressure to constant volume, surface-solar azimuth angle, deg difference solar declination angle, deg emissivity; absolute roughness, in.; effectiveness; effectiveness factor air exchange eff ciency saturation effectiveness wet coil effectiveness eff ciency combustion eff ciency compressor eff ciency compression eff ciency eff ciency of driving mechanism fan total eff ciency; f n eff ciency furnace eff ciency isentropic eff ciency mechanical eff ciency motor eff ciency overall eff ciency pump eff ciency fan static eff ciency, f n surface eff ciency saturation effectiveness system eff ciency for heating fan total eff ciency volumetric eff ciency wire-to-water eff ciency angle of incidence, deg; effective draft temperature, °F absolute viscosity, lb/ft s; degree of saturation; mechanical eff ciency kinematic viscosity, ft2/s density, lb/ft3; ref ectance density of suction vapor, lb/ft3 angle between tilting and horizontal surface, deg Stefan-Boltzmann constant, 0.1714  10 8 Btu/h ft2 °R4; standard deviation transmittance

NOMENCLATURE AND ABBREVIATIONS

 f  

relative humidity, percent; solar azimuth angle, deg f n resistance number surface azimuth angle, deg prof le angle, deg

A.2 ABBREVIATIONS ABMA abs. AC ACEC ACI ACP ADC ADPI AFUE AHU AI AMCA ANSI AO ARI ASHRAE ASME ASTM AVI BAS BHP, bhp BI BLAST BMS

American Boiler Manufacturers Association absolute air conditioning American Consulting Engineers Council adjustable current inverter alternate component package Air Diffusion Council air diffusion performance index, percent annual fuel utilization eff ciency air-handling unit analog input Air Movement and Control Association American National Standards Institute analog output Air Conditioning and Refrigeration Institute American Society of Heating, Refrigerating, and Air Conditioning Engineers American Society of Mechanical Engineers American Society of Testing and Materials adjustable voltage inverter building automation system brake horsepower binary or digital input, backward-inclined blade building loads analysis and systems thermodynamics building management system

BO BOCA CABDS CADD CFC CLF CLTD COP CTD DA dc DDC DECOS DOE DOP DSA DWDI DX ECB EEPROM EIA EMS EPA EPROM ETD FC FDA FOM GWHP GWP HEPA HR HSPF

A.9

binary or pulsed output Building Off cials and Code Administrators computer-aided building design system computer-aided design and drafting chlorof uorocarbon cooling load factor cooling load temperature difference coeff cient of performance condenser temperature difference direct-acting direct current direct digital control design energy cost Department of Energy di-octyl phthalate double-strength sheet glass double-width double-inlet direct expansion, dry expansion energy cost budget electrically erasable, programmable, read-only memory Energy Information Administration of the Department of Energy energy management system Environmental Protection Agency erasable programmable readonly memory equivalent temperature difference fan coil, forward-curved blade Food and Drug Administration f gure of merit groundwater heat pump global warming potential high-eff ciency particulate air heart rate heating seasonal performance factor

A.10

APPENDIX A

HVAC&R IAQ ILD I/O I-P IPLV IRS LiBr LiCl LPG MAU MCPC MPS NBC NC NCDC NFPA NIOSH NO NPL NPSH NWWA ODP PC PI PID PU PTAC PURPA PVC PWM RA

heating, ventilating, air conditioning, and refrigeration indoor air quality internal load density input/output inch-pound integrated part-load value Internal Revenue Service lithium bromide lithium chloride liquef ed petroleum gas makeup air unit microcomputer constructed psychrometric chart manual position switch National Broadcasting Corporation noise criteria, normally closed National Climatic Data Center National Fire Protection Association National Institute of Occupational Safety and Health normally open neutral pressure level net positive suction head National Water Well Association ozone depletion potential personal computer proportional plus integral proportional-integral-derivative packaged unit packaged terminal air conditioner Public Utility Regulatory Policies Act polyvinyl chloride pulse-width modulated inverter reverse-acting

RAM RAU RC RH ROM RTD RTS SBS SEER SEUF SMACNA SI SPF SSE SSU SWSI TA TARP TETD TFM TIMA TRAV UL ULPA VAV VDC VLSI VVVT WHO WSHP WWR

random access memory recirculating air unit room criteria relative humidity read-only memory resistance temperature detector room temperature sensor sick building syndrome seasonal energy eff ciency ratio seasonal energy utilization factor Sheet Metal and Air Conditioning Contractors’ National Association International System of units seasonal performance factor steady-state eff ciency Saybolt-seconds univeral viscosity single-width single-inlet time-averaging thermal analysis research program total equivalent temperature differential transfer function method Thermal Insulation Manufacturers Association terminal regulated air volume Underwriters’ Laboratories ultra low-penetration air f lters variable air volume volts of direct current very large-scale integrated variable-volume variable-temperature World Health Organization water-source heat pump window-to-wall ratio

APPENDIX B

PSYCHROMETRIC CHART, TABLES OF PROPERTIES, AND I-P UNITS TO SI UNITS CONVERSION

B.1

B.2 FIGURE B.1 Psychrometric chart. Based on ASHRAE Psychrometric Chart No. 1. Reprinted with permission from ASHRAE Inc. Sensible heat ratio (SHR), humidity ratio scale in grains/lb, and two cooling and dehumidifying curv es were added by author.

TABLE B.1 Thermodynamic Properties of Moist Air (at Atmospheric Pressure 14.696 psia) and Water Volume, ft3/lb dry air

Saturated water vapor

Enthalpy, Btu/lb dry air

Temp T, °F

Humidity ratio ws, lbw/lbda

va

vas

vs

ha

has

hs

psi

in. Hg.

32 33 34 35

0.003790 0.003947 0.004109 0.004277

12.389 12.414 12.439 12.464

0.075 0.079 0.082 0.085

12.464 12.492 12.521 12.550

7.687 7.927 8.167 8.408

4.073 4.243 4.420 4.603

11.760 12.170 12.587 13.010

0.08865 0.09229 0.09607 0.09998

0.18049 0.18791 0.19559 0.20355

36 37 38 39

0.004452 0.004633 0.004820 0.005014

12.490 12.515 12.540 12.566

0.089 0.093 0.097 0.101

12.579 12.608 12.637 12.667

8.648 8.888 9.128 9.369

4.793 4.990 5.194 5.405

13.441 13.878 14.322 14.773

0.10403 0.10822 0.11257 0.11707

40 41 42 43

0.005216 0.005424 0.005640 0.005863

12.591 12.616 12.641 12.667

0.105 0.110 0.114 0.119

12.696 12.726 12.756 12.786

9.609 9.849 10.089 10.330

5.624 5.851 6.086 6.330

15.233 15.700 16.175 16.660

44 45 46 47

0.006094 0.006334 0.006581 0.006838

12.692 12.717 12.743 12.768

0.124 0.129 0.134 0.140

12.816 12.8946 12.877 12.908

10.570 10.810 11.050 11.291

6.582 6.843 7.114 7.394

48 49 50 51

0.007103 0.007378 0.007661 0.007955

12.793 12.818 12.844 12.869

0.146 0.152 0.158 0.164

12.939 12.970 13.001 13.033

11.531 11.771 12.012 12.252

52 53 54 55

0.008259 0.008573 0.008897 0.009233

12.894 12.920 12.945 12.970

0.171 0.178 0.185 0.192

13.065 13.097 13.129 13.162

56 57 58 59

0.009580 0.009938 0.010309 0.010692

12.995 13.021 13.046 13.071

0.200 0.207 0.216 0.224

60 61 62 63

0.011087 0.011496 0.011919 0.012355

13.096 13.122 13.147 13.172

0.233 0.242 0.251 0.261

Absolute pressure p

Enthalpy, Btu/lb Sat. water liq. hf

Evap. hig/hfg

Sat. water vapor hg

 0.02 0.99 2.00 3.00

1075.15 1074.59 1074.02 1073.45

1075.14 1075.58 1076.01 1076.45

0.21180 0.22035 0.22919 0.23835

4.01 5.02 6.02 7.03

1072.88 1072.32 1071.75 1071.18

1076.89 1077.33 1077.77 1078.21

0.12172 0.12654 0.13153 0.13669

0.24783 0.25765 0.26780 0.27831

8.03 9.04 10.04 11.04

1070.62 1070.05 1069.48 1068.92

1078.65 1079.09 1079.52 1079.96

17.152 17.653 18.164 18.685

0.14203 0.14755 0.15326 0.15917

0.28918 0.30042 0.31205 0.32407

12.05 13.05 14.05 15.06

1068.35 1067.79 1067.22 1066.66

1080.40 1080.84 1081.28 1081.71

7.684 7.984 8.295 8.616

19.215 19.756 20.306 20.868

0.16527 0.17158 0.17811 0.18484

0.33650 0.34935 0.36263 0.37635

16.06 17.06 18.06 19.06

1066.09 1065.53 1064.96 1064.40

1082.15 1082.59 1083.03 1083.46

12.492 12.732 12.973 13.213

8.949 9.293 9.648 10.016

21.441 22.025 22.621 23.229

0.19181 0.19900 0.20643 0.21410

0.39054 0.40518 0.42031 0.43592

20.07 21.07 22.07 23.07

1063.83 1063.27 1062.71 1062.14

1083.90 1084.34 1084.77 1085.21

13.195 13.228 13.262 13.295

13.453 13.694 13.934 14.174

10.397 10.790 11.197 11.618

23.850 24.484 25.131 25.792

0.22202 0.23020 0.23864 0.24735

0.45204 0.46869 0.48588 0.50362

24.07 25.07 26.07 27.07

1061.58 1061.01 1060.45 1059.89

1085.65 1086.08 1086.52 1086.96

13.329 13.364 13.398 13.433

14.415 14.655 14.895 15.135

12.052 12.502 12.966 13.446

26.467 27.157 27.862 28.582

0.25635 0.26562 0.27519 0.28506

0.52192 0.54081 0.56029 0.58039

28.07 29.07 30.07 31.07

1059.32 1058.76 1058.19 1057.63

1087.39 1087.83 1088.27 1088.70

B.3

B.4 TABLE B.1 (Continued) Volume, ft3/lb dry air

Saturated water vapor

Enthalpy, Btu/lb dry air

Temp T, °F

Humidity ratio ws, lbw/lbda

va

vas

vs

64 65 66 67

0.012805 0.013270 0.013750 0.014246

13.198 13.223 13.248 13.273

0.271 0.281 0.292 0.303

13.468 13.504 13.540 13.577

15.376 15.616 15.856 16.097

68 69 70 71

0.014758 0.015286 0.015832 0.016395

13.299 13.324 13.349 13.375

0.315 0.326 0.339 0.351

13.613 13.650 13.688 13.726

72 73 74 75

0.16976 0.017575 0.018194 0.018833

13.400 13.425 13.450 13.476

0.365 0.378 0.392 0.407

76 77 78 79

0.019491 0.020170 0.020871 0.021594

13.501 13.526 13.551 13.577

80 81 82 83

0.022340 0.023109 0.023902 0.024720

84 85 86 87

Absolute pressure p

Enthalpy, Btu/lb

hs

psi

in. Hg.

Sat. water liq. hf

13.942 14.454 14.983 15.530

29.318 30.071 30.840 31.626

0.29524 0.30574 0.31656 0.32772

0.60112 0.62249 0.64452 0.66724

32.07 33.07 34.07 35.07

1057.07 1056.50 1055.94 1055.37

1089.14 1089.57 1090.01 1090.44

16.337 16.577 16.818 17.058

16.094 16.677 17.279 17.901

32.431 33.254 34.097 34.959

0.33921 0.35107 0.36328 0.37586

0.69065 0.71478 0.73964 0.76526

36.07 37.07 38.07 39.07

1054.81 1054.24 1053.68 1053.11

1090.88 1091.31 1091.75 1092.18

13.764 13.803 13.843 13.882

17.299 17.539 17.779 18.020

18.543 19.204 19.889 20.595

35.841 36.743 37.668 38.615

0.38882 0.40217 0.41592 0.43008

0.79164 0.81883 0.84682 0.87564

40.07 41.07 42.06 43.06

1052.55 1051.98 1051.42 1050.85

1092.61 1093.05 1093.48 1093.92

0.422 0.437 0.453 0.470

13.923 13.963 14.005 14.046

18.260 18.500 18.741 18.981

21.323 22.075 22.851 23.652

39.583 40.576 41.592 42.633

0.44465 0.45966 0.47510 0.49100

0.90532 0.93587 0.96732 0.99968

44.06 45.06 46.06 47.06

1050.29 1049.72 1049.16 1048.59

1094.35 1094.78 1095.22 1095.65

13.602 13.627 13.653 13.678

0.487 0.505 0.523 0.542

14.089 14.132 14.175 14.220

19.222 19.462 19.702 19.943

24.479 25.332 26.211 27.120

43.701 44.794 45.913 47.062

0.50736 0.52419 0.54150 0.55931

1.03298 1.06725 1.10250 1.13877

48.06 49.06 50.05 51.05

1048.03 1047.46 1046.89 1046.33

1096.08 1096.51 1096.95 1097.38

0.025563 0.026433 0.027329 0.028254

13.703 13.728 13.754 13.779

0.561 0.581 0.602 0.624

14.264 14.310 14.356 14.403

20.183 20.424 20.664 20.905

28.055 29.021 30.017 31.045

48.238 49.445 50.681 51.949

0.57763 0.59647 0.61584 0.63575

1.17606 1.21442 1.25385 1.29440

52.05 53.05 54.05 55.05

1045.76 1045.19 1044.63 1055.06

1097.81 1098.24 1098.67 1099.11

88 89 90 91

0.029208 0.030189 0.031203 0.032247

13.804 13.829 13.855 13.880

0.646 0.669 0.692 0.717

14.450 14.498 14.547 14.597

21.145 21.385 21.626 21.866

32.105 33.197 34.325 35.489

53.250 54.582 55.951 57.355

0.65622 0.67726 0.69889 0.72111

1.33608 1.37892 1.42295 1.46820

56.05 57.04 58.04 59.04

1043.49 1042.92 1042.36 1041.79

1099.54 1099.97 1100.40 1100.83

92 93 94 95

0.033323 0.034433 0.035577 0.036757

13.905 13.930 13.956 13.981

0.742 0.768 0.795 0.823

14.647 14.699 14.751 14.804

22.107 22.347 22.588 22.828

36.687 37.924 39.199 40.515

58.794 60.271 61.787 63.343

0.74394 0.76740 0.79150 0.81625

1.51468 1.56244 1.61151 1.66189

60.04 61.04 62.04 63.03

1041.22 1040.65 1040.08 1039.51

1101.26 1101.69 1102.12 1102.55

ha

has

Abridged from ASHRAE Handbook 1997, Fundamentals. Reprinted with permission.

Evap. hig/hfg

Sat. water vapor hg

PSYCHROMETRIC CHART, TABLES OF PROPERTIES, AND UNITS CONVERSION

B.5

TABLE B.2 Physical Properties of Air (at Atmospheric Pressure 14.696 psia) Temp. T, °F

, lbm/ft3

cp, Btu/lbm °F

  105, lbm /ft s

  103, ft2/s

k, Btu/h ft °F

, ft2/h

Pr

  103, 1/°F

g2/2, 1/°F3 ft3

0 30 60 80 100 150 200 250 300 400 500 600 800 1000 1500

0.0862 0.0810 0.0764 0.0735 0.0710 0.0651 0.0602 0.0559 0.0523 0.0462 0.0413 0.0374 0.0315 0.0272 0.0203

0.240 0.240 0.240 0.240 0.240 0.241 0.241 0.242 0.243 0.245 0.247 0.251 0.257 0.263 0.277

1.09 1.15 1.21 1.24 1.28 1.36 1.45 1.53 1.60 1.74 1.87 2.00 2.24 2.46 2.92

0.126 0.142 0.159 0.169 0.181 0.209 0.241 0.274 0.306 0.377 0.453 0.535 0.711 0.906 1.44

0.0132 0.0139 0.0146 0.0152 0.0156 0.0167 0.0179 0.0191 0.0203 0.0225 0.0246 0.0270 0.0303 0.0337 0.0408

0.639 0.714 0.798 0.855 0.919 1.06 1.24 1.42 1.60 2.00 2.41 2.88 3.75 4.72 7.27

0.721 0.716 0.711 0.708 0.703 0.698 0.694 0.690 0.686 0.681 0.680 0.680 0.684 0.689 0.705

2.18 2.04 1.92 1.85 1.79 1.64 1.52 1.41 1.32 1.16 1.04 0.944 0.794 0.685 0.510

4.39  106 3.28 2.48 2.09 1.76 1.22 0.840 0.607 0.454 0.264 0.163 79.4  103 50.6 27.0 7.96

Source: Fundamentals of Momentum Heat and Mass Transfer, Welty et al., 1976. John Wiley & Sons. Reprinted with permission.

TABLE B.3 Physical Properties of Water (at Atmospheric Pressure 14.696 psia) T, °F

, lbm /ft3

cp, Btu/lbm °F

  103 lbm /ft s

32 60 80 100 150 200 250 300 400 500 600

62.4 62.3 62.2 62.1 61.3 60.1 58.9 57.3 53.6 49.0 42.4

1.01 1.00 0.999 0.999 1.00 1.01 1.02 1.03 1.08 1.19 1.51

1.20 0.760 0.578 0.458 0.290 0.206 0.160 0.130 0.0930 0.0700 0.0579

  105 k, ft2 /s Btu/ft °F 1.93 1.22 0.929 0.736 0.474 0.342 0.272 0.227 0.174 0.143 0.137

Source: Fundamental of Momentum Heat and Mass permission.

0.319 0.340 0.353 0.364 0.383 0.392 0.395 0.395 0.382 0.349 0.293

, ft2 /h 5.06 5.45 5.67 5.87 6.26 6.46 6.60 6.70 6.58 5.98 4.58

Pr

  103, 1/°F

13.7  0.350 8.07 0.800 5.89 1.30 4.51 1.80 2.72 2.80 1.91 3.70 1.49 4.70 1.22 5.60 0.950 7.80 0.859 11.0 1.07 17.5

g2 /  2 1/°F ft3 17.2 48.3 107 403 1010 2045 3510 8350 17350 30300

Transfer, Welty et al., 1976 John Wiley & Sons. Reprinted with

TABLE B.4 Conversion of Inch-Pound (I-P) Units to International System of Units (SI) Unit atm

Btu (British thermal unit) Btu ft/h ft2 °F Btu/h

Equivalents

Unit

14.696 lbf /in

33.91 ft of water

29.92 in. Hg.

1.013 bars

101,325 Pa

778 ft lbf

1055 J

252 cal

1.731 W/m °C

0.293 W 2

Btu/h cfm Btu/h ft Btu/h ft2 Btu/h ft2 °F Btu in./h ft2 °F Btu/lb Btu/lb °F Btu/lb ft Btu/yr ft2 clo (clothing insulation)

Equivalents

06209 W s/L

0.961 W/m

3.155 W/m2

5.678 W/m2 °C

0.1442 W/m °C

2.326 kJ/kg

4.187 kJ / kg °C

7.63 kJ / kg m

0.000293 kWh / yr ft2

0.155 m2 °C/W

B.6

APPENDIX B

TABLE B.4 Conversion of Inch-Pound (I-P) Units to International System of Units (SI) (Continued) Unit clo ft3 /lb cfm (cubic foot per minute) cfm/ft cfm/ft2 cfm/tonref $/ft2 °F fc fpm ft ft2 ft3 ft lbf ft lbf /min ft/s, fps ft2/s(kinematic viscosity) ft WC gal gpm (U.S) gpm/tonref h °F/Btu h ft2 °F/Btu hp hp (boiler) in. (inch) in. Hg (mercury) in. WC (water column) in. WC/(cfm)2 in. WG (water gauge) J (joule) kBtu/ft2 yr kg km kW kWh

Equivalents

0.88 h ft °F/Btu

0.0624 m3 /kg

7.481 gpm

0.4719 L/s

0.02832 m3/min

1.548 L / s m

5.078 L / s m2

18.2 m3 /h m2

0.1342 L / s k Wref

10.76 $/m2

(°F  32)/(1.8)°C

10.76 lx

0.01136 mi / h

0.00508 m / s

0.3048 m

304.8 mm

144 in.2

0.0929 m2

0.748 gal

1.356 J

0.0226 W

0.3048 m / s

92,900 mm2/s

0.4334 lbf /in2

2.99 kPa

0.1337 ft3

8.35 lb of water

3.785 L

0.0631 L / s

0.0179 L / s kW

1.911 °C / W

0.176 m2 °C/W

33,000 ft lbf /min

550 ft lbf /s

0.746 kW

33,476 Btu / h

9808 W

25.4 mm

0.4912 lbf /in.2

3.3 kPa

0.0361 lbf /in.2

5.20 lbf /ft2

248.6 Pa

5.27  105Pa s/m6

248.6 Pa  1 atm

9.48  10 4 Btu

3.153 kWh / m2 yr

2.2046 lb

3281 ft

0.6214 mi

3413 Btu / h

1.341 hp

3413 Btu 2

Unit kW/ton L lbf lb/Btu lb °F/Btu lbf /ft2 lb/ft3 lb/ft h lbf /ft s lb/h lb/h ft2 lb/lb lb (mass) lb of water L/s m met mg mil mi mi/h mm mm Hg mg mm oz Pa (pascal) pint ppm (mass) psia (absolute) psig (gauge) quad (quadrillion) qt (quart) rad (radian)

Equivalents

[3.516/(kW/ton)] COPref

0.001 m3

0.0353 ft3

4.45 N

0.4786 kg / kJ

0.2659 kg °C/kJ

0.0069 lbf /in2

4.88 kg / m2

16.0 kg / m3

0.413 mPa s

1490 mPa s

0.126 g / s

4.88 kg / h m2

1.0 kg / kg

7000 gr

16 oz

0.4536 kg

0.01602 ft3

0.12 gal

2.119 cfm

15.85 gpm

1.094 yard

3.281 ft

39.37 in.

58.2 W/ m2

18.46 Btu / h ft2

0.01543 gr

0.001 in.

25.4 mm

5280 ft

1.61 km

88 fpm

0.44 m / s

0.03937 in.

133.3 Pa

1  10 6 g

1  10 6 m

3.94  10 5in.

0.0625 lb

28.35 g

1 N / m2

28.37 in.3

0.4732 L

1 mg/kg

2.307 ft water abs.

703.1 kg / m2abs.

6.895 kPa abs.

1 lbf /in.2  1 atm

1  1015 Btu

1.055 EJ

57.75 in.3

0.9461 L

57.3°

PSYCHROMETRIC CHART, TABLES OF PROPERTIES, AND UNITS CONVERSION

TABLE B.4 Conversion of Inch-Pound (I-P) Units to International System of Units (SI) (Continued) Unit rpm therm ton h ton (long)

Equivalents

1 r/min

100,000 Btu

105.5 MJ

12,000 Btu

3.516 kWh

2240 lb

1016 kg

Unit ton (metric) tonref (refrigeration) ton (short) torr W

Most of the conversion equivalents are based on values in ASHRAE Handbook 1997, Fundamentals.

Equivalents

1000 kg

12,000 Btu / h

3.516 kW

2000 lb

1 mm Hg

3.413 Btu / h

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Abbreviations, A.9 – A.10 Absolute zero, 2.5 Absorption chiller-heaters, 14.20 – 14.22 actual performance, 14.22 heating cycle, 14.20 – 14.22 Absorption chillers, double-effect, direct-fired 14.6 – 14.18 absorber and solution pumps, 14.6 – 14.7 air purge unit, 14.8 – 14.9 capacity control and part-load operation, 14.16 – 14.17 coefficient of performance 14.14 condenser, 14.7 – 14.8 condensing temperature, 14.19 – 14.20 controls, 14.16 – 14.18 cooling water entering temperature, 14.19 cooling water temperature control, 14.17 – 14.18 corrosion control, 14.20 crystallization and controls, 14.17 difference between absorption and centrifugal chillers, 14.18 – 14.19 evaporating temperature, 14.19 evaporator and refrigerant pump, 14.6 fl w of solution and refrigerant, 14.9 – 14.11 generators, 14.7 – 14.8 heat exchangers, 14.6 – 14.7 heat removed from absorber and condenser, 14.19 mass fl w rate of refrigerant and solution, 14.11 – 14.12 monitoring and diagnostics, 14.18 operating characteristics and design considerations, 4.18 – 4.20 performance of, 14.11 – 14.16 rated conditions, 14.20 safety and interlocking controls, 14.18 series fl w, parallel fl w, and reverse parallel fl w, 14.8 – 14.9 Standard 90.1 – 1999 minimum efficien y requirements, 14.20 system description, 14.6 – 14.8 thermal analysis, 14.12 – 14.14 throttling devices, 14.8

Absorption heat pumps, 14.22 – 14.24 case study: series connected, 14.22 – 14.24 functions of, 14.22 Absorption heat transformer, 14.24 – 14.26 coefficient of performance 14.26 operating characteristics, 14.24 – 14.25 system description, 14.24 – 14.25 Accuracy, 2.6 Adiabatic process, 2.11 Adiabatic saturation process, ideal, 2.11 Air: atmospheric, 2.1 dry air, 2.1 – 2.2 mass, 3.25 moist air, 2.1 primary, 20.4 process, 1.4 – 1.5 recirculating, 20.4 regenerative, 1.4 – 1.5 secondary, 20.4 transfer, 20.4 ventilation, 4.29 Air cleaner, electronic, 15.69 – 15.70 Air conditioning, 1.1 – 1.2, industry, 1.15 project development, 1.16 – 1.17 Air conditioning processes, 20.41 – 20.53 adiabatic mixing, 20.50 – 20.52 air washer, 20.46 bypass mixing, 20.52 – 20.53 cooling and dehumidifying, 20.47 – 20.50 heating element humidifie , 20.46 humidifying, 20.45 – 20.47 oversaturation, 20.46 – 20.47 reheating, recooling and mixing, 20.74 – 20.75 relative humidity of air leaving coil, 20.49 – 20.50 sensible heat ratio, 20.41 – 20.43 sensible heating and cooling, 20.44 – 20.45 space conditioning, 20.43 – 20.44 steam injection humidifie , 20.45 – 20.46

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Air conditioning systems, 1.2 air, cooling and heating systems designation, 26.2 – 26.3 central, 1.6 central hydronic, 1.6 classification basic approach, 26.1 – 26.2 classification of 1.3 – 1.10, 26.2 – 26.3 clean room, 1.5 comfort, 1.2 – 1.3 desiccant-based, 1.4 evaporative-cooling, 1.4 individual room, 1.4 packaged, 1.6 space, 1.5 space conditioning, 1.5 thermal storage, 1.5 unitary packaged, 1.6 Air conditioning systems, individual, 26.8 – 26.9 advantages and disadvantages, 26.9 basics, 26.8 – 26.9 Air conditioning systems, packaged terminal, 26.13 – 26.15 equipment used, 26.13 – 26.14 heating and cooling mode operation, 26.13 – 26.14 minimum efficien y requirements, ASHRAE/IESNA Standard 90.1 – 1999, 26.14 – 26.15 system characteristics, 26.13, 26.15 Air conditioning systems, room, 26.9 – 26.13 configuration 26.10 – 26.11 controls, 26.12 cooling mode operation, 26.11 energy performance and energy use intensities, 26.11 – 26.12 equipment used in, 26.9 – 26.10 features, 26.12 system characteristics, 26.12 – 26.13 Air conditioning systems, selection: applications and building occupancies, 26.4 – 26.5 energy efficien y, 26.7 fire safety and smo e control, 26.7 – 26.8 indoor air quality, 26.5 – 26.6 initial cost, 26.8 maintenance, 26.8 requirements fulfilled 26.4 selection levels, 26.3 – 26.4 sound problems, 26.6 – 26.7 space limitations, 26.8 system capacity, 26.5 zone thermal control, 26.6

Air conditioning systems, space conditioning, 28.1 – 28.3 advantages and disadvantages, 28.2 – 28.3 applications, 28.1 – 28.2 induction systems, 28.3 Air contaminants, indoor, 4.27 – 4.28, 15.61 Air duct design, principles and considerations, 17.43 – 17.51 air leakage, 17.48 – 17.50 critical path, 17.48 design procedure, 17.51 – 17.52 design velocity, 17.45 – 17.46 duct layout, 17.52 – 17.53 duct system characteristics, 17.52 ductwork installation, 17.50 fire protection 17.50 – 17.51 optimal air duct design, 17.43 – 17.45 sealing requirements of ASHRAE Standard 90.1 – 1999, 17.49 – 17.50 shapes and material of air ducts, 17.50 system balancing, 17.46 – 17.47 Air expansion refrigeration cycle, 9.45 – 9.49 fl w processes, 9.47 – 9.48 thermodynamic principle, 9.45 – 9.47 Air filters 15.64 – 15.68 classification of 15.65 coarse, 15.65 filter installation 24.7 – 24.8 filtration mechanism 15.64 – 15.65 high-efficien y, 15.66 – 15.67 low-efficien y, 15.65 – 15.66 medium-efficien y, 15.66 – 15.67 service life, 24.7 ultrahigh-efficien y, HEPA and ULPA filters 15.68 Air filters rating and assessments, 15.61 – 15.62 dust-holding capacity, 15.62 efficien y, 15.61 pressure drop, 15.61 – 15.62 service life, 15.62 Air filters test methods, 15.62 – 15.64 composite efficien y curves, 15.63 – 15.64 di-octylphthalate (DOP), 15.62 – 15.63 dust spot, 15.62 minimum efficien y reporting values (MERVs), 15.64 – 15.65 penetration, 15.63 removal efficien y by particle size, 15.63 selection, 15.71 – 15.72 test unit, 15.64 weight arrestance, 15.62 Air filters to rem ve contaminants, 24.6 – 24.8 filter selection for I Q, 24.6 – 24.7 remove indoor air contaminants, 24.6

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Air filtration and industrial air cleaning 15.60 – 15.61 Air fl w, basics, 17.2 – 17.8 Bernoulli equation, 17.2 equation of continuity, 17.7 – 17.8 laminar fl w and turbulent fl w, 17.6 – 17.7 pressure, 17.3 stack effect, 17.5 – 17.6 static pressure, 17.3 – 17.4 steady fl w energy equation, 17.2 – 17.3 total pressure, 17.5 velocity distribution, 17.3 velocity pressure, 17.4 – 17.5 Air fl w, characteristics, 17.8 – 17.10 air duct, types, 17.8 pressure characteristics, 17.8 – 17.10 static regain, 17.9 system pressure loss,17.10 Air-handling units, 1.8, 16.1 – 16.12 casing, 16.4 classification of 16.2 – 16.4 coil face velocity, 16.8 – 16.9 coils, 16.5 component layout, 16.6 – 16.8 controls, 16.6 draw-through or blow-through unit, 16.2 exhaust section, 16.6 factory fabricated or field- uilt AHU, 16.3 fans, 16.4 – 16.5 filters 16.5 functions of, 16.1 – 16.2 horizontal or vertical unit, 16.2 humidifiers 16.5 – 16.6 mixing, 16.6 – 16.7 outdoor air intake, 16.6 outdoor air (makeup air) or mixing AHU, 16.2 selection, 16.9 – 16.12 single zone or multizone, 16.2 – 16.3 rooftop or indoor AHU, 16.4 Air jets, 18.5 – 18.11 Archimedes number, 18.11 centerline velocities, 18.8 – 18.9 characteristic length, 18.8 confined 18.8 – 18.10 confined airfl w pattern, 18.9 – 18.10 core zone, 18.5 entrainment ratio, 18.7 envelope, 18.5 free isothermal, 18.5 – 18.7 free nonisothermal, 18.10 – 18.11 main zone, 18.6 surface effect, 18.8 terminal zone, 18.6

Air jets (Cont.) throw, 18.7 transition zone, 18.6, velocity profile 18.6 Air movements, 4.20 – 4.23 Air systems, 1.6 – 1.8, 20.2 – 20.4 air conditioning rules, 20.63 air distribution system, 20.3 air economizer mode, 22.5 air-handling system, 20.2 classification 20.39 constant volume systems, 20.40 – 20.41 cooling and heating mode, 22.4 mechanical ventilation system, 20.3 minimum outdoor air recirculating mode, 22.5 mixing-exhaust section, 22.8 occupied and unoccupied mode, 22.5 operating modes, 22.4 – 22.5 part-load operation, 22.4 – 22.5 purge-mode, 22.5 regenerative systems, 20.3 – 20.4 reheating, recooling, and mixing, 20.74 – 20.75 smoke control systems, 20.4 terminals, 20.4 ventilation systems, 20.3 warmup, colddown, and nighttime setback mode, 22.5 Air temperature: comfort air conditioning systems, 4.20 – 4.21 indoor, 4.20 – 4.23 processing air conditioning systems, 4.23 Air washer, 1.11 Amplifiers 2.7 Annual energy use, HVAC&R systems, 1.14 Artificial intelligence 5.45 – 5.53 Artificial neural net orks (ANN), 5.50 – 5.53 learning method, 5.52 – 5.53 neuron, 5.51 neuron activation transfer 5.51 – 5.52 net topology, 5.51 ASHRAE/IESNA Standard 90.1 – 1999, building envelope trade-off option, 3.50 compliance for building envelope, 3.48 – 3.50 controls, 5.66 – 5.67 off-hour controls, 5.66 – 5.67 Atmospheric dust, 15.61 Atmospheric extinction coefficient 3.26 Automated computer-aided drafting (AutoCAD), 1.26

I.3

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Bernoulli equation, 17.2 Boilers, hot water, 8.9 – 8.15 cast-iron sectional, 8.12 chimney or stack, 8.14 combustion efficien y, 8.13 condensing and noncondensing , 8.13 electric, 8.17 fire-tube 8.10 fl w processes, 8.10 – 8.12 forced-draft arrangements, 8.12 gas and oil burners, 8.13 heating capacity control, 8.14 minimum efficien y requirements, 8.13 – 8.14 safety control, 8.14 – 8.15 Scotch Marine packaged boiler, 8.10 – 8.12 selection of fuel, 8.9 – 8.10 types of, 8.10 Boiling point, 2.4 – 2.5 Building: energy star, 25.10 green, 25.8 – 25.10 shell building, 3.48 speculative building, 3.48 Building automation and control network (BACnet), 5.41 Building automation systems, 5.2 Building envelope, 3.2 ceiling, 3.2 energy-efficient and cost-e fective measures, 3.50 – 3.51 exterior floo , 3.2 exterior wall, 3.2 fenestration, 3.2 partition wall, 3.2 roof, 3.2 skylight, 3.2 slab on grade, 3.2 Standard 90.1 – 1999, 3.48 – 3.50 wall below grade, 3.2 window, 3.2 Building material: closed-cell, 3.16 open-cell, 3.13 Building tightness, or building air leakage, 20.5 – 20.6 air change per hour at 50 Pa (ACH50), 20.6 effective leakage area, 20.5 exfiltration 20.14 fl w coefficient fl w, in cfm/ft2, 20.6 infiltration 20.14 volume fl w rate of infiltration 20.14

Campus-type water systems, 7.53 – 7.58 building entrance, 7.56 control of variable-speed distribution pump, 7.56 distribution pipes, 7.58 multiple-source distributed building loop, 7.57 – 7.58 plant-distributed building loop, 7.56 – 7.57 plant-distribution building loop, 7.54 – 7.56 pressure gradient of distribution loop, 7.54 Carbon adsorbers, activated, 15.70 – 15.71 reactivation, 15.71 Cascade systems, 9.40 – 9.43 advantages and disadvantages, 9.40 – 9.41 performance, 9.42 – 9.43 Central plant, 1.8 – 1.9 Central systems, 30.2 air and water temperature differentials, 30.5 – 30.6 control at part load, 30.4 controls in water, heating, and refrigerating systems, 30.4 floo -by-floor systems vs. air systems servin many floors 30.2 – 30.3 influence of inlet anes on small centrifugal fans, 30.5 – 30.7 separate air system, 30.2 – 30.3 size of air system, 30.2 types of VAV central systems, 30.7 Central systems, clean-room, 30.14 – 30.24 airfl w, 30.14 – 30.16 case-study: integrated-circuit fabrication, 30.16 – 30.24 design considerations, 30.24 effect of filter pressure drop di ference on system performance, 30.22 – 30.24 energy use of components, 30.17 indoor requirements, 30.16 – 30.17 operating characteristics, 30.18 – 30.19 part-load operation and controls, 30.19 – 30.20 pressurization, 30.16 summer mode operation, 30.19 system characteristics, 30.13 system description, 30.14 – 30.15, 30.17 – 30.18 system pressure, 30.21 temperature and relative humidities, 30.16 winter mode operation and controls, 30.20 – 30.21 Central systems, dual-duct VAV, 30.10 – 30.11 system characteristics, 30.8 system description, 30.10 – 30.11

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Central systems, fan-powered VAV, 30.11 – 30.13 case-study: Taipei World Trade Center, 30.12 – 30.13 supply volume fl w rate and coil load, 30.11 system characteristics, 30.13 system description, 30.11 Central systems, single zone VAV, 30.7 – 30.9 supply volume fl w rate and coil load, 30.7 – 30.8 system characteristics, 30.8 system description, 30.7 zone temperature control, 30.8 Central systems, VAV cooling, VAV reheat, and perimeter-heating VAV, 30.9 – 30.10 supply volume fl w rate and coil load, 30.9 system characteristics, 30.8 system description, 30.9 zone temperature control, 30.10 Centrifugal chiller, 1.12 air purge, 13.24 auxiliary condenser, 13.9 – 13.11 capacity control, 13.19 – 13.21 capacity control by variable speed, 13.20 capacity control using inlet vanes, 13.20 chilled water leaving temperature control, 13.22 comparison between inlet vanes and variable speed, 13.21 condenser water temperature control, 13.23 controls, 13.22 – 13.24 difference between centrifugal compressors and fans, 13.19 double-bundle condenser, 13.9 – 13.10 evaporating and condensing temperatures at part-load, 13.26 – 13.27 faults detection and diagnostics, 13.24 functional controls and optimizing controls, 13.22 incorporating heat recovery, 13.9 – 13.13 operating characteristics, 13.24 – 13.35 operating modes, 13.9 – 13.11 part-load operation, 13.25 – 13.27 part-load operation characteristics, 13.25 – 13.26 performance rating conditions, 13.8 – 13.9 refrigerant fl w, 13.7 – 13.8 required system head at part-load operation, 13.19 – 13.20 safety controls, 13.23 – 13.24 sequence of operations, 13.24 – 13.25 short-cycling protection, 13.23 surge protection, 13.24

Centrifugal chiller (Cont.) system balance at full load, 13.25 system characteristics, 13.12 – 13.13 system description, 13.9 temperature lift at part-load, 13.29 – 13.31 water-cooled, 13.7 – 13.9 Centrifugal chiller, multiple-chiller plant, 13.33 – 13.36 chiller staging, 13.34 design considerations, 13.35 – 13.36 parallel and series piping, 13.33 – 13.34 Standard 90.1 – 1999 minimum efficien y requirements, 13.35 Centrifugal compressor: performance map,13.15 – 13.18 surge of, 13.15 – 13.16 Centrifugal compressor map: at constant speed, 13.16 – 13.18 at variable speed, 13.17 – 13.18 Centrifugal pumps, 7.30 – 7.34 cavitation, 7.33 net positive suction head (NPSH), 7.33 net static head, 7.32 performance curves, 7.32 – 7.33 pump efficien y, 7.32 pump power, 7.32 selection, 7.33 – 7.34 total head, 7.30 – 7.32 volume fl w, 7.30 Centrifugal refrigeration systems, 13.1 – 13.7 compressor, 13.3 – 13.4 free refrigeration, 13.31 – 13.33 free refrigeration, principle of operation, 13.31 – 13.32 free refrigeration capacity, 13.32 – 13.33 purge unit, 13.5 – 13.7 refrigerants, 13.2 – 13.3 system components, 13.4 – 13.5 Chilled-water storage systems, stratified 31.18 – 31.23 basics, 31.18 – 31.19 case-study, 31.23 – 28 charging and discharging, 31.18, 31.26 – 31.27 charging and discharging temperature, 31.22 – 31.23 chilled water storage system, 31.23 – 31.25 concentric double-octagon diffusers, 31.24 – 31.26 diffusers, 31.20 – 31.22 figure of merit 31.19 inlet Reynolds number, 31.21 – 31.22 part-load operation, 31.27 – 31.28

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Chilled-water storage systems, stratified (Cont.) self-balancing, 31.22 storage tanks, 31.19 stratified tanks 31.19 – 31.20 system characteristics, 31.10 system description, 31.18 system performance, 31.28 thermocline and temperature gradient, 31.20 – 31.21 Chlorofluorocarbons (CFCs) 1.12 Clean room, 4.31 Clean space, 4.31 Clearness number of sky, 3.26 Clothing: efficien y, 4.6 insulation, 4.7 CLTD/SCL/CLF method of cooling load calculation, 6.26 – 6.32 exterior walls and roofs, 6.26 – 6.28 fenestration, 6.28 infiltration 6.31 internal loads, 6.29 – 6.31 night shutdown mode, 6.32 wall exposed to unconditioned space, 6.28 – 6.29 Codes and standards, 1.23 – 1.25 Cogeneration, 12.25 – 12.26 using a gas turbine, 12.28 – 12.29 Coil accessories, 15.56 – 15.57 air stratification 15.58 – 15.59 air vents, 15.56 coil cleanliness, 15.57 coil freeze protection, 15.58 – 15.60 condensate collection and drain system, 15.57 – 15.58 condensate drain line, 15.58 condensate trap, 15.58 drain pan, 15.58 Coil characteristics, 15.32 – 15.39 coil construction parameters, 10.3 – 10.4 contact conductance, 15.37 – 15.39 direct-expansion (DX), 15.33 fins 15.33 – 15.37 interference, 15.38 steam heating, 15.33 types of, 15.33 – 15.34 water circuits, 15.38 – 15.39 water cooling, 15.33 water heating, 15.33 Coils, DX (wet coils), 10.2 – 10.10 (See also DX coils) Coils, sensible cooling and heating (dry coils), 15.39 – 15.48 Chilton-Colburn j-factor, 15.41

Coils, sensible cooling and heating (dry coils) (Cont.) effectiveness , 15.42 fin e ficien y f , 15.41 – 15.42 fin sur ace efficien y s , 15.41 fluid elocity and pressure drop, 15.44 heat transfer in sensible cooling process, 15.39 – 15.41 heating coils, 15.44 JP parameter, 15.41 number of transfer units (NTU), 15.43 part-load operation, 15.44 surface heat transfer coefficients 15.41 – 15.42 Coils, water cooling (dry-wet coils), 15.48 – 15.52 dry-part, 15.50 dry-wet boundary, 15.48 – 15.49 part-load operation, 15.50 – 15.51 selection, 15.51 – 15.52 wet-part, 15.50 Cold air distribution, 18.28 – 18.30 case-study, Florida Elementary School, 18.29 characteristics, 18.29 vs. conventional air distribution, 18.28 with fan-powered VAV boxes, 18.30 high induction nozzle diffusers, 18.28 – 18.29 performance of ceiling and slot diffusers, 18.29 – 18.30 surface condensation, 18.30 Commissioning, 32.1 cost of HVAC&R commissioning, 32.5 necessity of HVAC&R commissioning, 32.1 – 32.2 scope of, 32.2 – 32.3 team of HVAC&R commissioning, 32.4 when to perform, 32.4 – 32.5 Compound systems with flash cooler coefficient of performance 9.33, 9.38 coil core surface area Fs , 15.40 enthalpy of vapor mixture, 9.32 – 9.33 fl w processes, 9.31 fraction of evaporated refrigerant in flas cooler, 9.31 – 9.32, 9.35 – 9.37 three-stage, 9.35 – 9.38 two-stage, 9.31 – 9.33 Compound system with vertical intercooler, two-stage, 9.38 – 9.40 comparison between flash coolers and inter coolers, 9.40 Compressibility factor, 2.2 – 2.3 Compressors, reciprocating, 11.5

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Computational fluid dynamics (CFD) 18.51 – 18.54 conducting CFD experiments, 18.54 numerical methods, 18.52 – 18.53 Reynolds-averaged Navier-Stokes equations, 18.52 velocity vectors of the airfl w in a duct section, 18.53 Computer-aided design, 1.25 – 1.26 Computer-aided design and drafting (CADD), 1.25 – 1.26 Computer-aided design and interface, 17.73 Computer-aided drafting, 1.26 Computer-aided duct design and drafting, 17.72 – 17.73 Computer-aided duct drafting, 17.72 Computer-aided running processes of duct system, 19.73 Computer-aided schedules and layering, 17.72 – 17.73 Computer-aided piping design and drafting, 7.58 – 7.60 computer-aided design capabilities, 7.59 – 7.60 computer-aided drafting capabilities, 7.58 – 7.59 input data and reports, 7.60 pressure losses and network technique, 7.59 pump and system operations, 7.59 system and pipe size, 7.59 Condensation: in buildings, 3.17 – 3.18 concealed condensation in building envelopes, 3.18 visible surface, 3.17 – 3.18 Condensation process, 10.20 – 10.21 heat rejection factor, 10.21 – 10.22 total heat rejection, 10.21 – 10.22 Condensers, 10.20 – 10.36 automatic brush cleaning for, 13.13 – 13.15 effect of brush cleaning system,13.14 – 13.15 principle and operation, 13.13 – 13.14 type of, 10.22 Condensers, air-cooled, 10.26 – 10.30 clearance, 10.29 condenser temperature difference, 10.28 – 10.29 condensing temperature, 10.29, construction, 10.26 – 10.28 cooling air temperature rise, 10.28 dirt clogging, 10.29 heat transfer process, 10.26 – 10.28 low ambient control, 10.29 – 10.30

Condensers, air-cooled (Cont.) oil effect, 10.29 selections, 10.30 subcooling, 10.29 volume fl w, 10.28 warm air circulation, 10.29 Condensers, evaporative, 10.30 – 10.33 condensation process, 10.30 cooling air, 10.32 heat transfer, 10.30 – 10.32 low ambient air control, 10.33 selection and installations, 10.33 site location, 10.32 – 10.33 water spraying, 10.32 Condensers, water-cooled, 10.22 – 10.26 capacity, 10.26 double-tube condenser, 10.22 – 10.23 effect of oil, 10.25 heat transfer, 10.24 – 10.25 part-load operation, 10.26 performance, 10.25 – 10.26 shell-and-tube condensers, 10.22 – 10.25 subcooling, 10.25 types of, 10.22 Conduit induction system, 1.11 Constant-volume multizone system with reheat, 20.74 – 20.78 control systems, 20.75 – 20.76 operating parameters and calculation, 20.76 – 20.78 reheating, recooling and mixing, 20.74 – 20.75 system characteristics, 20.78 Constant-volume single-zone systems, cooling mode operation, 20.53 – 20.59 air conditioning cycle, cooling mode operation, 20.53 – 20.54 cooling mode operation in summer, 20.53 – 20.56 cooling mode operation in winter with space humidity control, 20.55 – 57 cooling mode operation in winter without space humidity control, 20.55 – 57 outdoor ventilation air and exhaust fans, 20.58 – 20.59 part-load operation and controls, 20.58 two-position or cycling control, 20.58 water fl w rate modulation, 20.58 Constant-volume single-zone systems, heating mode operation, 20.69 – 20.74 dual-thermostat, year-round zone temperature control, 20.73 – 20.74 heating mode with space humidity control, 20.71 – 20.73

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Constant-volume single-zone systems, heating mode operation (Cont.) heating mode without space humidity control, 20.69 – 20.70 part-load operation, 20.73 Constant-volume systems, 20.40 – 20.41 energy per unit volume fl w, 20.41 system characteristics, 20.40 – 20.41 Control loop, 5.5 closed, 5.5 open, 5.5 Control medium, 5.11 Control methods, 5.7 – 5.9 comparison of, 5.8 – 5.9 direct-digital-control (DDC), 5.7 electric or electronic control, 5.7 – 5.8 pneumatic control, 5.7 Control modes, 5.9 – 5.16 compensation control or reset, 5.15 differential, 5.9 floating control 5.11 modulation control, 5.10 offset or deviation, 5.13 proportional band, 5.12 proportional control, 5.11 – 5.13 proportional-integral-derivative (PID) control, 5.14 – 5.15 proportional plus integral (PI) control, 5.13 – 5.14 step-control, 5.10 – 5.11 throttling range, 5.12 two-position, 5.9 – 5.10 Control systems, 5.2 direct digital control (DDC), 1.9 dual-thermostat year-round zone temperature control, 20.73 – 20.74 Control valves, 5.26 – 5.31, actuators, 5.26 – 5.27 equal-percentage, 5.28 fl w coefficient 5.31 linear, 5.28 quick-opening, 5.29 rangeability, 5.29 three-way, 5.27 two-way, 5.27 Controlled device, 5.5 Controlled variable, 5.2 Controllers, 5.21 – 5.26 direct-acting and reverse-acting, 5.21 – 5.22 direct digital, 5.23 – 5.26 electric and electronic, 5.23 electric erasable programmable read-only memory (EEPROM), 5.24

Controllers (Cont.) flash erasable programmable read-only mem ory (flash EP OM), 5.25 normally closed or normally open, 5.22 pneumatic, 5.22 – 5.23 random-access memory (RAM), 5.24 read-only memory (ROM), 5.23 system, 5.23 – 5.26, 5.38 – 5.39 unit, 5.23 – 5.26, 5.39 Controls: alarming, 5.60 discriminator, 5.60 functional, 5.58 – 5.61 generic, 5.59 – 5.60 graphical displays, 5.59 scheduling, 5.59 – 5.60 specific 5.60 – 5.61 trending, 5.59 Cooling coil load, 6.32 – 6.34 duct heat gain, 6.33 fan power, 6.33 temperature of plenum air, 6.34 ventilation load, 6.34 Cooling coil load, components, 6.7 – 6.8 Cooling load: components, 6.6 – 6.7 external, 6.7 internal, 6.7 Cooling load calculations: historical development, 6.11 – 6.12 heat balance, 6.12 – 6.14 transfer function, 6.14 – 6.16 Cooling media, 9.3 Cooling towers, 10.34 – 10.36 approach, 10.36, 10.41 blowdown, 10.36 construction materials, 10.43 counterfl w forced draft, 10.35 – 10.36 counterfl w induced draft, 10.34 – 10.35 crossfl w induced draft, 10.34 – 10.35 factors affecting performance, 10.40 fill configuratio 10.42 – 10.43 heat and mass transfer process, 10.37 – 10.39 makeup, 10.36 optimum control, 10.43 – 10.44 outdoor wet-bulb temperature, 10.41 part-load operation, 10.43 performance, 10.40 – 10.43 range, 10.36, 10.40 thermal analysis, 10.36 – 10.39 tower capacity, size, 10.37 – 10.39 tower coefficient (NTU) 10.36 – 10.39, 10.41 water-air ratio, 10.41

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Cooling towers (Cont.) water circulating rate, 10.40 water distribution, 10.43 Cooling towers, operating considerations, 10.46 – 10.48 blowdown, 10.47 fogging, 10.46 – 10.47 freeze protection, 10.46 interference, 10.46 Legionnaires’ disease, 10.47 maintenance, 10.47 – 10.48 recirculation, 10.46 Coordination, 1.19 Copenhagen Amendments and Vienna Meeting, 9.10 – 9.11 Corrosion, 7.25 Daily range, mean, 4.39 Dalton’s law, 2.3 – 2.4 Dampers, 5.32 – 5.38 actuators, 5.33 butterfl , 5.32 characteristic ratio, 5.35 – 5.37 gate, 5.32 opposed-blade, 5.33, 5.35 – 5.37 parallel-blade, 5.33 , 5.35 – 5.37 sizing, 5.37 – 5.38 split, 5.32 – 33 DDC programming, 5.53 – 5.55 evolution, 5.53 graphical, 5.53 – 5.54 for mechanical cooling control, 5.55 templates, 5.54 DDC tuning controllers, 5.55 – 5.56 adaptive control, 5.56 PI controllers, 5.55 self-tuning, 5.55 Degree days: cooling with a base temperature of 50 °F, 4.39 heating with a base temperature of 65 °F, 4.39 number of, 4.39 Degree of saturation, 2.8 Demand-controlled ventilation (DCV), CO2based, 23.5 – 23.12 application of, 23.11 – 23.12 ASHRAE Standard 62 – 1999, 23.7 base ventilation, 23.9 – 23.10 CO2-based DCV system, 23.10 – 23.11 CO2 sensor or mixed-gases sensor, 23.7 location of CO2 sensor, 23.7 – 23.8 minimum outdoor air recirculation mode, 23.6

Demand-controlled ventilation (DCV), CO2based (Cont.) purge mode, 23.10 substantial lag time in space CO2 concentration dilution process, 23.8 – 23.8 vs. time-based constant-volume control, 23.5 – 23.6 Depletion of the ozone layer, 1.15 Desiccant-based air conditioning systems, 29.22 – 29.27 applications, 29.34 – 29.35 conditions to apply, 29.34 – 29.35 desiccant dehumidification and sensible cool ing, 29.22 – 29.24 desiccants, 29.24 – 29.26 lithium chloride, 29.26 molecular sieves, 29.26 – 29.27 rotary desiccant dehumidifiers 29.27 silica gel, 29.26 system characteristics, 29.21 Desiccant-based air conditioning systems, for operating rooms, 29.32 – 29.34 indoor environment, 29.32 – 29.33 system description, 29.33 – 29.34 Desiccant-based air conditioning systems, for retail store, 29.31 – 32 operating characteristics, 29.31 – 29.32 performance, 29.32 system description, 29.31 – 29.32 Desiccant-based air conditioning systems, for supermarket, 29.27 – 29.31 air conditioning cycle, 29.30 – 29.31 gas heater, 29.30 heat-pipe heat exchanger, 29.29 – 29.30 indirect evaporative cooler, 29.30 loads in supermarkets, 29.27 operating parameters in rotary desiccant dehumidifie , 29.29 part-load operation and controls, 29.31 refrigeration, 29.30, space conditioning line, 29.28 – 29.29 system description, 29.25, 29.28 of the control systems, 1.20 – 1.21 Design documents, 1.21 – 1.22 Design-bid, 1.17 Design-build, 1.17 Design intent, 32.1 Desorption isotherm, 3.11 Diagram: pressure-enthalpy, 9.17 – 9.18 temperature-entropy, 9.18 – 9.19 Direct expansion (DX) coil, 1.4

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Discharge air temperature controls, 23.18 – 23.23 basics, 23.18 discharge air temperature reset, 23.22 – 23.23 operation of air economizer, 23.21 – 23.22 outdoor air intake, 23.21 – 23.22 system description, 23.19 – 23.21 Distribution of systems usage, 1.10 Diversity factor, 1.20 Drawings, 1.22 air duct diagram, 1.22 control diagrams, 1.22 detail, 1.22 equipment schedule, 1.22 floor plans 1.22 legends, 1.22 piping diagram, 1.22 sections and elevations, 1.22 Duct cleaning, 17.74 – 17.75 Duct construction, 17.12 – 17.18 duct hanger spacing, 17.17 fibe glass ducts, 17.18 flame speed and smo e developed, 17.13 flat val ducts, 17.17 – 17.18 fl xible ducts, 17.18 material, 17.12 – 17.13 maximum pressure difference, 17.12 rectangular ducts, 17.13 rectangular metal duct construction, 17.15 round ducts, 17.17 thickness of galvanized sheets, 17.14, 17.17 transverse joint reinforcement,17.16 Duct friction losses, 17.22 – 17.31 absolute and relative roughness, 17.22 – 17.24 circular equivalents, 17.27 – 17.31 Colebrook formula, 17.24 Darcey-Weisbach equation, 17.22 duct friction chart, 17.24 – 17.26 17.25 – 17.26 duct roughness, 17.25 friction factor, 17.22 – 17.24 Moody diagram, 17.22 – 17.23 roughness and temperature corrections, 17.25 Rouse limit, 17.24 Swamee and Jain formula, 17.24 Duct insulation, 17.19 – 17.22 duct insulation by ASHRAE Standard 90.1 – 1999,17.19 – 17.21 temperature rise and drop, 17.19 temperature rise curves, 17.21 – 17.22 Duct liner, 17.74 Duct sizing methods, 17.53 – 17.56 constant velocity method, 19.53 – 19.54 equal friction method, 17.53

Duct sizing methods (Cont.) static regain method, 17.54 – 17.55 T-method, 17.55 – 17.56 Duct static pressure and fan controls, 23.23 – 23.26 comparison between adjustable-frequency drives and inlet vanes, 23.24 – 23.26 duct static pressure control, 23.23 – 23.24 sensor’s location, 23.24 set point, 23.24 Duct systems with certain pressure losses in branch takeoffs, 17.56 – 17.66 condensing two duct sections, 17.59 – 17.60 cost optimization, 17.56 – 17.59 design characteristics, 17.56 local loss coefficients for d verging tees and wyes, 17.60 – 17.62 return or exhaust duct systems, 17.63 Duct systems with negligible pressure loss at branch ducts, 17.66 – 17.72 local loss coefficients 17.68 – 17.69 pressure characteristics of airfl w in supply ducts, 17.66 – 17.68 rectangular supply duct with transversal slots, 17.67 return or exhaust duct systems, 17.71 – 17.72 supply duct systems, 17.66 DX coils, wet coils, 10.2 – 10.10 air-side pressure drop, 10.8 construction and installation, 10.3 – 10.4 DX coil effectiveness, 10.6 – 10.7 face velocity, 10.7 – 10.8 part-load operation, 10.8 – 10.10 selection of DX coils, 10.10 simultaneous heat and mass transfer, 10.5 – 10.6 superheated region, 10.5 two-phase region, 10.4 – 10.5 two-region model,10.4 – 10.5 Dynamic losses, 17.31 – 17.38 converging and diverging tees and wyes, 17.34 – 17.37 elbows, 17.31 – 17.34 entrances, exits, enlargements, and contractions, 17.38 Earth-sun distance, 3.25 Economizer cycle, economizers, and economizer control, 21.8 – 21.16 air economizers, 21.8 ANSI/ASHRAE Standard 90.1 – 1999 economizer control specifications 21.14 – 21.16

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Economizer cycle, economizers, and economizer control (Cont.) comparison of air and water economizers, 21.14 comparison of enthalpy-based and temperature-based, 21.10 – 21.12 differential enthalpy, electronic enthalpy, and fi ed enthalpy, 21.8 – 21.9 enthalpy (-based) economizer control, 21.8 – 21.9 fi ed dry-bulb and differential dry bulb, 21.9 – 21.10 sequence of operations of a differential drybulb, 21.10 sequence of operations of a differential enthalpy, 21.9 water economizer, 21.8, water economizer control, 21.12 – 21.14 Effective temperature, 4.14 Electric heating fundamentals, 8.15 – 8.16 electric duct heaters, 8.17 electric furnaces and electric heaters, 8.16 – 8.17 Electricity deregulation, 25.14 – 25.15 California approach, 25.15 case-study: automatic control of RTP, 25.16 – 25.17 prior to deregulation, 25.14 real-time pricing (RTP), 25.15 – 25.16 Energy conservation measures, 25.10 – 25.11 case-study-for an office 25.12 Energy cost budget method, ASHRAE/IESNA Standard 90.1 – 1999, 25.28 Energy efficien y, 1.13 – 1.15, 25.1 – 25.2, 25.5 25.10 during design, construction, commissioning, and operation, 25.2 energy audits, 25.6 energy retrofits 25.6 – 25.7 energy service companies (ESCOs), 25.7 federal mandates, 25.5 performance contracting, 25.7 – 25.8 reduction of unit energy rate, 25.2 – 25.3 Energy management and control systems (EMCS), 5.3 Energy management systems, 5.3 Energy use (energy consumption), 1.13 – 1.15, 25.1 – 25.2 between HVAC&R system characteristics, 25.12 – 25.13 building energy consumption and thermal storage systems, 31.2 fan, motor, and drive combined efficien y, 25.13 – 25.14

Energy use (energy consumption) (Cont.) heating-cooling equipment, 25.13 Energy use, index, 9.55 – 9.55 energy efficien y ratio (EER), 9.55 energy use intensities, 25.5 – 25.6 heating season performance factor (HSPF), 9.55 integrated part-load value (IPLV), 9.56 kW/ton, 9.55 – 9.56 seasonal energy efficien y ratio (SEER), 9.56 Engineering responsibilities, 1.18 – 1.19 Engineer’s quality control, 1.20 Environment: cleanest, 1.13 most precise, 1.13 quietest, 1.13 Environmental problems, 1.15 Equation of state: of an ideal gas, 2.2 of a real gas, 2.2 Evaporative coolers, add-on, 27.18 – 27.24 indirect-direct cooler to a DX packaged system, 27.18 – 27.20 tower and coil combination, 27.22 – 27.23 tower coil and rotary wheel combination, 27.20 – 27.22 Evaporative cooling, 27.1 air washers, 27.4 direct, 27.2 direct evaporative coolers, 27.3 – 27.4 evaporative pads, 27.4 operating characteristics, 27.6 rigid media, 27.4 rotary wheel, 27.4 – 27.6 saturation efficien y, 27.2 – 27.4 Evaporative cooling, indirect, 27.6 – 27.13 effectiveness, 27.10 – 27.11 heat transfer process, 27.7 – 27.10 operating characteristics, 27.11 – 27.12 part-load operation and control, 27.12 – 27.13 process, 27.6 Evaporative cooling, indirect-direct two-stage systems, 27.13 – 27.18 case study: Nevada’s College, 27.16 – 27.18 energy efficien y ratio and energy use intensities, 27.16 indirect-direct two-stage evaporative cooler, 27.13 – 27.15 system characteristics, 27.17 – 27.18 using outdoor air as cooled and wet air, 27.15 using return air as wet air and outdoor-return air mixture as cooled air, 27.15 – 27.16

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Evaporative cooling systems, 27.1 – 27.2 beware of dampness, sump maintenance, and water leakage, 27.24 design considerations, 27.24 – 27.26 scope of applications, 27.24 selection of summer outdoor design conditions, 27.24 – 27.26 Evaporative heat loss, 4.7 – 4.9 diffusion, 4.8 – 4.9 maximum, 4.7 – 4.8 due to regulatory sweating, 4.7 – 4.8 respiration losses, 4.7 from skin surface, 4.7 Evaporators, 10.2 – 10.20 air-cooler, 10.2 circulating rate, 10.20 counterfl w or parallel fl w, 10.20 direct-expansion liquid cooler, 10.18 down-feed or up-feed, 10.20 DX coil (wet coils) 10.2 – 10.10 flooded liquid coole , 10.12 – 1020 liquid cooler, 10.2 liquid overfeed cooler, 10.18 – 10.20 mechanical pump or gas pump, 10.20 Energy, 9.19 Expansion tank: closed, 7.21 diaphragm, 7.21 – 23 fill pressure 7.21 open, 7.20 – 7.21 water logging, 7.24 – 7.25 Factors affecting control processes, 5.56 – 5.58 climate change, 5.56 – 5.57 disturbance, 5.57 intermittent operation, 5.57 load, 5.56 performance of control processes, 5.57 – 5.58 system capacity, 5.57 thermal capacitance, 5.58 turndown ratio, 5.57 Fan capacity modulation, 15.20 – 15.24 ac inverter, 15.20 – 15.21 adjustable pitch, 15.24 blade pitch, 15.24 controllable pitch, 15.24 fan speed with adjustable frequency drives, 15.20 – 15.21 inlet cone, 15.23 – 15.24 inlet-vanes, 15.21 – 15.23 pulse-width-modulated inverter, 15.21 variable-speed drives (VSDs), 15.20 – 15.21 Fan coil, 1.5

Fan coil systems, 28.3 – 28.5 operating characteristics, 28.3 – 28.5 system description, 28.3 Fan coil systems, four-pipe, 28.9 – 28.15 chilled water supplied to coils, 28.11 – 28.12 dedicated ventilation system, 28.10 – 28.11 exhaust air to balance outdoor ventilation air, 28.12 general description, 28.9 – 28.10 operating parameters, 28.14 – 28.19 part-load operation, 28.13 space recirculation systems, 28.11 system characteristics, 28.14 – 28.15 zone temperature control and sequence of operations, 28.13 – 28.14 Fan coil systems, two-pipe, 28.20 – 28.24 applications, 28.24 changeover two-pipe systems, 28.23 – 28.24 nonchangeover two-pipe systems, 28.20 – 28.23 system characteristics, 28.15 Fan coil units, 28.5 – 28.9 coils, 28.7 cooling and dehumidifying, 28.8 – 28.9 fan, 28.6 – 28.7 filters 28.7 sound power level, 28.9 volume fl w rate, 28.7 – 28.8 Fan combinations, 22.4 operating modes, 22.4 – 22.5 Fan combinations, supply and exhaust fans, 22.8 – 22.14 air-economizer mode, 22.13 operating characteristics, 22.9 – 22.10 pressure variation at the mixing box, 22.13 – 22.14 recirculating mode and design volume fl w rate, 22.9 – 22.12 recirculating mode, 50% design fl w rate, 22.12 – 22.13 system characteristics, 22.8 – 22.9 warmup and colddown mode, 22.13 Fan combinations, supply and relief fans, 22.14 – 22.18 air economizer mode and design volume fl w rate, 22.14 – 22.16 air economizer mode, 50% design fl w, 22.17 design considerations and controls, 22.17 – 22.18 recirculating mode, 22.14 – 22.15 warmup and cool-down mode, 22.17 Fan combinations, supply and return fans, 22.18 – 22.21 air economizer mode, 22.20 – 22.21

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Fan combinations, supply and return fans (Cont.) comparison of three fan combination systems, 22.21 – 22.22 controls, 22.21 recirculating mode, 22.18 – 22.20 Fan construction and arrangements, 15.25 – 15.29 drive arrangements and direction of discharge, 15.26 – 15.28 high-temperature fans, 15.27 safety devices, 15.28 – 15.29 sizes and class standards, 15.25 – 15.26 spark-resistant construction, 15.28 width and inlets, 15.26 – 15.27 Fan-duct systems, 20.14 – 20.17 fan laws, Buckingham  method, 20.15 – 20.17 inlet system effect, 20.18 – 20.19 inlet system effect loss, 20.19 inlet system effect loss coefficient 20.19 – 20.20 outlet system effect, 20.20 – 20.22 outlet system effect loss coefficient 20.22 – 20.23 selecting fans considering system effect losses, 20.23 – 20.24 system effect, mechanism, 20.17, system operating point, 20.15 Fan-duct systems, combination, 20.24 – 20.31 connected in series, 20.25 – 20.26 fan combined in parallel and connected in series with a duct system, 20.26 – 20.27 two parallel fan-duct systems with another duct system, 20.28 – 20.30 Fan-duct systems, modulation, 20.31 – 20.38 blade pitch variation of axial fan, 20.35 – 20.36 modulation curve, 20.31 – 20.32 using dampers, 20.33 using inlet cone, 20.34 – 20.35 using inlet vanes, 20.34 varying fan speed, 20.35 – 20.36 Fan energy use, criteria of Standard 90.1 – 1999, 17.10 – 17.12 for constant volume systems, 17.10 – 17.11 for VAV systems, 17.11 – 17.12 Fan-powered VAV box, 1.8 Fan room, 16.24 – 16.28 isolated, 16.24 – 16.25 layout considerations, 16.25 – 16.28 open, 16.24 types of, 16.24 – 16.25 Fan selection, 15.29 – 15.32 case-study, 15.32

Fan selection (Cont.) comparison between various type of fans, 15.31 – 15.32 estimated fan sound power level, 15.30 – 15.31 Fans, fundamentals, 15.2 – 15.7 air temperature increase through fan, 15.5 blower, 15.2 compression ratio, 15.2 functions, 15.2 influence of el vation and temperature, 15.6 – 15.7 performance curves, 15.5 – 15.6 power and efficien y, 15.4 – 15.5 pressure, 15.4 types of, 15.2 – 15.3 volume fl w rate or capacity, 15.4 Fan stall, 15.24 – 15.25 Fan surge, 15.24 Fans, axial, 15.14 – 15.20 hub ratio, 15.14 – 15.15 number of blades, 15.20 performance curves, 15.17 – 15.19 power-volume fl w curves, 15.18 – 15.19 pressure-volume curves, 15.17 propeller, 15.15 reverse operation, 15.20 static pressure developed, 15.17 tip clearance, 15.20 total efficien y-volume fl w curves, 15.18 – 15.19 tube-axial, 15.15 – 15.16 typical vane-axial fan, 15.19 – 15.20 types of, 15.14 – 15.16 vane-axial, 15.15 – 15.16 velocity triangles, 15.16 – 15.17 Fans, centrifugal, 15.7 – 15.4 backward-curved, 15.8 – 15.10 blades, 15.7 blast area, 15.8 energy losses, 15.9 forward-curved, 15.11 – 15.12 impeller (fan wheel), 15.7 – 15.8 power-volume fl w curves, 15.10 – 15.11 pressure-volume curves, 15.9 radial-bladed, 15.10 – 15.12 roof ventilators, 15.14 total efficien y-volume fl w curves, 15.10 total pressure increase at fan impeller, 15.7 – 15.8 tubular or in-line, 15.12 – 15.13 unhoused plug/plenum,15.12 – 15.14 velocity triangles, 15.8 Fans, crossfl w, 15.3 – 15.4

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Fault detection and diagnostics, 5.61 – 5.65 ANN models, 3.64 ARX models, 5.63 – 5.64 comparison of ARX and ANN models, 5.65 expert systems rule-based, 5.62 – 5.63 system and component models, 5.64 Fenestration, 3.29 – 3.31 Fiberglass in HVAC&R systems, 19.17 – 19.18 problems, 19.17 – 19.18 recommendations, 19.18 Field experience, 1.21 Finite difference method, 6.34 – 6.39 cooling loads, 6.39 interior nodes, 6.36 – 6.37 simplify assumptions, 6.36 space air temperature, 6.38 – 6.39 surface nodes, 6.37 – 6.38 Flooded liquid cooler, 10.12 – 10.20 construction, 10.12 – 10.14 cooling capacity, 10.17 evaporating temperature, 10.16 fouling factor, 10.14 – 10.15 heat transfer, 10.14 oil effect, 10.17 part-load operation, 10.17 – 10.18 performance, 10.16 – 10.17 pool boiling and force convection model, 10.15 – 10.16 temperature difference Tee - Tel , 10.16 – 10.17 Flow resistance, 17.38 – 17.43 connected in parallel, 17.41 – 17.42 connected in series, 17.40 – 17.41 of duct system, 17.42 – 17.44 of Y-connection, 17.42 – 17.43 Flow sensors, 5.19 – 5.20 Fouling factor, 10.14 – 10.15 Fuzzy logic, 5.45 – 5.47 fuzzy logic controller, 5.47 fuzzy sets, 5.45 membership function, 5.45 production rules, 5.45 – 5.47 Gas cooling, 12.25 – 12.29 engine jacket heat recovery, 12.28 exhaust gas heat recovery, 12.27 – 12.28 gas-engine chiller, 12.25 – 12.27 gas engines, 12.27 Gaseous contaminants adsorbers and chemisorbers, 24.8 – 24.12 activated carbon adsorbers, 24.9 chemisorption, 24.11 chemisorption performance, 24.11

Gaseous contaminants adsorbers and chemisorbers (Cont.) granular activated carbon (GAC) applications, 24.10 – 24.11 granular activated carbon (GAC) performance, 24.9 – 24.10 indoor gaseous contaminants, 24.8 – 24.9 Gibbs-Dalton law, 2.4 Global radiation, 3.27 – 3.28 Global warming, 1.15, 25.3 – 25.5 CO2 release, 25.4 effect, 1.15 Kyoto Protocol, 25.3 mitigating measures, 25.4 – 25.5 refrigerant emissions, 25.4 – 25.5 total equivalent warming impact, 25.3 – 25.4 Goal to provide an HVAC&R system, 1.17 Green buildings, 25.8 – 25.10 basics, 25.8 – 25.9 case-studies, 25.9 – 25.10 green building assessment (GBA), 25.9 Greenhouse effect, 1.15 Heat: convective, 6.2 latent, 2.10 radiative, 6.2 sensible, 2.10 stored, 6.2 Heat capacity, 3.8 Heat of sorption, 3.12 Heat pipe heat exchangers, 12.23 – 12.24 Heat pump, 12.1 – 12.3 classification of 12.3 cycle, 12.2 – 12.3 Heat pump systems, air-source, 12.5 – 12.13 capacity and selection, 12.13 compressor, 12.6 – 12.7 controls, 12.13 cooling mode, 12.9 cycling loss and degradation factor, 12.11 defrosting, 12.12 – 12.13 heating mode, 12.9 indoor coil, 12.7 – 12.8 outdoor coil, 12.8 reversing valve, 12.7 – 12.8 Standard 90.1 – 1999 minimum efficien y requirements, 12.12 suction line accumulator,12.8 – 12.9 system performance, 12.9 – 12.11 Heat pump systems, ground-coupled and surface water, 12.17 – 12.19 Heat pump systems, groundwater, 12.13 – 12.17

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Heat pump systems, groundwater (Cont.) groundwater systems, 12.14 for hospital, 12.14 – 12.15 for residences, 12.15 – 12.16 Standard 90.1 – 1999 minimum efficien y requirements, 12.17 Heat recovery, air-to-air, 12.19 – 12.24 comparison between various heat exchangers, 12.24 effectiveness, 12.19 – 12.20 fi ed-plate heat exchangers, 12.20 – 12.21 heat pipe heat exchangers, 12.23 – 12.24 rotary heat exchangers, 12.12.21 – 12.23 runaround coil loops,12.21 types of, 12.19 Heat recovery systems, 12.3 – 12.5 heat balance and building load analysis, 12.4 – 12.5 Heat rejecting systems, 10.48 – 10.51 comparison between various systems, 10.48 – 10.50 Standard 90.1 – 1999, 10.50 – 10.51 types of, 10.48 Heat transfer: conductive, 3.3 – 3.4 convective, 3.4 – 3.5 fundamentals, 3.2 overall, 3.6 – 3.7 radiant, 3.5 – 3.6 Heat transfer coefficients 3.8 – 3.11 forced convection, 3.9 natural convection, 3.10 radiant, 3.8 – 3.9 surface, 3.10 – 3.11, 4.5 Heating load, 6.39 – 6.42 basic principles, 6.39 heat loss from products, 6.41 infiltration 6.41 latent heat loss, 6.41 night shutdown operation, 6.41 – 6.42 pickup load and oversizing factor, 6.42 setback, night, 6.41 – 6.42 transmission loss, 6.38 – 6.40 unheated spaces, 6.40 – 6.41 Heating systems, 8.1 – 8.2 control and operations of multizones, 8.30– 8.31 design considerations, 8.30 design nomograph, 8.30 low-pressure ducted warm air, 8.17 – 8.22 radiant floor panel 8.27 – 8.31 selection of, 8.2 system characteristics, 8.31 thermal characteristics of floor panel 8.28 – 8.29

Henry’s equation, 7.23 Hot water heating systems: design considerations, 8.25 – 8.26 finned-tube heaters 8.24 – 8.25 part-load operation and control, 8.26 two-pipe individual loop, 8.23 – 8.24 types of, 8.23 using finned-tube heaters 8.23 – 8.26 Humidifiers 15.72 – 15.85 humidifying load, 15.72 – 15.73 selection and design, 15.83 – 15.84 space relative humidity, 15.72 types of, 15.73 Humidifiers atomizing and wetted element, 15.76 – 15.78 air washers, 15.79 – 15.82 bypass control, 15.81 characteristics, 15.82 – 15.83 construction of air washer, 15.79 – 15.80 case study: White Plains ultrasonic project, 15.77 centrifugal atomizing, 15.77 – 15.78 functions of air washer, 15.80 humidification process 15.76 oversaturation, 15.81 performance of air washer, 15.80 – 15.81 pneumatic atomizing, 15.78 single-stage or multistage, 15.81 – 15.82 ultrasonic, 15.77 wetted element, 15.78 Humidifiers steam and heating element, 15.73 – 15.76 characteristics and requirements, 15.76 heating element, 15.75 steam grid, 15.73 – 15.74 steam humidifiers with separators 15.74 – 15.75 Humidity: comfort air conditioning systems, 4.23 – 4.24 process air conditioning systems, 4.24 Humidity ratio, 2.7 Humidity sensors, 5.18 – 5.19 HVAC&R industry, 1.15 h-w chart, 2.19 Hygrometers: capacitance, 2.17 – 2.18 Dunmore resistance, 2.16 – 2.17 electronic, 2.16 – 2.17 ion-exchange resistance, 2.16 – 2.17 mechanical, 2.16 Hysteresis, 3.11 – 3.12 Ice point, 2.4 – 2.5

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Ice storage systems: comparison of various systems, 31.17 – 31.18 types of, 31.5 Ice storage systems, encapsulated, 31.13 – 31.15 charging and discharging, 31.15 chiller priority and storage priority, 31.15 controls, 31.14 – 31.15 encapsulated ice containers, 31.13 location of chiller and storage tank, 31.14 system characteristics, 31.10 Ice storage systems, ice-harvesting, 31.15 – 31.17 chiller operation, 31.17 ice making or charging, 31.16 – 31.17 system characteristics, 31.10 system description, 31.15 – 31.16 Ice storage systems, ice-on-coil, external melt, 31.10 – 31.13 case-study, 31.13 ice builders, 31.11 ice-charging control, 31.11 refrigeration feed, 31.1 system characteristics, 31.10, 31.11 – 31.13 system description, 31.10 – 31.11 Ice storage systems, ice-on-coil, internal melt, 31.6 – 31.10 brine and glycol solution, 31.6 – 31.7 case-study: operation modes, 31.7 – 31.8 direct cooling, 31.9 ice-burning or ice melting, 31.9 ice-charging or ice making, 31.8 ice storage tank, 31.7 – 31.8 on-peak, 31.9 system characteristics, 31.9 – 31.10 system description, 31.6 Indicator, 2.6 Indoor air contaminants, 4.27 – 4.28 bioaerosols, 4.28 combustion products, 4.28 nicotine, 4.28 occupant-generated contaminants, 4.28 radon, 4.28 total particulates concentration, 4.28 volatile organic compounds, 4.28 Indoor air quality (IAQ), 4.27 acceptable, 4.29 basic strategies to improve, 4.29 IAQ problems, 24.1 – 24.2 IAQ procedure, 4.29 ventilation rate procedure, 4.29 – 4.31 Indoor design conditions, 4.1 – 4.2 Infrared heaters: electric, 8.32 – 8.33 gas, 8.32

Infrared heating, 8.31 – 8.35 basics, 8.31 – 8.32 beam radiant heaters, 8.32 design and layout, 8.33 – 8.35 Insufficient communication 1.17 Insulation material, 3.19 moisture content, 3.19 – 3.21 Interoperability, 5.41 system integration, 5.41 Knowledge-based systems (KBS), 5.47 – 5.51 development of KBS, 5.49 expert-systems, 5.47 – 5.51 knowledge acquisition, 5.49 knowledge-base, 5.48 inference engine, 5.48 testing, verification and validation, 5.49 user interface, 5.48 – 5.49 Legal responsibility for IAQ cases, 24.13 – 24.15 HVAC&R engineer, 24.14 – 24.15 sick building syndrome or IAQ cases, 24.13 who is legally responsible, 24.13 – 24.14 Legionnaires’ disease, 10.47 Liquid absorbents, 9.3 Lithium-bromide solution, properties of, 14.3 – 14.6 enthalpy-concentration diagram, 14.5 – 14.6 equilibrium chart, 14.4 mass balance in solution, 14.3 vapor pressure, 14.3 – 14.4 Load: block, 6.9 – 6.10 coil, 6.3 DX coil, 6.3 heating coil, 6.3 peak load, 6.9 – 6.10 profile 6.9 refrigeration, 6.3 space cooling, 6.3 Load calculation method: CLTD/SCL/CLF method, 6.15, 6.26 – 6.31 finite di ference, 6.34 – 6.39 TETD/TA method, 6.15 – 6.16 transfer function (TFM), 6.14 – 6.26 Load ratio, 5.13 Machinery room, refrigerating, 9.58 – 9.59 Maintenance, HVAC&R, 32.5 – 32.6 contractors and personnel, 32.5 – 32.6

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Maintenance, HVAC&R (Cont.) fault detection and diagnostics assisting predictive maintenance, 32.6 Maintenance to guarantee IAQ, 24.12 – 24.13 coils and ductwork, 24.12 – 24.13 inspection, service, and access, 24.12 monitoring of operation conditions, 24.12 Mass-transfer coefficients convective, 3.15 Masterformat, 1.23 Measurements, pressure and airfl w, 17.75 – 17.78 equal-area method, 17.77 – 17.78 log-linear rule for round duct, 17.77 – 17.78 log Tchebycheff rule, 17.7717.78 manometer, 17.75 – 17.77 measurements in air ducts, 17.76 – 17.77 Pitot tube, 17.75 – 17.77 Mechanical work, 4.4 Metabolic rate, 4.4 Microbial growth, eliminating, 24.4 – 24.6 basics, 24.4 eliminate water leaks, 24.5 microbial growth, 24.4 – 24.5 pressurization control, 24.5 prevent damped surface and material, 24.5 purge, 24.5 ultraviolet germicidal irradiation, 24.5 – 24.6 Moist air, 2.1 – 2.2 calculation of the properties of, 2.3 density, 2.10 enthalpy, 2.8 – 2.9 moist volume, 2.9 – 2.10 sensible heat, 2.10 – 2.11 Moisture content, 3.11 Moisture migration in building materials, 3.13 – 3.14 Moisture permeability index, 4.8 Moisture-solid relationship, 3.12 – 3.13 Moisture transfer, 3.11 – 3.17 from the surface, 3.14 – 3.15 in building envelopes, 3.16 – 3.17 Montreal Protocol and Clean Air Act, 9.10 – 9.11 Multistage vapor compression systems, 9.29 – 9.31 compound systems, 9.29 – 9.30 interstage pressure, 9.30 – 9.31 flash cooler and intercoole , 9.31

Night shutdown operating mode (Cont.) night shutdown period, 6.3 – 6.4 warm-up period, 6.4 – 6.6 Noise, 4.32 airfl w, 19.5 – 19.6 from chiller and pumps, 19.4 – 19.5 diffusers and grilles, 19.6 maximum duct velocities, 19.5 – 19.6 poor fan entry and discharge, 19.6 Noise control, recommended procedure, 19.3 – 19.4 Noise control for typical air system, 19.25 – 19.26 combination of supply fan noise and terminal noise, 10.25 environment adjustment factor, 19.26 estimated sound pressure level for space served by terminal units, 19.25 – 19.26 plenum ceiling effect, 19.26 Nomenclature, A.1 – A.6 Greek letter symbols, A.8 – A.9 subscripts, A.6 – A.8 Open data communication protocol, 5.41 application layer, 5.42 – 43 ARCNET, 5.44 BACnet, 5.41 – 5.44 data link/physical layer, 5.43 – 5.44 Ethernet, 5.43 – 5.44 local area networks (LANs), 5.43 LonTalk, 5.44 LonTalk LAN, 5.44 master-slave/token passing (MS/TP), 5.44 network layer, 5.43 network technology, 5.43 – 5.44 point-to-point, 5.44 proprietary network, 5.44 Outdoor air requirements for occupants, 4.30 – 4.31 Outdoor design conditions, 4.38 – 4.42 Outdoor design temperature, 4.38 – 4.42 1.0% summer wet-bulb, 4.39 summer dry-bulb, 4.39 summer mean coincident wet-bulb, 4.39 winter dry-bulb, 4.39 Overlooked commissioning, 1.17

Network technology, 5.43 – 5.44 Night shutdown operating mode, 6.3 – 6.6 conditioning period, 6.6 cool-down period, 6.4 – 6.6 influence of stored heat 6.6

Packaged systems, 29.2 -29.4 applications, 29.3 – 29.4 comparison between packaged and central systems, 29.2 – 29.3 types of, 29.4

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Packaged systems, fan-powered VAV, 29.18 – 29.22 case-study: rooftop packaged unit, 29.20 – 29.22 controls, 29.20, supply volume fl w rate and coil load, 29.19 – 29.20 system characteristics, 29.21 system description, 29.18 – 29.19 Packaged systems, perimeter-heating VAV, 29.18 system characteristics, 29.6 Packaged systems, single-zone constant-volume, 29.4 -29.6 controls, 29.5 energy use intensities, 29.5 supply volume fl w rate and coil loads, 29.4 – 29.5 system characteristics, 29.5 – 29.6 system description, 29.4 Packaged systems, single-zone VAV, 29.7 – 29.8 controls, 29.7 – 29.8 system calculations, 29.7 system characteristics, 29.6 system descriptions, 29.7 Packaged systems, VAV cooling, 29.9 – 29.12 duct static pressure control, 29.10 – 29.12 pressure characteristics, 29.10 supply volume fl w rate and coil load, 29.10 system characteristics, 29.6 system description, 29.9 – 29.10 Packaged systems, VAV reheat, 29.12 – 29.18 air-cooled, water-cooled, and evaporativecooled condensers, 29.17 air-side economizer mode, 29.15 case-study for precision manufacturing, 29.17 – 29.18 discharge air temperature control, 29.15 – 29.16 evenly distributed airfl w at DX coils, 29.14 – 29.15 fan modulation, 29.16 – 29.17 initiation of cooling stages, 29.15 – 29.16 night setback and morning warm-up, 29.14 reset, 29.16 sound control, 29.17 supply volume fl w rate and coil load, 29.12 – 29.14 system characteristics, 29.6 system description, 29.12 – 29.13 Packaged terminal air conditioner (PTAC), 1.4 Packaged terminal heat pump (PTHP), 1.4 Packaged units, 16.12 – 16.23 controls, 16.18 – 16.19

Packaged units (Cont.) indoor air quality, 16.18 indoor environmental control, 16.17 – 16.18 scroll compressors and evaporative condensers, 16.18 selection of, 16.19 – 16.22 Standard 90.1 – 1999 minimum efficien y requirements, 16.19 types of, 16.12 Packaged units, indoor, 16.15 – 16.16 Packaged units, rooftop, 16.12 – 16.15 compressors, 16.14 – 16.15 condensers, 16.15 curb, 16.13 DX-coils, 16.13 – 16.14 electric heating coil, 16.14 gas-fired furnace 16.14 heat pump, 16.15 humidifiers 16.14 supply, return, relief, and exhaust fans, 16.14 Packaged units, rooftop, sound control, 19.29 – 19 – 32 basics, 19.29 discharge side duct breakout, 19.31 sound source on return side, 19.31 – 19.32 sound sources and paths, 19.30 – 19.31 structure-borne noise, 19.32 Packaged units, split, 16.16 – 16.17 Panel heating and cooling, 28.33 Personal computer workstation, 5.39 – 5.40 Plant-building-loop, 7.43 – 7.51 balance valves, 7.49 – 7.50 building loop, 7.43 coil discharge air temperature control, 7.43 common pipe thermal contamination, 7.51 low T, 7.49 plant-loop, 7.43 pressure differential control, 7.45 sequence of operations, 7.46 – 7.49 staging control, 7.43 – 7.44 system characteristics, 7.45 – 7.46 variable-speed pumps connected in parallel, 7.49 water leaving chiller temperature control, 7.43 Plant-distributed pumping, 7.52 – 7.53 Plant-through-building loop, 7.40 – 7.42 bypass throttling fl w, 7.40 – 7.41 distributed pumping, 7.41 variable fl w, 7.41 – 7.42 Point or object, 5.25 Poor indoor air quality, 1.17 Precision, 2.6

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Pressure fl w characteristics, 22.22 – 22.24 fan characteristics, 22.7 mixing-exhaust section and conditioned space, 22.8 supply and relief fan combination, field sur vey system pressure characteristics, 22.23 – 22.24 supply and return fan combination system, 22.22 – 22.23 system pressure diagram, 22.5 – 22.8 VAV systems, fi ed part, 22.5 VAV systems, variable part, 22.5 variation of pressure in mixing box, 22.23 Pressure sensors, 5.19 reference pressure, 5.19 Primary ambient-air quality standard, 4.29 Profile angle 3.42 Properties of air, physical, A.15 Properties of moist air, thermodynamic, A.13 – A.14 Properties of water, physical, A.15 Psychrometric chart, A.12 Pump-piping systems, 7.34 – 7.38 connected in series, 7.35 – 7.36 modulation of, 7.36 – 7.37 operating point, 7.34 – 7.35 parallel-connected, 7.35 – 7.36 pump laws, 7.37 system curve, 7.34 Psychrometer, 2.12 – 2.13 aspiration, 2.14 – 2.15 sling, 2.14 – 2.15 Psychrometrics, 2.1 R-value, 3.7 overall, 3.7 Radiant heat loss from building, 3.46 – 3.47 Radiated noise, 19.18 – 19.19 break-out and break-in, 19.18 – 19.19 break-out and break-in sound power level, 19.19 – 19.20 duct rumble, 19.19 Radiation, atmospheric, 3.47 Reciprocating compression, performance, 11.29 – 11.34 condenser, 11.33 – 11.34 evaporator, 11.32 – 11.33 power input, 11.30 – 11.32 refrigeration capacity, 11.30 Reciprocating refrigeration systems, 11.2 – 11.42 air-cooled reciprocating chiller, 11.2 – 11.3 air-cooled reciprocating DX cooler, 11.2

Reciprocating refrigeration systems (Cont.) balance of capacities of selected components, 11.35 – 11.36 capacity control, 11.24 – 11.26 compressor components, 11.5 – 11.8 crankcase heater, 11.7 – 11.8 cylinder block and piston, 11.7 cylinder unloader, 11.24 filter dryer and straine , 11.10 – 11.11 frost control, 11.27 hot-gas bypass control, 11.26 liquid overfeed,11.3 – 11.4 liquid receiver, 11.8 liquid-suction heat exchanger, 11.8 – 11.10 low-pressure and high-pressure controls, 11.26 – 11.27 low-temperature control, 11.27 minimum performance, ASHRAE/IESNA Standard 90.1 – 1999, 11.41 – 11.42 motor overload control, 11.29 multistage, 11.4 oil lubrication, 11.7 oil-pressure failure control, 11.27 – 11.29 on/off control, 11.24 pressure relief valves, 11.11 – 11.12 real cycle of a single-stage, 11.4 – 11.5 reciprocating compressors, 11.5 refrigerant charge valve, 11.12 safety controls, 11.26 – 11.29 service valves, 11.11 – 11.12 solenoid valves, 11.11 speed modulation control, 11.24 – 11.26 suction and discharge valves, 11.7 system balance, 11.34 – 11.36 Reciprocating refrigeration systems, air-cooled direct-expansion, 11.36 – 11.42 compressor short cycling, 11.40 defrosting, 11.40 – 11.41 liquid slugging, 11.40 main problems, 11.40 – 11.42 oil returns, 11.40 operating balance, 11. 36 – 11.37 part-load operation using an unloader, 11.38 – 11.39 pressure characteristics, 11.37 – 11.38 proper refrigerant charge, 11.41 – 11.42 pump-down control, 11.39 – 11.40 Refrigerant fl w control devices, 10.51 – 10.58 advantages of electric expansion valves, 10.56 analog valves, 10.55 – 10.56 capacity superheat curve, 10.52 capillary tubes, 10.57 – 10.58 cross charge, 10.53 – 10.54 electric expansion valves, 10.55 – 10.56

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Refrigerant fl w control devices (Cont.) external equalizer, 10.52 – 10.53 float alves, high-side, 10.56 float alves, low-side, 10.56 – 10.57 hunting of thermostatic expansion valve, 10.10.54 – 10.55 limited liquid charge, 10.53 – 10.54 liquid charge, 10.53 – 10.54 operating characteristics, 10.51 – 10.52 pulse-width-modulated valve, 10.55 – 10.56 step motor valve, 10.55 straight charge, 10.53 – 10.54 thermostatic expansion valves, 10.51 – 10.53 Refrigerant piping for reciprocating refrigeration system, 11.12 – 11.23 copper tubing, 11.12 – 11.13 discharge line, 11.20 – 11.21 discharge line sizing, 11.20 – 11.21 double riser, 11.16 – 11.17 liquid line, 11.21 – 11.23 liquid line sizing, 11.22 – 11.23 maximum pressure drop, 11.17 minimum refrigeration load for oil entrainment up hot-gas riser, 11.20 minimum refrigeration load for oil entrainment up suction riser, 11.19 oil trap and piping pitch, 11.15 – 11.16 parallel connections, 11.23 piping design, 11.13 pressure drop of valves, and fitting 11.15 – 11.16 size of copper tubing, 11.14 sizing procedure, 11.14 – 11.15 suction line, 11.15 – 11.20 suction line sizing, 11.18 – 11.19 suction line sizing chart, 11.17 – 11.18 Refrigerants, 9.3 azeotropic, 9.3 blends, 9.3 CFCs replacements, 9.13 classification 9.13 – 9.16 concentration shift, 11.46 – 11.47 conversions and replacements, 9.11 glide, 9.3 – 9.4, 11.46 – 11.47 global warming potentials, 9.7 – 9.10 chlorofluorocarbons (CFCs) and halons 9.16 hydrochlorofluorocarbons (HCFCs) 9.15 – 9.16 hydrofluorocarbons (HFCs) 9.13 – 9.14 inorganic compounds, 9.16 near azeotropic, 9.3 numbering of, 9.4 ozone depletion potentials, 9.7 – 9.10 phase-out of CFC’s and halons , 9.10

Refrigerants (Cont.) recovery, recycle, and reclaiming, 9.11 – 9.13 reducing leakage and preventing deliberate venting, 9.11 – 9.13 restrict production of HCFCs, 9.10 – 9.11 storage of, 9.59 use of, 9.7 zeotropic, 9.3 Refrigerants, properties, 9.5 – 9.7 effectiveness of refrigeration cycle, 9.5 evaporating and condensing pressure, 9.6 inertness, 9.6 leakage detection, 9.6 – 9.7 oil miscibility, 9.6 physical properties, 9.6 refrigeration capacity, 9.6 safety requirements, 9.5 thermal conductivity, 9.6 Refrigerants safety, 9.56 Refrigerating machinery room, 9.58 – 59 storage of refrigerants, 9.59 Refrigeration, 9.2 unit of, 9.17 Refrigeration compressors, 9.51 – 9.56 direct-drive, belt drive, and gear drive, 9.53 energy use index, 9.55 – 9.56 hermetic, semihermetic, and open, 9.53 isentropic, and polytropic analysis, 9.54 – 9.55 motor, mechanical, and compression effi ciency, 9.54 performance, 9.53 – 9.56 positive displacement and nonpositive displacement, 9.51 – 9.53 volumetric efficien y, 9.53 – 9.54 Refrigeration cycles, 9.17 air expansion , 9.45 – 9.49 Carnot, 9.19 – 9.21 coefficient of performance 9.21 – 9.22 cycle performance, 9.22 – 9.24 determination of enthalpy by polynomials, 9.24 – 9.25 ideal vapor compression, single stage, 9.22 – 9.26 performance, 9.19 – 9.21 Refrigeration effect, refrigerating load, refrigerating capacity, 9.25 – 9.26 Refrigeration processes, 9.16 – 9.17 Refrigeration systems, 9.2 absorption, 9.2, 14.1 – 14.3 air or gas expansion, 9.2 cascade, 9.40 – 9.43 centrifugal, 13.1 – 13.7 classifications 9.49 – 9.51 compound, 9.31 – 9.40

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Refrigeration systems (Cont.) developments, recent ,9.51 high-probability systems, application rules, 9.56 – 9.57 low-probability systems, application rules, 9.57 – 9.58 multistage vapor compression, 9.29 – 9.31 reciprocating , 11.2 – 11.42 vapor compression, 9.2 Refrigeration systems, absorption, 14.1 – 14.3 applications, 14.3 cost analysis, 14.2 – 14.3 historical development, 14.2 types of, 14.1 – 14.2 Refrigeration systems, rotary, 11.42 – 11.43 main components, 11.43 rotary compressor, 11.42 – 11.43 system performance, 11.43 Refrigeration systems, screw, 11.55 air-cooled screw chillers, 11.55 ASHRAE/IESNA Standard 90.1 – 1999 minimum performance, 11.54 – 11.55 capacity control, 11.53 – 11.52 controls, 11.53 economizer, 11.54 electric expansion valves, 11.55 location of installation, 11.55 oil cooling, 11.51, 11.53 performance of twin-screw compressor, 11.52 – 11.53 screw compressors, 11.50 – 11.52 system performance, 11.55 types of, 11.50 variable volume ratio, 11.54 Refrigeration systems, scroll, 11.43 – 1150 capacity control and part-load performance, 11.47 – 11.48 chillers, 11.48 – 11.49 circulating concentration shift, 11.46 – 11.47 compressor performance, 11.46 concentration shift, 11.46 – 11.47 heat exchanger fl w configuration 11.47 radial and axial compliance, 11.44 – 11.45 scroll compressors, 11.44 – 11.45 system characteristics, 11.48 temperature glide, 11.46 – 11.47 types of, 11.43 – 11.44 Relative humidity, 2.7 – 2.8 Residuals, 5.61 normalized, 5.62 Resistance temperature detectors (RTDs), 2.6, 5.18 Retrofit remodeling, and replacement, 1.19

Return and exhaust inlets, 18.17 – 18.20 exhaust inlets, 18.19 light troffer diffuser, 18.19 – 18.20 return grilles, 18.18 – 18.19 return slots, 18.18 – 18.19 troffer diffuser slot, 18.18 – 18.19 Return and exhaust systems, 22.2 – 22.3 ANSI/ASHRAE Standard 90.1 – 1999 dampers specifications 22.3 enclosed parking garage ventilation, 22.3 exhaust hoods, 22.3 low-level return systems, 22.2 – 22.3 return ceiling plenum, 22.2 types of, 22.2 Room, 6.2 Room air conditioner, 1.4 Room heat pump, 1.4 Room sound power level and room sound pressure level, relationship, 19.23 – 19.24 array of ceiling diffusers, 19.24 single or multiple sound sources, 19.23 – 19.24 Safety factor, 1.20 Semiheated space, 3.49 Sensible heat exchange, 4.5 Sensing element, 5.16 Sensitivity, 2.6 Sensors, 2.6, 5.16 – 5.17 air, 5.16 – 5.18 air quality (VOC), 5.20 CO2, 5.20 drift, 5.16 intelligent network, 5.21 occupancy, 5.20 – 5.21 resistance temperature detectors (RTD), 5.18 temperature sensors, 5.18 wireless zone, 5.21 Sequence of operations, 5.5 – 5.6 Set point, 5.5 Shading coefficients 3.36 Shading devices, 3.40 – 3.43 draperies, 3.41 external, 3.42 – 3.43 indoor, 3.40 – 3.42 overhang, 3.42 roller shades, 3.41 – 3.42 side fin 3.42 venetian blinds, 3.40 – 3.41 Shading from adjacent buildings, 3.43 – 3.44 Sick building, 4.27 Sick building syndrome, 1.17, 4.27 Silencers, 19.12 – 19.17 characteristics, 19.14 – 19.15

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Silencers (Cont.) dissipative, 19.14 free area ratio, 19.15 insertion loss, 19.15 locations of, 19.15 – 19.16 packless, 19.14 pressure drop of, 19.15 reflection-dissipat ve, 19.14 selection of, 19.17 self noise of, 19.15 sound-attenuating plenum, 19.13 – 19.14 types of, 19.13 – 19.14 Silencers, active, 19.14 frequency limits, 19.16 operating characteristics, 19.16 performance, 19.17 system characteristics, 19.16 – 19.17 Simulation, energy software DOE-2.1E, 25.25 – 25.28 energy efficien y measures, 25.27 energy simulation software, 25.25 – 25.26 loads, 25.25 plant, 25.27 – 25.28 systems, 25.26 – 25.27 Simulation, system, 25.17 – 25.19 dynamic simulation, 25.18 energy simulation, 25.17 performance equations, 25.17 – 25.18 physical modeling, 25.18 sequential, 25.19 simultaneous, 25.19 steady-state, 25.18 – 25.19 Simulation of a centrifugal chiller, 25.19 – 25.25 centrifugal compressor model, 25.23 – 25.25 condenser model, 25.22 cooling tower model, 25.23 evaporator model, 25.20 – 25.21 operating parameter, 25.20 simulation methodology, 25.20 system model, 25.19 – 25.20 Skin wetness, 4.9 Smoke control and fire safet , 22.24 – 22.38 ANSI/NFPA 92A and 92B, 22.28 automatic sprinkler on fire protection 22.27 – 22.28 effective area and fl w rates, 22.27 fire safety in uildings 22.24 – 22.25 smoke control in atria, 22.28 smoke management in atria, malls, and large areas, 22.28 smoke movement in buildings, 22.25 – 22.27 zone smoke control, 22.31 – 22.32 zone smoke control, design considerations, 22.32

Software, load calculations and energy analysis, 6.42 – 6.49 building load analysis and system thermodynamics (BLAST), 6.42 TRACE-600, 6.42 – 6.49 Sol-air temperature, 3.47 Solar angles, 3.22 – 3.25 altitude angle, 3.23 – 3.24 angle of incidence, 3.23 – 3.25 hour angle, 3.22 – 3.24 latitude angle, 3.22 – 3.24 relationships, 3.23 – 3.24 solar azimuth angle, 3.23 – 3.24 solar declination angle, 3.22 – 3.24 surface-solar azimuth angle, 3.23 – 3.24 Solar constant, 3.25 Solar heat gain coefficient (SHGC) 3.33 Solar heat gain factors, 3.37 Solar intensity, 3.24 – 3.25 direct normal radiation, 3.26 Solar radiation, 3.25 – 3.29 apparent, 3.26 diffuse radiation, 3.26 direct radiation, 3.26 extraterrestrial intensity of, 3.25 for a clear sky, 3.26 – 3.283.28 – 3.29 reflection of 3.28 Sorption isotherm, 3.11 – 3.12 Sound, 4.32 airborne, 4.32 octave bands, 4.33 power, 4.32 power level, 4.32 – 4.33 pressure level, 4.32 – 4.33 Sound attenuation, along duct-borne path, 19.6 – 19.12 duct-borne crosstalk,19.11 in ducts, 19.6 – 19.9 at elbows and branch takeoffs, 19.9 – 19.10 end reflection loss 19.10 – 19.11 inner-lined round ducts, 19.7 lined fl xible ducts, 19.8 – 19.9 lined rectangular ducts, 19.8 unlined rectangular sheet-metal ducts, 19.7 unlined round ducts, 19.7 Sound control, 19.1 – 19.2 control at design stage, 19.3 Sound control criteria, 4.34 A-weighted sound level, 4.34 noise criteria (NC), 4.34 room criteria (RC), 4.34 Sound paths, 19.2 – 19.3 airborne, 19.2 duct-borne, 19.2

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pg I.23

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Smoke paths (Cont.) radiated sound, 19.2 structure-borne, 19.2 – 19.3 Space, 6.2 Space air diffusion, mixing fl w, design procedure, 18.31 – 18.34 choose an optimum throw/characteristic length ratio, 18.33 design characteristics of slot diffusers in perimeter zone, 18.33 – 18.34 drop of cold air jet, 18.34 final layout 18.34 select the type of supply outlet, 18.31 – 18.32 sound level, 18.34 total pressure loss of supply outlet, 18.34 volume fl w rate per outlet or unit length, 18.32 – 18.33 Space air diffusion, principles, 18.2 – 18.5 age of air, 18.4 – 18.5 air change effectiveness, 18.4 air diffusion performance index (ADPI), 18.3 – 18.4 draft, 18.2 draft temperature, effective, 18.2 – 18.3 nominal air change effectiveness, 18.5 nominal time constant, 18.5 space air velocity vs. space air temperature, 18.3 space diffusion effectiveness factor, 18.4 turbulence intensity, 18.2 – 18.3 ventilation effectiveness, 18.4 Space airfl w pattern, displacement fl w, 18.38 – 18.43 ceiling plenum, 18.41 supply air velocity, 18.41 unidirectional fl w, 18.38 – 19.39 unidirectional fl w for clean rooms, 18.39 – 18.40 ventilating ceiling, 18.40 – 18.41 Space airfl w pattern, mixing fl w, 18.20 – 18.28 airfl w pattern, 18.20 principles and characteristics, 18.21 reverse air streams in the occupied zone, 18.21 stratified mixing f w, 18.25 – 18.28 stratified mixing f w using nozzles, 18.27 – 18.28 types and locations of return and exhaust inlets, 18.21 types and locations of supply outlets, 18.21 using ceiling diffusers, 18.23 – 18.24 using high-side outlets, 18.21 – 18.23 using sill or floor outlets 18.24 – 18.25 using slot diffusers, 18.24

Space airfl w pattern, projecting fl w, 18.44 – 18.48 applications of desktop task conditioning systems, 18.48 benefits of 18.44 desktop task conditioning systems, 18.46 – 18.48 distance between target zone and supply outlet, 18.44 horizontal vs. vertical jet, 18.44 – 18.46 industrial spot cooling systems, 18.44 – 18.46 performance of desktop task conditioning systems, 18.47 – 18.48 recommendations in spot cooling design, 18.46 target velocities, 18.46 thermal sensation, 18.46 Space airfl w pattern, stratified displacemen fl w, 18.42 – 18.43 comparison of stratified displacement f w and mixing fl w, 18.43 operating characteristics, 18.42 – 18.43 two-zone stratified model 18.42 Space airfl w pattern, upward fl w underfloo air distribution, 18.48 – 51 applications, 18.51 consistent access plenum temperature, 18.50 design considerations, 18.50 – 18.51 floor plenum master zone air temperatur control, 18.50 heat unneutralized, 18.50 thermal storage of floor plenum 18.49 upward fl w from floor plenum 18.48 – 18.49 Space heat extraction rate, 6.3 Space heat gain, 6.3 Space pressurization and return/relief volume fl w controls, 23.16 – 23.18 characteristics of space pressure control, 23.16 – 23.17 VAV systems return/relief fan volume fl w control, 23.17 – 23.18 volume fl w of air leakage and effective leakage area, 23.17 Space pressurization control, 24.13 Space pressurization or building pressurization, 4.37 – 4.38, 20.7 – 20.14 airfl w balance, 20.11 – 20.13 air systems and mechanical ventilation systems, 20.11 characteristics, 20.7 by differential fl w, 20.11 – 20.13 differentials, 4.37 – 4.38 neutral pressure level, 20.7 – 20.9

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Space pressurization or building pressurization (Cont.) stack effect, 20.7 – 20.8 stack effect for high-rise buildings, 20.9 – 20.10 wind effect, 20.10 – 20.11 Specifications 1.22 – 1.23 Stairwell pressurization, 22.29 – 22.34 bottom single injection or bottom and top injection, 22.34 – 22.35 characteristics, 22.29 – 22.30 overpressure relief and feedback control, 22.30 – 22.31 pressure drop coefficient 22.34 stair and shaft vents, 22.31 system pressure loss, 22.33 – 22.35 volume fl w rate, 22.32 – 22.33 Standard 90.1 – 1999 for building envelope, 3.48 – 50 Standard 90.1 – 1999, simplified approach op tion for small and medium HVAC&R systems, 29.8 – 29.9 Steam point, 2.4 – 2.5 Subcooling, 9.26 Superheating, 9.26 – 9.27 Supply air condition, determination, 20.62 – 20.66 air conditioning rules, 20.63 graphical method, 20.63 – 20.64 influence of sensible heat ratio 20.64 – 20.66 Supply outlets, 18.11 – 18.17 ceiling diffusers, 18.12 – 18.14 gang-operated turning vanes, 18.17 grilles, 18.11 – 18.12 induction, 18.14 nozzle diffusers, 18.16 – 18.17 nozzles, 18.16 – 18.17 plenum box, 18.14 – 18.15 registers, 18.11 – 18.12 slot diffusers, 18.14 – 18.16 split dampers, 18.17 Supply volume fl w rate, 20.59 – 20.62 based on space cooling vs. heating load, 20.59 – 60 rated volume fl w of supply and return fans, 20.61 – 20.62 requirements other than cooling load, 20.60 – 20.62 temperature difference vs. enthalpy difference, 20.60 volumetric vs. mass fl w rate, 20.60 System pressure diagram, 22.5 – 22.8 duct static pressure control, 22.5 – 22.7

Supply volume fl w rate (Cont.) fan characteristics, 22.7 – 22.8 Temperature, 2.4 dew point, 2.11 globe, 4.9 mean radiant, 4.9 – 4.12 mean surface temperature of clothing, 4.5 measurements, 2.6 operative, 4.5 Temperature scales, 2.4 – 2.5 absolute scale, 2.5 Celsius, 2.4 – 2.5 Fahrenheit, 2.4 – 2.5 Kelvin, 2.4 – 2.5 Rankine, 2.4 – 2.5 thermodynamic, 2.5 Testing, adjusting, and balancing (TAB), 32.2 – 32.4 Thermal comfort, 4.15 – 4.20 ASHRAE comfort zones, 4..17 – 4.18 comfort-discomfort diagrams, 4.17 – 4.20 factors affecting, 4.14 – 4.15 Fanger’s comfort chart, 4.15 – 4.17 Fanger’s comfort equation, 4.15 – 4.17 heart rate (HR), 4.19 – 4.20 predicted mean vote (PMV), 4.15 – 4.17 thermal sensational scale, 4.16 Thermal insulation, 3.18 – 3.22 economic thickness, 3.21 Thermal interaction: between human body and indoor environment, 4.2 steady-state thermal equilibrium, 4.3 transient energy balance, 4.3 two-node model, 4.2 Thermal resistance, 3.4 of airspaces, 3.21 – 3.22 convective, 3.5 Thermal resistance ratio, 3.19 – 3.21 Thermal storage systems, 31.1 – 31.5 benefits and dr wbacks, 31.2 – 31.3 full storage or load shift, 31.3 – 31.5 ice-storage and chilled water storage, 31.5 impact of electric deregulation, 31.2 partial storage or load leveling, 31.3 – 31.5 system description, 31.1 – 31.2 Thermistors, 2.6 Thermodynamic wet bulb temperature, 2.12 Thermometer, globe, 4.9 Total shortwave irradiance, 3.34, 3.37 TRACE 600 input, 6.42 – 6.49 external loads, 6.45

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TRACE 600 input (Cont.) internal loads, 6.46 – 6.47 job, 6.44 – 6.45 load methodology, 6.43 – 6.44 schedules, 6.45 – 6.46 structure and basics, 6.42 – 6.43 TRACE 600, minimum input, run, and outputs, 6.47 Transducers, 5.21 Transfer function, method, 6.14 – 6.26 ceiling, floors and interior partition walls, 6.16 – 6.17 conversion of heat gain to cooling load, 6.24 – 6.25 electric motors, 6.21 – 6.23 equipment and appliances, 6.21 – 6.23 exterior wall and roofs, 6.16 heat extraction rate, 6.25 heat loss to surroundings, 6.25 – 6.26 heat to space, 6.20 infiltration 6.24 lighting, 6.18 – 6.21 space air temperature, 6.25 window glass, 6.17 Transmission losses, 19.19 – 19.23 for selecting building structures, 19.23 TLin for flat val ducts, 19.22 TLin for rectangular ducts, 19.22 TLin for round ducts, 19.22 TLout for flat val ducts, 19.21 TLout for rectangular ducts, 19.21 TLout for round ducts, 19.20 – 19.21 Transmitters, 5.21 Triple point, 2.4 – 2.5 T-w chart, 2.19 Unit conversion, Inch-Pound (I-P) units to SI units, A.15 – A.17 Updated technology, 1.17 Valves, 7.16 – 7.17 balancing, 7.17 check, 7.16 – 7.17 connections and ratings, 7.17 – 7.18 gate, 7.16 globe, 7.16 – 7.17 materials, 7.18 pressure relief, 7.17 Vapor retarders, 3.17, 3.18 Variable-air-volume (VAV) systems, 1.11, 21.2 – 21.56 comparison between various VAV systems, 21.56

Variable-air-volume (VAV) systems (Cont.) dew point control, 23.27 – 23.28 diagnostics, 23.28 functional controls, 23.26 – 23.28 interaction between controls, 23.29 – 23.30 nighttime setback and warmup or cooldown control, 23.26 – 23.27 override, 23.29 – 23.30 recommendations for VAV controls, 23.28 – 23.29 sequence control, 23.29 specific controls 23.2 steam humidifier control 23.27 types of, 21.2 – 21.3 VAV systems, dual duct, 21.33 – 21.44 case-study, 21.42 – 21.44 discharge air temperature control, 21.40 – 21.41 mixing mode operation, 21.38 mixing VAV box, 21.36 – 21.38 number of supply fans, 21.36 part-load operation, 21.43 – 21.44 system description, 21.33 – 36 winter heating and winter cooling mode operation, 21.43 zone control and sequence of operations, 21.38 – 21.40 zone supply fl w rate, 21.41 – 21.42 VAV systems, fan-powered, 21.44 – 21.56 design considerations, 21.55 – 21.56 fan energy use, 21.54 – 21.55 fan-powered VAV box, 21.48 – 21.50 parallel fan-powered VAV box, 21.48 – 21.50 parallel fan-powered VAV box, fan characteristics, 21.50 – 21.51 series fan-powered VAV box, 21.48 – 21.49 supply volume fl w rate,21.53 – 21.54 system description, 21.44 – 21.47 zone control and sequence of operations, 21.52 – 21.53 VAV systems, single-zone, 21.2 – 21.18 air conditioning cycle and system calculations, 21.4, 21.16 – 21.17 system description, 21.3 – 21.5 year-round operation of, 21.5 – 21.8 zone temperature control - sequence of operations, 21.17 – 21.18 VAV systems, VAV cooling, VAV reheat, and perimeter heating VAV systems, 21.18 – 21.33 air skin VAV system, 21.21 ANSI/ASHRAE Standard 90.1 – 1999 specifi cations, 21.20

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VAV systems, VAV cooling, VAV reheat, and perimeter heating VAV systems (Cont.) minimum ventilation, discharge air temperature, and duct static pressure controls , 23.30 – 23.34 perimeter heating VAV systems, 21.20 – 21.21 reheating VAV box, 21.23 sequence of operations, primary considerations, 23.30 – 23.35 stability of zone control, 21.26 – 21.27 VAV box, 21.21 – 21.23 VAV box, pressure dependent and pressure independent, 21.23 VAV box, sound level, 21.23 – 21.25 VAV cooling systems, 21.18 – 21.19 VAV reheat system, case-study, 21.27 – 21.33 VAV reheat system, cooling mode part-load operation, 21.29 – 21.30 VAV reheat system, dead-band mode, 21.25 – 21.26 VAV reheat system, winter cooling mode in interior zone, 21.32 – 21.33 VAV reheat system, winter reheating in perimeter zone, 21.30 – 21.32 VAV reheat system, volume fl w rate and coil load, 21.28 – 21.29 VAV reheat systems, 21.19 VAV reheat zone temperature control sequence of operations, 21.25 – 21.26 Ventilation, 24.2 air economizer, 24.2 – 24.3 minimum outdoor air damper and economizer damper, 24.3 minimum ventilation control, 24.3 outdoor air requirement, 24.2 purge operation, 24.2 – 24.3 time of operation, 24.2 Ventilation control, minimum, 23.2 – 23.5 ASHRAE Standard 62 – 1999, 23.3 -23.4 basic approach, 23.2 conference rooms, 23.16 direct measurement of minimum outdoor air intake, 23.15 fan tracking systems, 23.15 – 23.16 high-occupancy areas, 23.5 indoor air quality procedure, 23.3 – 23.4 outdoor air injection fan, 23.14 – 23.15 recirculation of unused outdoor air, 23.4 – 23.5 types of, 23.2 – 23.3 ventilation rate procedure, 23.3 Ventilation control, minimum, mixed-plenum pressure, 23.12 – 23.14 applications, 23.14

Ventilation control, minimum, mixed-plenum pressure (Cont.) monitoring plenum pressure, 23.12 – 23.13 monitoring pressure drop of louver and damper, 23.13 – 23.14 supply and return fans, 23.13 Volume fl w control, 5.33 – 5.35 branch fl w control, 5.33 – 5.34 bypass control, 5.35 – 5.34 mixed-air control, 5.33 – 5.34 Warm air furnace, 8.3 – 8.9 annual fuel utilization efficien y (AFUE), 8.7 circulating fan, 8.3 – 8.4 condensing or noncondensing, 8.7 control and operation, 8.8 – 8.9 gas burners, 8.3 gas-fired 8.3 heat exchangers, 8.3 ignition, 8.3 minimum efficien y, 8.8 power vent or natural vent, 8.7 steady state efficien y (SSE), 8.7 thermal efficien y, 8.6 – 8.7 types of, 8.3 venting arrangements, 8.4 Warm air heating system, low-pressure ducted, 8.17 – 8.23 duct efficien y, 8.20 duct leakage, 8.20 – 8.21 location of furnace, 8.20 part-load operation and control, 8.21 – 8.22 supply and return duct, 8.18 supply duct and return plenum, 8.18 system efficien y, 8.20 thermal stratification 8.21 Water: chilled, 1.8 column (WC), 4.38 condenser, 1.8 valves, 5.26 vapor, 2.1 – 2.2 Water heat gain factor, 10.21 Water impurities, 7.25 – 7.26 Water piping, 7.7 – 7.16 dimensions, copper, 7.10 – 7.11 dimensions, steel, 7.8 – 7.9 expansion and contraction, 7.14 – 7.15 fittings 7.18 – 7.19 insulation, 7.7.15 – 7.16 material, 7.7 supports, 7.14 – 7.15

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Water piping (Cont.) system accessories, 7.19 Water-source heat pumps, 1.5, 28.26 – 28.27 control, 28.30 – 28.3 energy performance by ASHRAE Standard 90.1 – 1999, 28.27 Water-source heat pump systems, 28.24 – 28.33 air system and maintenance, 28.29 case-study, 28.31 – 28.32 close-circuit evaporative water cooler, 28.27 – 28.29 controls, 28.30 – 28.31 design considerations, 28.32 – 28.33 loop temperatures, 28.25 – 28.26 operating characteristics, 28.24 – 28.25 safety controls, 28.30 storage tank, 28.29 system characteristics, 28.15 system description, 28.24 – 28.35 water heater, 28.29 water-loop temperature control, 28.30 Water systems, 1.8, 7.2 accessories, 7.18 – 7.19 air in, 7.23 – 7.24 campus type, 7.53 – 7.58 chilled, 7.2 chiller plant, 7.39 closed, 7.2 condenser or cooling, 7.2 dual-temperature, 7.2 evaporative-cooled, 7.2 friction chart, copper pipes, 7.6 friction chart, plastic pipes, 7.7 friction chart, steel pipes 7.6 hot, 7.2 maximum allowable pressures, 7.12 – 7.13 once through, 7.4 open, 7.2

Water systems (Cont.) oxidation, 7.24 – 25 pressure drop, 7.5 – 7.7 pressurization control, 7.19 – 7.20 pump location, 7.23 temperature difference, 7.4 – 7.5 types of, 7.40 variable fl w, 7.40 volume fl w, 7.4 – 7.5 volume fl w, chilled water, 7.38 – 7.39 water velocity, 7.5 waterlogging, 7.24 – 25 wire-to-water efficien y, 7.37 – 7.38 Water treatments, 7.27 – 7.28 chemical feeding, 7.27 microbiological control, 7.26 scale and corrosion control, 7.26 Wet bulb: constant, 2.13 depression, 2.13 temperature, 2.12 – 2.14 Window glass: clear plate, 3.29 double-strength sheet glass, 3.36 glass temperature, 3.35 heat gain for double-glazing, 3.34 – 3.36 heat gain for single-glazing, 3.32 – 3.34 insulating, 3.29 low-emissivity (low-E), 3.29 – 3.30 optical properties, 3.30 – 3.31 reflect ve coated, 3.29 spectral transmittance, 3.31 tinted heat-absorbing, 3.29 type of, 3.29 – 3.30 U-values, 3.33

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