HVAC Systems Testing Adjusting & Balancing 3rd Edition

HVAC SYSTEMS TESTING, ADJUSTING & BALANCING SHEET METAL AND AIR CONDITIONING CONTRACTORS’ NATIONAL ASSOCIATION, INC. ...

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HVAC SYSTEMS TESTING, ADJUSTING & BALANCING

SHEET METAL AND AIR CONDITIONING CONTRACTORS’ NATIONAL ASSOCIATION, INC.

HVAC SYSTEMS TESTING, ADJUSTING & BALANCING

THIRD EDITION — AUGUST, 2002

SHEET METAL AND AIR CONDITIONING CONTRACTORS’ NATIONAL ASSOCIATION, INC. 4201 Lafayette Center Drive Chantilly, VA 20151-1209

HVAC SYSTEMS TESTING, ADJUSTING & BALANCING COPYRIGHT2002 All Rights Reserved by

SHEET METAL AND AIR CONDITIONING CONTRACTORS’ NATIONAL ASSOCIATION, INC. 4201 Lafayette Center Drive Chantilly, VA 20151 Printed in the U.S.A.

FIRST EDITION - 1983 SECOND EDITION - JULY, 1993 THIRD EDITION - AUGUST, 2002

Except as allowed in the Notice to Users and in certain licensing contracts, no part of this book may be reproduced, stored in a retrievable system, or transmitted, in any form or by any means, electronic, mechanical, photocopying, recording, or otherwise, without the prior written permission of the publisher.

FOREWORD This handbook has been extensively updated for 2002 from the original 1983 publication and includes all of the many changes that have takes place in the industry since the 1990’s. We have added many new sections covering variable frequency drives (VFD), direct digital control (DDC) systems, lab hood exhaust balancing, and the latest changes in the balancing equipment and procedures. All of the system testing, adjusting, and balancing fundamentals that make up the original text has been updated, and all helpful reference tables and charts in the Appendix have been extensively updated. This handbook will provide any SMACNA contractor already familiar with mechanical system operation basics, with the information necessary to balance most heating, ventilation, and air conditioning (HVAC) systems. Chapters on both air and water side HVAC system adjusting and balancing are included, and the chapters on system controls have been totally rewritten to reflect the trend away from pneumatic controls and towards programmable micro−processor controls. Most of today’s HVAC systems are being designed with many more individually controlled temperature zones to im− prove occupant comfort, and variable speed fans and pumps are now commonplace to provide the exact amount of heating and cooling system capacity necessary to minimize energy usage. New occupant air ventilation codes are much more restrictive, at the same time building envelopes are becoming much tighter. The combination of constantly changing HVAC flows and increased demand for fresh and filtered ventilation air for all occupants is placing much more emphasis on proper HVAC system operation and balancing. Any SMACNA contractor wanting to become part of this rapidly growing field is strongly encouraged to read other related SMACNA publications available, and take part in the many training courses offered to become a certified TAB Contractor. The International Training Institute provides a Certified Technician program for journeyman sheet metal workers who already have a basic understanding of system testing and balancing, and many of these courses are avail− able in versions for home study. The building construction industry is experiencing a major growth in demand for trained and experienced contractors who can balance today’s much more complex HVAC systems.

SHEET METAL AND AIR CONDITIONING CONTRACTORS’ NATIONAL ASSOCIATION, INC.

HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

iii

TASK FORCE Bill Freese, Chairman International Testing & Balancing, Ltd. Seaford, New York

Ray Coleman Certified Testing & Balancing, Inc. Riverton, Utah

David Aldag Aldag−Honold Mechanical, Inc. Sheboygan, Wisconsin

Ben Dutton SMACNA, Inc. Chantilly, Virginia

John Brue Balancing Precision, Inc. Bloomington, Illinois

Eli P. Howard, III SMACNA, Inc. Chantilly, Virginia

OTHER CONTRIBUTORS J. R. Yago & Associates Consulting Engineers Manakin−Sabot, Virginia

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HVAC SYSTEMS Testing Adjusting & Balancing • Third Edition

NOTICE TO USERS OF THIS PUBLICATION

1.

DISCLAIMER OF WARRANTIES

a) The Sheet Metal and Air Conditioning Contractors’ National Association (“SMACNA”) provides its product for informational purposes. b) The product contains “Data” which is believed by SMACNA to be accurate and correct but the data, including all information, ideas and expressions therein, is provided strictly “AS IS”, with all faults. SMACNA makes no warranty either express or implied regarding the Data and SMACNA EXPRESSLY DISCLAIMS ANY IMPLIED WARRANTIES OF MERCHANTABILITY OR FITNESS FOR PARTICULAR PURPOSE. c) By using the data contained in the product user accepts the Data “AS IS” and assumes all risk of loss, harm or injury that may result from its use. User acknowledges that the Data is complex, subject to faults and requires verification by competent professionals, and that modification of parts of the Data by user may impact the results or other parts of the Data. d) IN NO EVENT SHALL SMACNA BE LIABLE TO USER, OR ANY OTHER PERSON, FOR ANY INDIRECT, SPECIAL OR CONSEQUENTIAL DAMAGES ARISING, DIRECTLY OR INDIRECTLY, OUT OF OR RELATED TO USER’S USE OF SMACNA’S PRODUCT OR MODIFICATION OF DATA THEREIN. This limitation of liability applies even if SMACNA has been advised of the possibility of such damages. IN NO EVENT SHALL SMACNA’S LIABILITY EXCEED THE AMOUNT PAID BY USER FOR ACCESS TO SMACNA’S PRODUCT OR $1,000.00, WHICHEVER IS GREATER, REGARDLESS OF LEGAL THEORY. e) User by its use of SMACNA’s product acknowledges and accepts the foregoing limitation of liability and disclaimer of warranty and agrees to indemnify and hold harmless SMACNA from and against all injuries, claims, loss or damage arising, directly or indirectly, out of user’s access to or use of SMACNA’s product or the Data contained therein.

2.

ACCEPTANCE

This document or publication is prepared for voluntary acceptance and use within the limitations of application defined herein, and otherwise as those adopting it or applying it deem appropriate. It is not a safety standard. Its application for a specific project is contingent on a designer or other authority defining a specific use. SMACNA has no power or authority to police or enforce compliance with the contents of this document or publication and it has no role in any representations by other parties that specific components are, in fact, in compliance with it.

3.

AMENDMENTS

The Association may, from time to time, issue formal interpretations or interim amendments, which can be of significance between successive editions.

4.

PROPRIETARY PRODUCTS

SMACNA encourages technological development in the interest of improving the industry for the public benefit. SMACNA does not, however, endorse individual manufacturers or products.

5.

FORMAL INTERPRETATION

a) A formal interpretation of the literal text herein or the intent of the technical committee or task force associated with the document or publication is obtainable only on the basis of written petition, addressed to the Technical Resources Department and sent to the Association’s national office in Chantilly, Virginia. In the event that the petitioner has a substantive disagreement with the interpretation, an appeal may be filed with the Technical Resources Committee, which has technical oversight responsibility. The request must pertain to a specifically identified portion of the document that does not involve published text which provides the requested information. In considering such requests, the Association will not review or judge products or components as being in compliance with the document or publication. Oral and written interpretations otherwise obtained from anyone affiliated with the Association are unofficial. This procedure does not prevent any committee or task force chairman, member of the committee or task force, or staff liaison from expressing an opinion on a provision within the document, provided that such person clearly states that the opinion is personal and does not represent an official act of the Association in any way, and it should not be relied on as such. The Board of Directors of SMACNA shall have final authority for interpretation of this standard with such rules or procedures as they may adopt for processing same. b) SMACNA disclaims any liability for any personal injury, property damage, or other damage of any nature whatsoever, whether special, indirect, consequential or compensatory, direct or indirectly resulting from the publication, use of, or reliance upon this document. SMACNA makes no guaranty or warranty as to the accuracy or completeness of any information published herein.

6.

APPLICATION

a) Any standards contained in this publication were developed using reliable engineering principles and research plus consultation with, and information obtained from, manufacturers, users, testing laboratories, and others having specialized experience. They are


v

subject to revision as further experience and investigation may show is necessary or desirable. Construction and products which comply with these Standards will not necessarily be acceptable if, when examined and tested, they are found to have other features which impair the result contemplated by these requirements. The Sheet Metal and Air Conditioning Contractors’ National Association and other contributors assume no responsibility and accept no liability for the application of the principles or techniques contained in this publication. Authorities considering adoption of any standards contained herein should review all federal, state, local, and contract regulations applicable to specific installations. b) In issuing and making this document available, SMACNA is not undertaking to render professional or other services for or on behalf of any person or entity. SMACNA is not undertaking to perform any duty owed to any person or entity to someone else. Any person or organization using this document should rely on his, her or its own judgement or, as appropriate, seek the advice of a competent professional in determining the exercise of reasonable care in any given circumstance.

7.

REPRINT PERMISSION

Non-exclusive, royalty-free permission is granted to government and private sector specifying authorities to reproduce only any construction details found herein in their specifications and contract drawings prepared for receipt of bids on new construction and renovation work within the United States and its territories, provided that the material copied is unaltered in substance and that the reproducer assumes all liability for the specific application, including errors in reproduction.

8.

THE SMACNA LOGO

The SMACNA logo is registered as a membership identification mark. The Association prescribes acceptable use of the logo and expressly forbids the use of it to represent anything other than possession of membership. Possession of membership and use of the logo in no way constitutes or reflects SMACNA approval of any product, method, or component. Furthermore, compliance of any such item with standards published or recognized by SMACNA is not indicated by presence of the logo.

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TABLE OF CONTENTS

TABLE OF CONTENTS FOREWORD . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

iii

TASK FORCE . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

iv

NOTICE TO USERS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

v

TABLE OF CONTENTS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

vii

CHAPTER 1 1.1 1.2 1.3

INTRODUCTION INTRODUCTION TO TAB WORK . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . THE TAB TECHNICIAN/TEAM . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . GENERAL REQUIREMENTS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

1.1 1.1 1.2

CHAPTER 2 HVAC FUNDAMENTALS 2.1 2.2 2.3 CHAPTER 3 3.1 3.2 3.3 3.4 3.5 3.6 CHAPTER 4 4.1 4.2 4.3 4.4 4.5 4.6 4.7 CHAPTER 5 5.1 5.2 5.3 5.4 5.5 CHAPTER 6 6.1 6.2 6.3 6.4 6.5

HEAT FLOW . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . PSYCHROMETRICS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . FLUID MECHANICS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

2.1 2.6 2.19

ELECTRICAL EQUIPMENT AND CONTROLS ELECTRICAL SYSTEMS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . ELECTRICAL SERVICES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . TRANSFORMERS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . MOTORS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . MOTOR CONTROLS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . VARIABLE FREQUENCY DRIVES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

3.1 3.1 3.5 3.5 3.8 3.9

TEMPERATURE CONTROL AUTOMATIC TEMPERATURE CONTROL SYSTEMS . . . . . . . . . . . . . . . . . . . . CONTROL LOOPS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . CONTROL DIAGRAMS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . CONTROL RELATIONSHIPS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . ATC SYSTEM ADJUSTMENT . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . TAB/ATC RELATIONSHIP . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . CENTRALIZED CONTROL SYSTEMS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

4.1 4.2 4.5 4.5 4.6 4.6 4.7

FANS FAN CHARACTERISTICS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . FAN CONSTRUCTION . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . FAN AIRFLOW AND PRESSURES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . FAN/SYSTEM CURVE RELATIONSHIP . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . FAN CAPACITY RATINGS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

5.1 5.4 5.10 5.13 5.17

AIR DISTRIBUTION AND DEVICES AIR TERMINAL BOXES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . VARIABLE AIR VOLUME (VAV) TERMINAL BOXES . . . . . . . . . . . . . . . . . . . . . OTHER AIRFLOW DEVICES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . AIR DISTRIBUTION BASICS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . ROOM AIR DISTRIBUTION . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .


6.1 6.3 6.3 6.6 6.9

vii

CHAPTER 7 7.1 7.2 7.3 7.4 7.5 CHAPTER 8 8.1 8.2 8.3 8.4 8.5 CHAPTER 9 9.1 9.2 9.3 9.4

AIR SYSTEMS INTRODUCTION . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . TYPES OF AIR SYSTEMS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . AIR SYSTEM DESIGN . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . DUCT SIZING EXAMPLES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . SUMMARY . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

7.1 7.2 7.9 7.11 7.14

HYDRONIC EQUIPMENT PUMPS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . PUMP / SYSTEM CURVE RELATIONSHIP . . . . . . . . . . . . . . . . . . . . . . . . . . . . . PUMP INSTALLATION CRITERIA . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . HYDRONIC HEATING AND COOLING SOURCES . . . . . . . . . . . . . . . . . . . . . . TERMINAL HEATING AND COOLING UNITS . . . . . . . . . . . . . . . . . . . . . . . . . . .

8.1 8.7 8.11 8.13 8.14

HYDRONIC SYSTEMS HYDRONIC SYSTEMS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . HYDRONIC SYSTEM DESIGN . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . HYDRONIC DESIGN PROCEDURES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . STEAM SYSTEMS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

9.1 9.8 9.13 9.14

CHAPTER 10 REFRIGERATION SYSTEMS 10.1 10.2 10.3 10.4 10.5 10.6 10.7 CHAPTER 11 11.1 11.2 11.3 11.4 11.5 11.6 11.7 11.8

REFRIGERATION SYSTEMS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . REFRIGERATION TERMS AND COMPONENTS . . . . . . . . . . . . . . . . . . . . . . . . SAFETY CONTROLS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . OPERATING CONTROLS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . REFRIGERANTS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . THERMAL BULBS AND SUPERHEAT . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . COMPRESSOR SHORT CYCLING . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

10.1 10.2 10.4 10.4 10.4 10.4 10.6

TAB INSTRUMENTS INTRODUCTION . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . AIRFLOW MEASURING INSTRUMENTS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . PRESSURE GAGE, CALIBRATED . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . ROTATION MEASURING INSTRUMENTS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . TEMPERATURE FUNCTION TACHOMETER MEASURING INSTRUMENTS ELECTRICAL MEASURING INSTRUMENTS . . . . . . . . . . . . . . . . . . . . . . . . . . . . COMMUNICATION DEVICES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . HYDRONIC FLOW MEASURING DEVICES . . . . . . . . . . . . . . . . . . . . . . . . . . . .

11.1 11.1 11.9 11.12 11.16 11.22 11.23 11.24

CHAPTER 12 PRELIMINARY TAB PROCEDURES 12.1 12.2 12.3 12.4 12.5 12.6

INITIAL PLANNING . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . CONTRACT DOCUMENTS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . SYSTEM REVIEW AND ANALYSIS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . THE AGENDA . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . PLANNING FIELD TAB PROCEDURES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . PRELIMINARY FIELD PROCEDURES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

12.1 12.1 12.2 12.4 12.5 12.6

CHAPTER 13 GENERAL AIR SYSTEM TAB PROCEDURES 13.1 13.2 13.3 13.4 13.5 13.6 13.7 viii

BASIC FAN TESTING PROCEDURES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . SYSTEM STARTUP . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . FAN TESTING . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . DEFICIENCY REVIEW . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . RETURN AND OUTSIDE AIR SETTINGS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . ANALYSIS OF MEASUREMENTS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . RECORDING DATA . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .


13.1 13.1 13.1 13.2 13.2 13.3 13.3

13.8 13.9 13.10 13.11 13.12 13.13 13.14 13.15 13.16 13.17 13.18 13.19 13.20 13.21 13.22 13.23 13.24 13.25

PROPORTIONAL BALANCING (RATIO) METHOD . . . . . . . . . . . . . . . . . . . . . . . 13.3 PERCENTAGE OF DESIGN AIRFLOW . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13.3 SYSTEM AIRFLOW . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13.5 BASIC OUTLET BALANCING PROCEDURES . . . . . . . . . . . . . . . . . . . . . . . . . . . 13.5 STEPWISE METHOD . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13.5 FAN ADJUSTMENT . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13.6 WET COIL CONDITIONS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13.6 AIRFLOW TOTALS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13.6 EXHAUST FANS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13.6 FAN DRIVE ADJUSTMENT . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13.6 DAMPER ADJUSTMENTS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13.7 DUCT TRAVERSES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13.7 SYSTEM DEFICIENCIES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13.7 FUME HOOD EXHAUST BALANCING PROCEDURES . . . . . . . . . . . . . . . . . . . 13.7 DUST COLLECTION AND EXHAUST BALANCING PROCEDURES . . . . . . . 13.8 AIR FLOW MEASUREMENTS ON DISCHARGE STACKS . . . . . . . . . . . . . . . . 13.11 INDUSTRIAL VENTILATION . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13.12 SELECTION OF INSTRUMENTS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13.12

CHAPTER 14 TAB PROCEDURES FOR SPECIFIC AIR SYSTEMS 14.1 14.2 14.3 14.4 14.5 14.6 14.7

INTRODUCTION . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . VARIABLE AIR VOLUME (VAV) SYSTEMS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . MULTI-ZONE SYSTEMS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . INDUCTION UNIT SYSTEMS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . DUAL DUCT SYSTEMS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . SPECIAL EXHAUST AIR SYSTEMS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . PROCESS EXHAUST AIR SYSTEMS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

14.1 14.1 14.13 14.14 14.14 14.16 14.17

CHAPTER 15 HYDRONIC SYSTEM TAB PROCEDURES 15.1 15.2 15.3 15.4 15.5 15.6 15.7

HYDRONIC SYSTEM MEASUREMENT METHODS . . . . . . . . . . . . . . . . . . . . . . 15.1 BASIC HYDRONIC SYSTEM PROCEDURES . . . . . . . . . . . . . . . . . . . . . . . . . . . 15.3 PIPING SYSTEM BALANCING . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 15.4 BALANCING SPECIFIC SYSTEMS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 15.5 VARIABLE VOLUME FLOW . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 15.9 PRIMARY-SECONDAR Y SYSTEMS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 15.11 SUMMER-WINTER SYSTEMS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 15.11

CHAPTER 16 TAB REPORT FORMS 16.1 16.2

PREPARING TAB REPORT FORMS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . DESCRIPTION OF USE . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

APPENDIX A

16.1 16.1

DUCT DESIGN TABLES & CHARTS DUCT DESIGN TABLES AND CHARTS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . HVAC EQUATIONS - (I-P) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . HVAC EQUATIONS - (SI) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . SI UNITS AND EQUIVALENTS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . SOUND DESIGN EQUATIONS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . FITTING EQUIVALENTS (WATER) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . PROPERTIES OF STEAM . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . STEAM PIPING (I-P) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . STEAM PIPING (SI) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . REFERENCES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

A.1 A.31 A.35 A.39 A.41 A.43 A.44 A.45 A.49 A.54

GLOSSARY . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

G.1

INDEX . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

I.1


ix

TABLES

5-1 6-1

Typical Fan Rating Table . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5.7 Typical Ratios of Damper to System Resistance for Flow Characteristic Curve . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 6.6 6-2 Guide to Use of Various Outlets . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 6.12 6-3 Recommended Return Air Inlet Face Velocities . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 6.14 6-4 Air Outlets and Diffusers Total Pressure Loss Average—in. wg (Pa) . . . . . . . . . . . 6.15 6-5 Supply Registers Total Pressure Loss Average—in. wg (Pa) . . . . . . . . . . . . . . . . . . 6.15 6-6 Return Registers Total Pressure Loss Average—in. wg (Pa) . . . . . . . . . . . . . . . . . . 6.15 8-1 Characteristics of Centrifugal Pumps . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8.3 8-2 Characteristics of Common Types of Pumps . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8.3 8-3 Flow vs Total Head (Cooling Tower Application) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8.11 9-1 Hydronic Trouble Analysis Guide . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9.8 11-1 Airflow Measuring Instruments . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11.9 11-2 Instruments for Hydronic Balancing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11.11 11-3 Hydronic Measuring Instruments . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11.11 11-4 Rotation Measuring Instruments . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11.13 11-5 Instrumentation for Air & Hydronic Balancing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11.16 11-6 Instruments for Air Balancing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11.17 11-7 Temperature Measuring Instruments . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11.21 15-1 Load-Flow Variations . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 15.10 A-1 Duct Material Roughness Factors . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.3 A-2 Circulation Equivalents of Rectangular Ducts for Equal Friction and Capacity (I-P) (2) Dimensions in Inches . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.5 A-2 Circulation Equivalents of Rectangular Ducts for Equal Friction and Capacity (I-P) (2) Dimensions in Inches (continued) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.6 A-3 Circular Equivalents of Rectangular Ducts for Equal Friction and Capacity (SI) (2) Dimensions in mm . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.7 A-3 Circular Equivalents of Rectangular Ducts for Equal Friction and Capacity (SI) (2) Dimensions in mm (continued) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.8 A-4 Velocities/Velocity Pressures (I-P) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.9 A-5 Velocities/Velocity Pressures (SI) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.10 A-6 Angular Conversion . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.10 A-7 Loss Coefficients for Straight-Through Flow . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.11 A-8 Recommended Criteria for Louver Sizing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.12 A-9 Typical Design Velocities for Duct Components . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.13 A-10 Elbow Loss Coefficients . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.14 A-1 1 Transition Loss Coefficients . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.17 A-12 Rectangular Branch Connection Loss Coefficients . . . . . . . . . . . . . . . . . . . . . . . . . . . A.19 A-13 Round Branch Connection Loss Coefficients . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.23 A-14 Miscellaneous Fitting Coefficients . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.27 HVAC Equations (I-P) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.31 A-15 Converting Pressure In Inches of Mercury to Feet of Water at Various Water Temperatures . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.33 A-16 Air Density Correction Factors (I-P) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.34 HVAC Equations (SI) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.35 A-17 Air Density Correction Factors (SI) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.38 A-18 SI Units And Equivalents . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.39 A-19 SI Equivalents . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.40 A-20 Sound Design Equations . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.41 A-21 Equivalent Length in Feet of Pipe for 90 Elbows . . . . . . . . . . . . . . . . . . . . . . . . . . A.43 A-22 Equivalent Length in Meters of Pipe for 90 Elbows . . . . . . . . . . . . . . . . . . . . . . . . A.43 A-23 Iron and Copper Elbow Equivalents . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.43 A-24 Properties of Saturated Steam (I-P) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.44 A-25 Properties of Saturated Steam (SI) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.44 A-26 Steam Piping (I-P) Flow Rate of Steam in Schedule 40 Pipe at Initial Saturation Pressure of 3.5 and 12 psig (Flow Rate expressed in Pounds per Hour) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.45 x


TABLES (continued) A-27 Comparative Capacity of Steam Lines at Various Pitches for Steam and Condensate Flowing in Opposite Directions (Pitch of Pipe in Inches per 10 Feet – Velocity in Feet per Second) . . . . . . . . . A.45 A-28 Pressure Drops In Common Use for Sizing Steam Pipe (For Corresponding Initial Steam Pressure) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.46 A-29 Length in Feet of Pipe to be Added to Actual Length of Run — Owing to Fittings — to Obtain Equivalent Length . . . . . . . . . . . . . . . . . . A.46 A-30 Steam Pipe Capacities for Low Pressure Systems (For Use on One-Pipe Systems or Two-Pipe Systems in which Condensate Flows Against the Steam Flow) . A.47 A-31 Return Main and Riser Capacities for Low-Pressure Systems—Pounds per Hour (Reference to this table will be made by column letter G through V) . . . . . . . . . A.48 A-32 Flow Rate in kg/h of Steam in Schedule 40 Pipe at Initial Saturation Pressure of 15 and 85 kPa Above Atmospheric . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.49 A-33 Comparative Capacity of Steam Lines at Various Pitches for Steam and Condensate Flowing in Opposite Directions . . . . . . . . . . . . . . . . . . . . . . . . . . A.49 A-34 Equivalent Length of Fittings to be Added to Pipe Run . . . . . . . . . . . . . . . . . . . . . . A.50 A-35 Steam Pipe Capacities for Low-Pressure Systems (For Use on One-Pipe Systems or Two-Pipe Systems in which Condensate Flows Against the Steam Flow) . A.51 A-36 Return Main and Riser Capacities for Low-Pressure Systems — kg/h . . . . . . . . A.52


xi

FIGURES

2-1 Heat Transfer by Conduction and Radiation . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2-2 Convection Heat Transfer . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2-3 Counterflow Airstreams . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2-4 Parallel Flow Airstreams . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2-5 Cross-flow Airstreams . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2-6 Parallel and Counterflow Heat Transfer Curves . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2-7 Psychrometric Chart (I-P) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2-8 Psychrometric Chart - Typical Condition Points (SI) . . . . . . . . . . . . . . . . . . . . . . . . . 2-9 Psychrometric Chart - Typical Condition Points . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2-10 Sensible Heating and Cooling (I-P) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2-1 1 Humidification and Dehumidification (I-P) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2-12 Psychrometric Chart - Processes . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2-13 Cooling and Dehumidifying (I-P) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2-14 Heating and Humidification . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2-15 Mixing of Two Airstreams (SI) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2-16 Tank Static Head . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2-17 Velocity Profile . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2-18 Pressure Changes During Flow in Ducts . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2-19 Sample Fitting Loss Coefficient Table . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2-20 Pump with Static Head and Suction Head . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2-21 Pump with Suction Lift . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3-1 Series-Parallel Circuit . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3-2 Single-Phase AC Service . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3-3 Current And Voltage-T ime Curves and Power Factor . . . . . . . . . . . . . . . . . . . . . . . . 3-4 220-Volt Three-Wire Delta Three-Phase Circuit . . . . . . . . . . . . . . . . . . . . . . . . . . . 3-5 220-Volt Delta Three-Phase Circuit with 110-V olt Single-Phase Supply . . . . . . . 3-6 120/208-Volt Four-Wire Wye Circuit . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3-7 Transformer with TaPped Secondary . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3-8 Typical Performance of Standard Squirrel Cage Induction Motors . . . . . . . . . . . . . 3-9 Interlocked Starters with Control Transformers . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3-10 VFD Added to Existing Air Handling Unit . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 4-1 Valve Throttling Characteristic Comparison . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 4-2 ATC Valve Arrangements . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 4-3 Typical Multiblade Dampers . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 4-4 Desktop Computer Displaying Status of Building HVAC Systems . . . . . . . . . . . . . . 4-5 Functional Block Diagram A Centralized Computer Control System . . . . . . . . . . . 4-6 HVAC Controls Panel with Original Pneumatic Controls. . . . . . . . . . . . . . . . . . . . . . 4-7 The Same HVAC Control Panel After Upgrading to Direct Digital Control (DDC). 4-8 Portable Computer Plugged Into Electronic Wall Thermostat During System Balancing. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-1 Centrifugal Fan Components . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-2 Characteristic Curves for FC Fans . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-3 Characteristic Curves for BI Fans . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-4 Characteristic Curves for Air Foil . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-5 Axial Fan Components . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-6 Characteristic Curves for Propeller Fans . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-7 Characteristic Curves for Vaneaxial Fans (High Performance) . . . . . . . . . . . . . . . . 5-8 Tubular Centrifugal Fan . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-9 Characteristic Curves for Tubular Centrifugal Fans . . . . . . . . . . . . . . . . . . . . . . . . . 5-10 Fan Class Standards (I-P) (SW BI Fans) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-1 1 Fan Class Standards (SI) (SW BI Fans) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-12 Drive Arrangements For Centrifugal Fans . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-13 Arrangement 1 In-Line Fans . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-14 Arrangement 4 in-line fans . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-15 Arrangement 9 in-Line fans . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-16 Centrifugal Fan Motor Locations . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-17 Direction of Rotation And Discharge . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . xii


2.2 2.2 2.3 2.3 2.4 2.4 2.9 2.10 2.11 2.12 2.13 2.14 2.15 2.15 2.17 2.20 2.21 2.22 2.24 2.28 2.29 3.2 3.2 3.3 3.4 3.4 3.4 3.5 3.7 3.9 3.10 4.3 4.4 4.4 4.7 4.8 4.9 4.10 4.10 5.1 5.1 5.2 5.2 5.2 5.3 5.3 5.3 5.4 5.4 5.4 5.5 5.8 5.9 5.9 5.10 5.11

FIGURES (continued) 5-18 Fan Total Pressure (TP) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-19 Fan Static Pressure (SP) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-20 Fan Velocity Pressure (VP) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-21 Tip Speed . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-22 System Resistance Curve . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-23 Operating Point . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-24 Variations from Design Air Shortage . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-25 Fan Law - RPM Change . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-26 Effect of Density Change (Constant Volume) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-27 Effect of Density Change (Constant Static Pressure) . . . . . . . . . . . . . . . . . . . . . . . . 5-28 AMCA Fan Test - Pitot Tube . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-29 Effect of Density Change (Constant Mass Flow) . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-30 Effects of System Effect . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-31 Fan Outlet Effective Duct Length . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-32 Non-Uniform Flow Conditions Into Fan Inlet . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 6-1 Constant Volume Fan-Powered Box . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 6-2 Bypass-Type Fan-Powered Box . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 6-3 Multiblade Volume Dampers . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 6-4 Flow Characteristics for a Parallel Operating Damper . . . . . . . . . . . . . . . . . . . . . . . 6-5 Flow Characteristics for an Opposed Operating Damper . . . . . . . . . . . . . . . . . . . . . 6-6 Volume Dampers . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 6-7 Surface (Coanda) Effect . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 6-8 Some Elements Affecting Body Heat Loss . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 6-9 Four Zones in Jet Expansion . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 6-10 Typical Supply Outlets . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7-1 Single Duct System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7-2 Typical Equipment for Single Zone Duct System . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7-3 Variable Air Volume (VAV) System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7-4 Terminal Reheat System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7-5 Induction Reheat System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7-6 Dual Duct High Velocity System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7-7 Multi-Zone System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7-8 System Layout (I-P Units) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7-9 System Layout (SI) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7-10 Fan Duct Connections . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8-1 Typical Centrifugal Pump Cross Section . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8-2 Descriptions of Centrifugal Pumps Used in Hydronic Systems . . . . . . . . . . . . . . . . 8-3 Coupling Alignment with Straight Edge . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8-4 Typical Required NPSH Curve . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8-5 Pump Curve for 1750 rpm Operation . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8-6 Typical Design Pump Selection Point (from Abbreviated Curve) . . . . . . . . . . . . . . 8-7 System Curve Plotted on Pump Curve . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8-8 Typical Open Systems . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8-9 Typical Cooling Tower Application . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8-10 System Curve for Open Circuit False Operating Point . . . . . . . . . . . . . . . . . . . . . . . 8-1 1 System Curve for Open Circuit True Operating Point . . . . . . . . . . . . . . . . . . . . . . . . 8-12 Pump Operating Points . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8-13 Multiple Pumps . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8-14 Pump and System Curves for Parallel Pumping . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8-15 Pump and System Curves for Series Pumping . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8-16 Gage Location . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8-17 Relative Gage Elevations . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8-18 Effect of Viscosity . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9-1 A Series Loop System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9-2 A One-Pipe System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9-3 Direct Return Two-Pipe System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9-4 Reverse Return Two-Pipe System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9-5 Example of Primary and Secondary Pumping Circuits . . . . . . . . . . . . . . . . . . . . . . . 9-6 Return Mix System Room Unit Controls . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

5.12 5.12 5.13 5.13 5.14 5.14 5.15 5.15 5.16 5.17 5.18 5.18 5.19 5.20 5.20 6.2 6.3 6.4 6.5 6.6 6.7 6.8 6.10 6.11 6.12 7.3 7.3 7.4 7.5 7.6 7.7 7.8 7.11 7.12 7.14 8.1 8.2 8.4 8.6 8.7 8.8 8.8 8.9 8.9 8.9 8.10 8.10 8.11 8.11 8.12 8.12 8.12 8.13 9.2 9.2 9.3 9.3 9.4 9.5 xiii

FIGURES (continued) 9-7 Four Pipe System Room Unit . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9-8 Boiler Piping for a Multiple-Zone, Multiple-Purpose Heating System . . . . . . . . . . 9-9 Water Cooled Condenser Connections for City Water . . . . . . . . . . . . . . . . . . . . . . . 9-10 Cooling Tower Piping System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9-1 1 Basic Piping Circuits for Gravity Flow of Condensate . . . . . . . . . . . . . . . . . . . . . . . 9-12 Basic Piping Circuits for Mechanical Return Systems . . . . . . . . . . . . . . . . . . . . . . . 9-13 Typical Two-Pipe Vacuum Steam System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9-14 Thermostatic Trap . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9-15 Inverter Bucket Trap . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9-16 Float and Thermostatic Trap . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9-17 Typical Connections to Finned Tube Heating Coils . . . . . . . . . . . . . . . . . . . . . . . . . . 10-1 Refrigerant Cycle . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 10-2 Locations of Thermal Bulbs . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-1 U-T ube Manometer Equipped with Over-Pressure Traps . . . . . . . . . . . . . . . . . . . . 11-2 Inclined-V ertical Manometer . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-3 Electronic/Multi-meter . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-4 Pitot Tube Connections . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-5 Pitot Tube . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-6 Magnehelic Gage . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-7 Rotating Vane Anemometer . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-8 Electronic Analog Rotating Vane Anemometer . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-9 Deflecting Vane Anemometer Set . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-10 Thermal Anemometer . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-1 1 Flow Measuring Hood . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-12 Calibrated Pressure Gages . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-13 Single Gage Being Used to Measure a Differential Pressure . . . . . . . . . . . . . . . . 11-14 Single Gage Being Used to Measure a Differential Pressure . . . . . . . . . . . . . . . . 11-15 Differential Pressure Gage . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-16 Chronometric Tachometer . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-17 Digital Optical Tachometer . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-18 Digital Contact Tachometer . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-19 Stroboscope . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-20 Multi-range, Dual Function (Optical/Contact Tachometer) . . . . . . . . . . . . . . . . . . 11-21 Glass Tube Thermometers . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-22 Dial Thermometer . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-23 Thermocouple . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-24 Thermistor Thermometer . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-24 Infrared Digital Thermometer . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-26 Resistance Temperature Detector . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-27 Electronic Thermometer . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-28 Sling Psychrometer . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-29 Digital Psychrometer . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-30 Thermohygrometer . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-31 Clamp-on Volt Ammeter . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-32 Accessing Automation System with Laptop Computer . . . . . . . . . . . . . . . . . . . . . . 11-33 Orifice as a Measuring Device . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-34 Flow Meter Types . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-35 Annular Flow Indicator . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-36 Calibrated Balancing Valve . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 12-1 Schematic Duct System Layout . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 12-2 Instruments Selected for a Specific Job . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13-1 Sample Supply Air Duct (Part) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13-2 Typical Air Diffuser CFM Measurement . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13-3 Measuring Exhaust Air Velocity on Lab Exhaust Hood with Sash Height . . . . . . 13-4 Example of Exhaust Hood Air Balance Label . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13-5 Sample Dust Collection Exhaust System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

xiv


9.5 9.7 9.12 9.13 9.14 9.16 9.16 9.17 9.17 9.18 9.18 10.2 10.5 11.1 11.2 11.2 11.3 11.4 11.5 11.6 11.6 11.7 11.7 11.8 11.10 11.12 11.12 11.13 11.14 11.14 11.15 11.15 11.15 11.18 11.18 11.19 11.19 11.19 11.20 11.20 11.22 11.22 11.23 11.23 11.24 11.25 11.26 11.26 11.26 12.3 12.5 13.4 13.6 13.7 13.8 13.9

FIGURES (continued) 14-1 Typical Variable Air Volume (VAV) System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 14-2 Open Loop Fan Volume Control . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 14-3 Closed Loop Fan Volume Control . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 14-4 Fan and System Curves, Constant Speed Fan . . . . . . . . . . . . . . . . . . . . . . . . . . . . 14-5 Fan and System Curves, Variable Speed Fan . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 14-6 Series Fan Powered VAV Unit . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 14-7 Parallel Fan Powered VAV Unit . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 14-8 Paper Strip at VAV Box Return Before Balancing . . . . . . . . . . . . . . . . . . . . . . . . . . 14-9 Paper Strip at VAV Box After Balancing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 14-10 Constant Fan VAV Box . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 14-1 1 Intermittent Fan VAV Box (Parallel) Cycle . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 14-12 Multi-zone System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 14-13 Dual Duct System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 14-14 Induction Unit System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 15-1 Hydronic Flow Measurement . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 15-2 External Ultrasonic Flow Sensor on Pipe with Insulation Removed . . . . . . . . . . . 15-3 Ultrasonic Flow Meter . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 15-4 Effects of Flow Variation on Heat Transfer 20F (11C) ∆t at 200F (93C) . . . 15-5 Percent Variation to Maintain 90% Terminal Heat Transfer . . . . . . . . . . . . . . . . . . 15-6 Chilled Water Terminal Flow Versus Heat Transfer . . . . . . . . . . . . . . . . . . . . . . . . . 15-7 Pump With Variable Speed Drive . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 15-8 Example of Primary and Secondary Pumping Circuits . . . . . . . . . . . . . . . . . . . . . . 15-9 Summer-Winter Systems . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A-1 Duct Friction Loss Chart (I-P) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A-2 Duct Friction Loss Chart (SI) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A-3 Duct Friction Loss Correction Factors . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A-4 Velocities/Velocity Pressures (I-P) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A-5 Air Density Friction Chart Correction Factors . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A-6 Louver Velocity . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . A-7 Elbow Equivalents of Tees at Various Flow Conditions . . . . . . . . . . . . . . . . . . . . . .


14.1 14.2 14.3 14.4 14.4 14.9 14.9 14.9 14.10 14.12 14.13 14.14 14.15 14.16 15.1 15.2 15.2 15.9 15.9 15.10 15.11 15.12 15.13 A.1 A.2 A.4 A.9 A.11 A.12 A.43

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CHAPTER 1

INTRODUCTION

CHAPTER 1 1.1

INTRODUCTION TO TAB WORK

1.1.1

New Buildings

Testing, adjusting and balancing (TAB) of all HVAC systems in a new building is needed to complete the installation and to make the systems perform as the de− signer intended. Assuming that the system design and installation meets the comfort needs of the building occupants, good testing, adjusting and balancing of the HVAC system provides occupant comfort with minimum en− ergy input. This is extremely important in this era of rising energy costs. It is also important to make sure all factory equipment startup service has been completed before beginning any TAB work. Most specifications on new building construction usually require a factory representative to be present during the initial startup and adjustment of central boilers, chillers, large variable speed motor drives, and cooling towers. This initial equipment checkout is also usually required to activate the factory warranties and are not be part of the TAB contractor’s responsibility. After this initial startup service has been completed, the TAB contractor should be in− formed that the systems are operating properly, that all safety interlocks and protective devices are function− ing, and the systems are ready to be balanced. The Testing, Adjusting, and Balancing or TAB phase of any building construction or renovation is intended to verify that all HVAC water and air flows and pres− sures meet the design intent and equipment manufac− turer’s operating requirements. It is rare to find an HVAC system of any size that will perform completely satisfactorily without the benefit of TAB work. This is why it is necessary for the designer to specify that TAB work be part of the HVAC system installation. A sam− ple TAB specification can be found in the Appendix. Commissioning services for any new building construction or renovation are intended to verify all HVAC, lighting, plumbing, electrical, and security systems operate properly and meet performance crite− ria.

INTRODUCTION It should be made clear that the Testing, Adjusting, and Balancing (TAB) services may be the only HVAC sys− tem testing services contracted on most projects, but TAB work is not intended to be ?commissioning." Most commissioning services are completed by firms having technicians experienced with each of the indi− vidual building systems mentioned above. These firms will usually subcontract the services of an independent TAB contractor to verify HVAC system balancing as part of their more inclusive commission− ing contract. 1.1.2

There are few buildings in existence that have not ex− perienced changes in internal loads and wall reloca− tions since they were designed and built. These build− ings should have their HVAC systems rebalanced to achieve maximum operating efficiency and comfort. Many buildings require rebalancing twice each year with the seasonal change from heating to cooling or the reverse. Firms with a good TAB team have had a natural lead−in to service contracts and retrofit work because the TAB work identifies system operating deficiencies. 1.2

THE TAB TECHNICIAN/TEAM

1.2.1

The Technician

Throughout this publication, TAB technician will be used to designate the person in charge of the TAB work being done on the HVAC system discussed. It will be apparent after reading this publication and observing TAB procedures on a complicated HVAC system that the TAB technician must be a highly skilled and knowledgeable individual. This person must know the fundamentals of airflow, hydronic flow, refrigeration and electricity and be familiar with all types of HVAC temperature control and refrigeration systems. They must also know how to take pressure, temperature and flow measurements; and be able to perform effective trouble−shooting. The days of bal− ancing using a wet finger and cigarette smoke are long gone! 1.2.2

Commissioning also includes the testing of all build− ing controls for each mode of operation to verify all systems are being sequenced correctly with each other, and that all interlocks are functioning. The commis− sioning agent must document the results of each equip− ment test performed as it is completed.

Existing Buildings

The Team

There are TAB jobs that can be done by one person. However, many HVAC systems need a TAB team to complete the TAB work in a reasonable time period. It is equally important that the other members of the TAB team be trained and become knowledgeable in


1.1

the basic fundamentals and procedures of TAB work. Many of the local Joint Apprentice Training Programs have TAB courses, and the International Training In− stitute (ITI) has a Testing Adjusting and Balancing Bu− reau (TABB) training program.

1.3

GENERAL REQUIREMENTS

In addition to having the training to meet the demand− ing requirements of a TAB technician, a complete cali− brated set of balancing instruments is necessary to do TAB work on any commercial or institutional project.

1.2

The required instruments are detailed in Chapter 11CTAB Instruments. Sample test report forms may be found in Chapter 16CTAB Report Forms. These TAB report forms may be copied and used by SMACNA Contractors who fol− low the procedures and methods outlined in this manu− al. The forms are preceded by a description of their use. A sample outline specification has been included in the Appendix that can be used by the HVAC system de− signer to obtain a good, accurate TAB report based on the methods and procedures found in this manual.


CHAPTER 2

HVAC FUNDAMENTALS

CHAPTER 2 2.1

HEAT FLOW

2.1.1

Introduction

Large HVAC systems must be designed by profession− al engineers who have received a high degree of educa− tion and practical experience in the fundamentals of heating, ventilating, and air conditioning. The TAB technician does not have to be an expert in the funda− mentals of HVAC systems, but must have had experi− ence in these systems and a basic knowledge of these fundamentals in order to perform a good balancing job and to understand what is happening. This chapter on HVAC fundamentals will include ba− sic thermodynamic and fluidic fundamentals that in− clude heat transfer, psychometrics, and fluid mechan− ics. 2.1.2

There are two fundamental laws of thermodynamics which can be stated in different, but equivalent ways: First Law

Energy can neither be created nor destroyed (the net increase in the energy content of a particular system in a given period is equal to the energy content of the ma− terial leaving the system, plus the work done on the system, plus the heat added to the system). 2.1.2.2

Second Law

It is impossible for a self−acting machine, unaided by any external agency, to convey heat from a body of lower temperature to one of higher temperature (heat flow always occurs from the higher temperature level to the lower temperature level). 2.1.3

scale of a thermometer, but the Celsius scale is used in the rest of the world. A typical temperature spectrum for the HVAC industry is where water freezes at 32F (0C) and boils at 212F (100C). The temperature at which a substance has no molecu− lar action is called absolute zero, which is −460F (−273C). The absolute temperature used in tempera− ture/pressure/volume calculations can be obtained by using the following equations: Equation 2-1 (I-P) R °F 460°F Where: R Absolutetemperature(Rankine) °F Fahrenheittemperature

Thermodynamics

Heat is one of the several forms of energy which can be converted by various methods to, or from, energy in mechanical, chemical, electrical, and nuclear forms. Thermodynamics is the science of heat energy and its transformations to and from these other forms of energy.

2.1.2.1

HVAC FUNDAMENTALS

Units of Measurement

The intensity of heat of a substance traditionally has been measured in the United States on the Fahrenheit

Equation 2-1 (SI) K °C 273°C Where: K Absolutetemperature(Kelvin) °C Celsiustemperature The quantity or amount of heat in a substance is mea− sured in British Thermal Units (Btu) which is the heat required to heat one pound of water one degree Fah− renheit. It is easy to realize that a swimming pool full of water at 95F, needs substantially more heat than a cup of water at 95F to increase each of them to 96F. In the metric system, the amount of heat required to heat one kilogram of water one degree Celsius is 4.18 kilojoules (kJ). 2.1.4

Heat Transfer

Heat flow is adding or removing heat at a given rate, which is measured in Btu per hour (Btuh) in the U.S. system and is measured in joules per second (J/s) or watts (W) in the metric system (1 J/s = 1 W). Figures 2−1 and 2−2 give examples of heat conduction, convection, and radiation, which are the three methods of heat transfer in environmental systems. Equation 2−2 may be used for heat flow through a ma− terial(s) that separates different air temperatures.


2.1

Radiant heat

Heat Flow Heat flows from warmer body to cooler body by conduction

from flame

Rod heated by flame becomes hot as heat flows by conduction from one end to the other.

FIGURE 2-1 HEAT TRANSFER BY CONDUCTION AND RADIATION Q A U Dt

Equation 2-2

Where: Q Rateofheattransfer Btuh(W) A Areaofsurface sq.ft.(m2) U Coefficientofheattransfer Dt Temperaturedifference °F(°C) ?Delta" ( ), as used above, usually indicates a small change or difference; in this case, Dt is the tempera− ture difference. In Equation 2−2, neither Dt or Q represent heat. In the past, heat was thought to be a tangible quantity like a gallon of water, a pint of milk, or a bushel of wheat. Despite this carryover from the past, one cannot feel ?hot," as heat is not a tangible quantity. What one does feel is temperature. Temperature can be ?hot" when compared to some other reference point such as 98.6F (37C) body temperature, but one cannot feel how much heat is in an object. The amount of heat con− tained within the object varies with the object. The amount of heat contained within the object varies with the object’s temperature, mass, and substance. The

amount of heat in any given object at any given tem− perature can be calculated, but the HVAC industry does not find that a particularly useful function. What is useful is to know how fast heat is given up from that object, or the rate of heat transfer (Q) expressed in Btu per hour or watts. There also are other equations for calculating ?Q". Figures 2−3, 2−4, and 2−5 show the difference between counterflow, parallel flow, and cross−flow airstreams in coils or heat exchangers. The importance of the il− lustration in Figure 2−6 is that the final temperature at− tained by both of the mediums is affected by the direc− tion of the flow of the two different mediums at the heat transfer points.

Example 2.1 (I−P) A room (70F) has two separate walls which have an unheated space on the other side. The wall exposed to outdoors (30F) is 20’ × 8’ and has a ?U" factor of 0.12. The 24’ 8’ wall exposed to the unoccupied space (55F) has a ?U" factor of 0.30. Which wall has the greatest heat loss?

Heating Coil

Steam or Hot Water in Pipes

Airflow

(A) Heat Flow by Natural Convection

(B) Heat Flow by Forced Convection

FIGURE 2-2 CONVECTION HEAT TRANSFER 2.2


B

A

B

A

B A

A

B A

B

A B

B

FIGURE 2-3 COUNTER FLOW AIRSTREAMS Solution Outside wall: Q A U Dt 20 8 0.12 (70°F 30°F) 160 0.12 40°F 768Btuh Inside wall: Q A U Dt 24 8 0.30 (70°F 55°F) 192 0.30 15°F 864Btuh (or the higher loss wall)

Example 2.1 (SI) A room (21C) has two separate walls which have an unheated space on the other side. The wall exposed to outdoors (0C) is 6 m 2.5 m and has a ?U" factor of 1.2. The 7 m 2.5 m wall exposed to the unoccupied space (12C) has a ?U" factor of 3.0. Which wall has the greatest heat loss? Solution Outside wall: Q =MA U nt = 6 2.5 1.2 (21_C 0C) =M15 1.2 21C = 378 watts Inside wall: Q =MA × U × nt = 7 × 2.5 × 3.0 × (21_C − 12C) =M17.5 × 3 × 9_C = 473 watts

FIGURE 2-4 PARALLEL FLOW AIRSTREAMS

(or the higher loss wall) To illustrate the difference between temperature and the heat flow rate, and to show that a system can be bal− anced to either, assume that the airflow being supplied to different rooms is to be balanced so that each room has the same temperature reading. This procedure can be used in existing buildings when original engineer− ing calculations are not available and when the build− ing is experiencing large differences in temperatures between rooms. The TAB technician would then at− tempt to balance the system by adjusting the airflow so that each room had the designed space temperature. The other balancing procedure usually is required in new buildings where the designer has calculated the airflow rates that normally establish equal tempera− tures for each of the various room spaces. The TAB technician then balances the airstream flow rate to that scheduled for each room. The rooms should achieve the desired temperatures if the design calculations were correct. If a change in temperatures is required after occupancy, the additional balancing should not be done at no cost unless there was a provision for this extra work included in the TAB contract. In any event, the TAB technician balances to the flow rate of the me− dium, which actually is balancing to the heat flow rate that is being transferred by the medium. Balancing by heat flow and temperature are therefore not the same. In Equation 2−2 (Q = A U nt), ?U" is the variable affected by building insulation materials. Insulation ?values" can become quite confusing. They can be given per inch of thickness of material or for the actual thickness of the material, such as a ?six inch thick batt." Values can be given as conductance or resistan− ce. Chapter 25 of the 2001 ASHRAE Fundamentals


2.3

A1

A2

”C” AIRSTREAM

each component of the wall is the only way the total resistance (RT) can be obtained for a particular wall construction. Other building enclosure surfaces are treated in a similar manner (windows, floors, ceilings, etc.). ?R" values of air surface films must be used as indicated After the total resistance is obtained by adding the in− dividual resistances, the ?U" value is obtained by tak− ing its reciprocal as shown in Equation 2−4. Insulating materials play a large role in reducing the heat flow rate into or from buildings, ducts, pipes, etc. Obvious− ly, exterior surfaces having high resistances or low ?U" values will have less heat transfer between the outside environment and the interior of the structure, duct, or pipe.

EXCHANGER C2

“A” AIRSTREAM

Equation 2-3 RT =R1 +R2 +R3 +. . . +Rn

FIGURE 2-5 CROSS-FLOW AIRSTREAMS

Equation 2-4 Handbook entitled Thermal and Water Vapor Trans− mission Data contains thermal resistance values for different materials. Conductivity (k) indicates how much heat will pass through an inch of a material. Conductance (C) is a somewhat similar value, but is for a given thickness of material. Resistances (R), which are the reciprocals of ?k" or ?C", can be added sequentially for the heat flow through combinations of different materials. Conduc− tivity (k) and conductance (C) values can not be added in this manner. Therefore, addition of the resistance of

U 1 RT Where: U = Coefficient of heat transfer C Btuh/ ft2N⋅N°F (W/m2N⋅N°C) RT = Total of the resistances The values of U for the SI system are about 17.6 per− cent of the values in I−P Units, i.e., a ?U" of 1.0 in I−P Units is 0.176 in SI units.

TB (WARM)

TA (HOT)

TA (WARM)

TB (COLD) Distance Through Exchanger

TEMPERATURE

TEMPERATURE

TA (HOT)

TA (WARM)

TB (WARM)

TB (COLD) Distance Through Exchanger

FIGURE 2-6 PARALLEL AND COUNTERFLOW HEAT TRANSFER CURVES 2.4


Example 2.2 An outside wall of a building has the following resist− ances:

Outside surface film

— 0.17

Masonry

— 1.60

Furring (air space)

— 0.94

Drywall

— 0.45

Inside surface film

— 0.68

Find the coefficient of heat transfer U for this wall.

Solution R T R1 R2 R3 R n R T 0.17 1.6 0.94 0.45 0.68 R T 3.84 U 1 1 0.26 RT 3.84 2.1.5 EQUIPMENT HEAT FLOW Heat flow in HVAC equipment is normally from a fluid (or gas) through a thin wall into another fluid (or gas); or it is into a transfer substance which moves into or to another cooler fluid (or gas) to deposit its energy. Major factors in the transfer of heat by conduction are: a.

temperature difference

b.

size and shape of the transfer surface

c.

type of fluid (or gas) and flow velocity

d.

conductivity of heat transfer material

e.

conductivity of the boundary layer

There also are many other factors to consider such as film coefficients, fouling, corrosion, condensables, frost or freezing, and poor maintenance.

Going back to the ?laws of thermodynamics" stated earlier in this chapter, the concept can be restated that the same amount of heat that is given up by one me− dium is gained by the other, and that all heat can be ac− counted for (energy can neither be created nor de− stroyed). This is not quite true in this atomic era, but it can be used as a basic principle for this TAB work. Equation 2-5 (I-P) Hydronic systems: Q 500 gpm Dt Where: Q Heatflow(Btuh) gpm Gallonsperminute(water) Dt Temperaturedifference(°F) Equation 2-5 (SI) Q 4190 m3s Dt or Q 4.19 Ls Dt Where: Q Heatflow(kW) m 3s Cubicmeterspersecond(water) Ls Literspersecond(water) Dt Temperaturedifference(°C) Note that in Equation 2−5 the 500 (4190 or 4.19) is a ?constant" that is used specifically for water. This constant will change when the system medium is other than water, such as a glycol mixture, steam, refriger− ant, or air. In fact, the comparable equation for sensible heat flow of air is shown below. Equation 2-6 (I-P) Air systems: Q 1.08 cfm Dt Where: Q Sensibleheatflow(Btuh) cfm Cubicfeetperminute(air) Dt Temperaturedifference(°F)


2.5

Equation 2-6 (SI) Q 1.23 Ls Dt Where: Q Sensibleheatflow(watts) Ls Literspersecond(air) Dt Temperaturedifference(°C) Sensible heat is defined as the heat associated with temperature differences only as measured by a ther− mometer. This is not affected by the method of heat transfer, such as radiation, convection, and conduc− tion. Transmission heat gains are those which occur through conduction of the heat through a surface as a wall. Convection has been taken into account by film coefficients on each side of the wallCwhether inside or outside. Example 2.3 (I−P) 30 gpm of water at 200F circulates through a heating coil. If 4000 cfm of air increases in temperature from 50F to 120F, determine the leaving temperature of the water. Solution Q 1.08 cfm Dt 1.08 4000 (120°F 50°F) 302, 400Btuh 500 gpm Dt Q 302, 400 Dt 500 gpm 500 30 20.16°F (T1 T2) T2% T1MM∆tNN200F ON20.16F N179.84F Water with a temperature of 180F is leaving the coil. Example 2.3 (SI) 2 L/s of 93F water circulates through a heating coil. If 2000 L/s of air increases in temperature from 10F to 50C, determine the leaving temperature of the wa− ter. Solution Q 1.23 Ls Dt 1.23 2000 (50°C 10°C) 98, 400watts(98.4kW) 4.19 Ls Dt Q 98.4kW 4.19 Ls 4.19 2Ls ∆tPM11.74CMM(T1 M–MT2 ) 2.6

T2 PMT1M – ∆tMM93CM–M11.74CM=M81.26C Water with a temperature of 81.3C is leaving the coil. 2.2

PSYCHROMETRICS

2.2.1

Introduction

Psychrometrics is a study of the thermodynamic prop− erties of moist air and the application of these proper− ties to the environment and environmental systems. Thermodynamics previously has been defined as the science of heat energy and its transformation, or change, from one form of energy to another. Since air is the final environment and one of the major fluids of the systems, whatever affects air affects the systems and the environment. Whatever happens to the air and the moisture it contains, under both natural circumstances and artificial conditions imposed by the systems and the environment, is of concern to the TAB technician. The language of psychrometrics, and to be able to use psychrometric charts and tables as tools to change existing conditions to those desired or re− quired, is a requirement for a good TAB technician. 2.2.2

Properties of Air

Dry air is an unequal mixture of gases consisting prin− cipally of nitrogen, oxygen, and small amounts of neon, helium, and argon. The percentage of each gas normally will be the same from sample to sample, al− though carbon dioxide and pollutants might be present in varying quantities. Air in our atmosphere, however, is not dry but contains small amounts of moisture in the form of water vapor, and the percentage may vary from sample to sample. This air−water mixture is the moist air referred to in the subject of psychrometrics. The amount of water vapor in atmospheric air normally represents less than 1 per− cent of the weight of the moist air mixture. If the aver− age weight of air is approximately 0.075 pounds per cubic foot (1.204 kg/m3), the moisture contained therein will weight less than 0.00075 pounds per cubic foot (0.012 kg/m3). This would seem to be an insignificant amount to cause so much concern. Normally, atmospheric air contains only a portion of the water it is able to absorb (partially saturated). If the proper conditions occur, air will absorb additional moisture until it can absorb no more, and it is then saturated. Air and water vapor be− have as though the other were not present. Each will act as an independent gas and exert the same pressure as if it were alone. The barometric pressure is the sum


of the two pressuresCthe partial pressure of the air, plus the partial pressure of the water vapor. The dry air component of the moist air exists only as a gas under all environmental conditions. it cannot be liquified by pressure alone, and therefore acts as a per− fect gas. However, water vapor does coexist with water as a liquid at all environmental temperatures, and it can be liquified by pressure. As it is not a perfect gas, the properties of water vapor are determined experi− mentally. 2.2.3

Air-V apor Relationship

Throughout the normal ranges of atmospheric pres− sures and temperatures, the air and water vapor mix− ture behaves as a perfect gas provided no condensation or evaporation takes place. A perfect gas is one where the relationship of pressure, temperature, and volume may be defined and predicted by Equation 2−7. Equation 2-7 PV R T

T 70°F(21°C) P 14.696psi(101.325kPa)] V 13.33ft 3lb(0.831m3kg) d 0.075lbft 3(1.204kgm3) Example 2.4 (I−P) Using Equation 2−9, any condition other than standard may be calculated if one of the final conditions is known. For example, if one pound of standard air was heated to 700F, as in a process application, find the new volume of air per pound. Solution VT V 2 1 2 13.33 460 700 T1 460 70 1160 13.33 29.2ft 3lb 530 Similar calculations may be made for any value of temperature so long as the pressure remains constant. Example 2.4 (SI) If one kilogram of standard air is heated to 370C, find the new volume of air per kilogram.

Where: P'=MAbsoluteM pressureMMlb/Mft 2M(kPa) V'=MVolumeMMft3M(m3) T'=MAbsoluteMtemperature 460M + FM(273M+MC) R'=MGas constant The actual value of R has little meaning here, but the fact that R remains constant for any given perfect gas is extremely important. It is possible to equate the P, V, T values for two different conditions for the same gas: Equation 2-8 P 1V 1 PV 2 2 T1 T2

Solution 0.831 (273°C 370°C) V 1T 2 T1 (273°C 21°C) 8 0.831 643° 1.817m 3kg 294°

V2

Example 2.5 (I−P Units) It is possible to make similar calculations which in− clude pressure variations. An example might be to find the correct volume for an altitude of 5000 feet. Assume that the temperature is the same at both points for con− venience.

If it is assumed that atmospheric pressure remains es− sentially constant at a given elevation on earth, then: Equation 2-9 VT V2 1 2 T1 Using Standard Air, which is the fixed reference for air conditioning calculations, the following properties oc− cur:

Solution P 1V 1 PV 2 2 , T 1 T 2 and P1V 1 P 2V2 T1 T2 P orV 2 V1 1 P2 The standard atmospheric pressure at 5000 feet is 24.90 inches of mercury (see Table A−16).


2.7

V 2 V 1

P1 13.33 29.92 16.02ft 3lb P2 24.90

Referring to the sample charts in Figures 2−7 and 2−8, note that the groups or families of curves have been la− beled to indicate which part or parts of the graph are for the different properties of most air. Definitions of the properties may be found in the Glossary.

Example 2.5 (SI) 2.2.4.1 Find the correct volume for an altitude of 1500 meters (T1 = T2 ) at standard conditions.

Solution V P 0.831 101.3 V2 1 1 P2 95.1(fromTableA 17) 0.989m3kg The conclusion to be reached from these examples is that large deviations from standard air temperature and pressure values require corrections to the calculations, while relatively small variations may be ignored. In the case of pressure, a correction normally is not used in applications below 2000 feet (600 m) above sea le− vel. Above 2000 feet (600 m) it becomes necessary to make the correction, since air has a significant reduc− tion in its ability to carry heat. By a series of calcula− tions and laboratory measurements, a long list of val− ues are obtained and are listed in Air Density and Correction Factor Tables in Appendix ACEngineer− ing Data, Tables and Charts. To be meaningful, these values must have a reference, which has been stated to be standard atmospheric pressure at sea level: 29.92 inches of mercury (101.325 kPa). Since the variations caused by pressure are not serious below an altitude of 2000 feet (600 m), the values obtained by maintaining the 29.92 inches of mercury (101.325 kPa) are ade− quate for use with HVAC systems. However, as seen in Example 2−4, temperature varia− tions in process systems can cause wide variations in air density: density = 1/29.2 = 0.034 lb/ft3, or a reduc− tion of 54 percent. For this reason, many engineering catalogs use standard cfm (scfm) which is ?cfm" cor− rected to ?standard air" conditions. Similar conditions apply to the SI system. 2.2.4

Using the Psychrometric Chart

A psychrometric chart is a series of graphs or curves arranged in such a way that, by knowing a specific val− ue of each of two different properties, a point can be obtained that will determine the values for all other properties under the same conditions. Charts are avail− able for different elevations above sea level, higher or lower temperatures, and many other variations. 2.8

Basic Grid, Humidity Ratio (Specify humidity)

The horizontal parallel and equidistant grid lines (with the scale displayed along the right side of the chart) in− dicate the grains of moisture per pound of dry air or pounds of moisture per pound of dry air. The bottom line of the chart is zero humidity and represents totally dry air. The chemical industry prefers to use mol−ra− tios, and grams of moisture per kilogram of dry air is used in the SI system. 2.2.4.2

Enthalpy (Total Heat)

Enthalpy lines are slanted from the top−left to the bot− tom−right. Enthalpy is designated by the letter ?h" and, as all values on the chart, is referred to a pound (kilo− gram) of dry air. This is the only value which does not change through various processes, such as heating− cooling, compression−expansion, and humidification− dehumidification. Other constant value lines are su− perimposed upon this basic grid by plotting and are neither equidistant nor parallel, although they may ap− pear to be so. 2.2.4.3

Dry Bulb Temperature

Constant dry bulb temperature lines on the chart are nearly vertical. They diverge slightly towards the top of the chart. The scale is at the bottom, along the dry air line (zero humidity ratio) from left to right, and val− ues are given in F in Inch−Pound units, English or cus− tomary, and in C in SI Units, the International System of Units, Metric. 2.2.4.4

Saturation Line, Dew Point and Wet Bulb

The curved upper borderline of the chart, which runs from the bottom left to the mid−top, is an experimen− tally plotted saturation line. It shows the maximum amount of water vapor (in pounds or kilograms) which can be associated with a pound (kilogram) of dry air at a given dry bulb temperature. Air is said to be saturated with moisture at this point. The temperature at this point is known as the saturation temperature. It can be seen from the saturation line that at higher temperatures air can hold more moisture than at lower temperatures. Conversely, the capacity to hold mois−


Pounds of moisture per pound of dry air Enthalpy at saturation Btu per pound of dry air Grains of moisture per pound of dry air 85 90 95 100 105 .35 180 .025 170 .024 160

150

.023

.021 140

130

.45

.020 .019 .018

120

.40

.022

.50

.55

.017 .60

110

.016 .65 .015

100 .014 90

80

.013 .012

.70 .75 .80 .85 .90 .95

.011 70

60

.010

Sensible heat factor

.009 .008

50 .007 40

.006 .005

30 .004 20

80% 60%

10

40%

0 25

30

.002 .001

20% 20

.003

35

40

45

50

55

60

65

70

75

80

85

90

95

100

105

0 110

Wet-Bulb, Dewpoint or Saturation Temperature F

FIGURE 2-7 PSYCHROMETRIC CHART (I-P) ture is less at lower temperatures. Therefore, if the al− ready saturated moist air is being cooled, the excess moisture instantly begins to separate from the moist air by condensation either in the form of fog (the left side of the saturation curve is the fog area) or as dew if it condenses on a cold surface. The point of saturation (saturation temperature) is also called the dew point. At this point the saturation tem− perature, the dry bulb temperature and the wet bulb temperature are all the same value. The saturation temperature values (also dew point or wet bulb tem− peratures) are shown in F (C) along the saturation line where it is crossed by the same value dry bulb lines. 2.2.4.5

Specific Volume

Specific volume lines run at a steep angle from top left to bottom right. The numerical values, along the bot− tom of the chart at the end of these lines, are given in cubic feet per pound of dry air in I−P Units (and in cu− bic meters per kilogram in the SI System).

The slant and spacing of the specific volume lines show that moist air becomes lighter (by expansion) with an increase in temperature and with an increase in moisture content (moist air is lighter than dry air at the same temperature). The specific volume of dry air can be found at the intersection with the bottom, zero humidity ratio line. The volume of the water vapor can be found by subtracting the volume of the dry air from the volume of the moist air. 2.2.4.6

Relative Humidity

The curves of constant relative humidity (RH) lie be− tween the zero humidity ratio line at the bottom and the curved saturation line above. The curvature decreases as curves approach the dry air line. Relative humidity expresses the proximity of the sub− ject moist air to that of saturated air at the same tem− perature. The saturation line represents 100 percent RH and the bottom line of the chart is 0 percent RH. Another term used to define proximity of the moist air to saturation is the degree of saturation which is the ra− tio of the humidity ratio of the moist air to that of the saturated moist air at the same temperature and pres−


2.9

30

28 30

26

H H

24

SENSIBLE HEAT TOTAL HEAT

25

=

22

HUMIDITY RATIO (w) GRAMS MOISTURE PER KILOGRAM DRY AIR

20

20

15

18

16

14

12 0.85 0.90

10

0.95 1.0

8

10

6 5

4

2

0

20

30

40

50

DRY BULB TEMPERATURE C

FIGURE 2-8 PSYCHROMETRIC CHART - TYPICAL CONDITION POINTS (SI) sure. The difference between both meanings is small but still noticeable within the comfort conditions. While the degree of saturation of 50 percent lies on the dry bulb temperature line directly in the middle be− tween 0 percent and 100 percent, the 50 percent RH point is slightly below the mid point. Measurement of relative humidity depends on changes in humidity responsive materials. While low cost and quite accurate, these instruments require frequent cal− ibration. 2.2.4.7

Wet Bulb Temperature

The wet bulb method was developed to obtain a more practical way to measure relative humidity and de− pends on the evaporation of water around the bulb of a mercury thermometer. Since evaporation and the de− pression of the wet bulb temperature depends on the relative humidity of the moist air, it can be used to measure relative humidity by use of conversion tables.

bulb temperatures are equal at this point and are also equal to the dew point temperature (saturation temper− ature). As the wet bulb lines run so close to the enthal− py lines, most psychrometric charts use the same lines for both wet bulb temperatures and enthalpy and show correction of either enthalpy values or wet bulb values by another family of curves. 2.2.4.8

Plotting Conditions

Consider point A, in Figure 2−9, representing summer outdoor design conditions. By finding the point which represents 95F DB and 78F WB, the values for the other properties are:

Dew point temperature = 71.98F Relative humidity = 48% Enthalpy (total heat) = 41.6 Btu/lb dry air Moisture content = 118 grains/lb dry air

At the saturation point, there is no evaporation. Since the wet bulb depression is zero, the wet bulb and dry 2.10

or 0.0169 lb moisture/lb dry air.


Enthalpy at saturation Btu per pound of dry air 85

90

Grains of moisture per pound of dry air 95

100

Pounds of moisture per pound of dry air

105 .35

180 .025

170 .024

.023

160

.40

.022 150

41.6 BTU/Lb

78F WB

.021

POINT “A”

.45

140

.020

.019

.50

130 .018 .55 120

.017 .60 .016

110

118 GR/Lb

.65 .015

71.9F DP

.70

100 .014

.75

48% RH

62.7F WB

.0169 Lb/Lb

.013

90

.80 .85

.012

.90 .95

80

28.3 BTU/Lb

.011

70

.010

.009


60

POINT “B”

65 GR/Lb

.008

50 .007

50% RH

55F DP

.006 40

.0093 Lb/Lb

.005 30 .004

.003

20 80%

Wet-Bulb Dewpoint or Saturation Temperature F

60%

.

40%

95F DB

75F DB

.002 10 .001

20% 0

Dry-Bulb F

20

25

30

35

40

45

50

55

60

65

70

75

80

85

90

95

100

105

0 110

FIGURE 2-9 PSYCHROMETRIC CHART - TYPICAL CONDITION POINTS Now consider point B in Figure 2−9 representing sum− mer indoor design conditions. By finding the point which represents 75F DB and 50 percent RH, the val− ues for the other properties are: Dew point temperature = 55F Wet bulb temperature = 62.7F Enthalpy (total heat) = 28.3 Btu/lb dry air Moisture content = 65 grains/lb dry air or 0.0093 lb moisture/lb dry air.

a change are the environmental systems. The designer, by knowing what design conditions must be satisfied, but without knowing the airflow rate, is able to plot the various changes in different portions of the system and the environment. The designer then is able to deter− mine what systems are capable of accomplishing the necessary results. In addition, the heat values are used in the design calculations to see immediately if the de− sign conditions are practical or impossible. But first, the various condition changes must be illus− trated and understood. For this purpose, skeleton psychrometric charts have been used in the related dia− grams, alternating between I−P Units and SI units. 2.2.5.1

In this way, any condition may be plotted on the chart for normal environmental systems. Other charts are available for plotting from tabular data for conditions not found here, but the need for such variations is not common. 2.2.5

Condition Changes on Psychrometric Charts

Using the chart, simple logic leads to the conclusion that there must be some way to change the properties of the air, initially at the condition of point A, to the conditions of point B. The means to accomplish such

Sensible Changes

Sensible heat, by definition, indicates only dry bulb temperatures changes. Therefore, any heating system not including humidification is a sensible heat process or system and is represented on the chart as a horizon− tal straight line. Conversely, any cooling system utiliz− ing a dry coil which does not dehumidify or where the surface of the coil does not fall to or below the air dew point is also a horizontal straight line. a.

Heating

Assume that air which is at 70F and 20 percent RH is heated to 105F. This process is indicated in Figure


2.11

2−10. Conditions other than those mentioned here have been determined from the psychrometric chart.

a change in latent heat. This process is indicated on the sample metric chart in Figure 2−11. Humidification and dehumidification are defined as the addition or subtraction of moisture from air. Since each is a change of state from liquid to gas and gas to liquid, each occurs at a constant dry bulb temperature, but a varying wet bulb temperature. Note that this is the same process as the addition or subtraction of latent heat and is also the same vertical line on the chart in Figure 2−11 at a constant dry bulb temperature. Hu− midification and dehumidification are both latent heat processes, and both are shown on the same chart.

105F DB 8% RH 20F DP 64F WB

70F DB 20% RH 28F DP 50F WB

HEATING

COOLING

FIGURE 2-10 SENSIBLE HEATING AND COOLING (I-P)

In this example, the only constant value is the dry bulb temperature; all other properties increase for humidifi− cation and decrease for dehumidification. Note that this process is essentially an illustration and cannot normally be reproduced in environmental systems. 2.2.5.3

By considering the values, it may be seen that the dew point has not changed and the total moisture content in grains/lb has not changed. The dry bulb temperature has gone up, the heat content has gone up, the wet bulb temperature has gone up, and the ability of the air to absorb moisture has gone up as indicated by the de− crease in relative humidity. The example is theoretical because moisture from people, cooking, infiltration, etc., will make some contribution to the moisture in the air. However, this is the design diagram for heating without deliberate humidification. b.

Cooling

Now consider the process and ignore the extraneous moisture sources noted above. In ideal conditions, air gives up its heat along the same line to maintain the oc− cupied spaces at the given 70F DB and 20 percent RH. A cooling coil, selected to cool air from 105F DB and 20F DP to 70F DB would also produce condi− tions along the same line as long as the coil surface temperature was above 20F. In most systems, this would be impractical if not impossible, since the coil would immediately clog with frost. 2.2.5.2

Latent Changes

Latent heat, by definition, involves a change of state to or from a fluid; and in the case of air, this means the addition or removal of moisture. It must be remem− bered that the dry bulb temperature does not change during this addition or removal of moisture. Therefore, a vertical line on the psychrometric chart between any two points at constant dry bulb temperature represents 2.12

Combination Changes

Combination sensible−latent heat processes are the rule in most systems. The addition or subtraction of la− tent and sensible heat appears as a combination pro− cess with all changes occurring simultaneously. The result is neither a horizontal nor vertical line but a slanted one tilted in the direction dictated by process. Referring to Figure 2−12, consider the general rules be− low based on the two endpoints of the process; the first being the initial condition of the air; and the second be− ing the final condition after the process or a portion of the air treatment has been completed. On the chart, all processes have the same initial point, and the arrow point indicates each arbitrary final point: Sensible heating is a horizontal line from left to right. Sensible cooling is a horizontal line from right to left. Humidification is a vertical line upward. Dehumidification is a vertical line downward. Heating and humidification is a line sloping upward to the right. Cooling and dehumidification is a line sloping down− ward to the left. Evaporative cooling is a line sloping upward to the left. Chemical dehydration or dehumidification is a line sloping downward to the right. Now consider the specific cooling−dehumidifying ex− ample which might represent the conditions obtained


Humidification—Addition of Moisture and Latent Heat by Evaporation of Steam Injection

Dehumidification— Removal of Moisture and Latent Heat by Condensation

41C DB 28.2C WB 24C DP 38% RH 90 kJ/kg 18.9 g/kg 41C DB 20.1C WB 6.8C DP 13% RH 58.9 kJ/kg 6.2 g/kg

FIGURE 2-11 HUMIDIFICATION AND DEHUMIDIFICATION (I-P) from a cooling coil using 100 percent recirculated air. Assume that the entering air and room conditions are 80F DB and 50 percent RH. Also assume that the cooling coil can produce leaving conditions of the air at 55F DB and 54F WB. The illustration of the pro− cess is shown in Figure 2−13. The line drawn between the initial and final conditions represents the change of air properties produced by the cooling coil and is conveniently drawn straight. In ac− tual fact, the process follows a curve, but the deviation is not usually important to the system analysis. To the coil designer, however, the curvature is critical, since it indicates the heat transfer conditions from point−to− point through the coil depth. The amount of moisture that was removed from the air (condensed on the coil) was 16 grains/lb of dry air (77T–T61T=T16). The reverse operation, heating and humidifying, could be explained by working the cooling−dehumidifying diagram backwards. However, a more practical ap− plication may be obtained by using a new diagram. As− suming that 21C DB, 40 percent RH air returns to a heating coil and that a humidifier has been added, the

combination process, assuming the required leaving conditions to be 41C DB and 38 percent RH, is illus− trated in Figure 2−14. 12.7 grams of moisture per kilo− gram of dry air was added in the process (18.9T–T6.2T=T 12.7) Equation 2−10 is used for the change in the total heat content of the air, including the moisture content. Equation 2-10 (I-P) Q(Total) 4.5 cfm Dh Where: Q Totalheatflow(Btuh) cfm Airflow Dh Enthalpydifference(Btuhlbdryair) Equation 2-10 (SI) Q(Total) 1.2 Ls Dh Where: Q Totalheatflow Ls Airflow Dh Enthalpydifference(kJkgdryair)


2.13

49

Pounds of moisture per pound of dry air

Grains of moisture per pound of dry air

48

Enthalpt at saturation Btu per pound of dry air

90

100

95

105

47

85

.35

46

180

45

.025 170

43

44

.024

.023

.40

42

80

160

41

.022 150 .45

39

40

.021

.020

.019

75

37

38

-0.1 Btu

140

.50

130

-0.2 Btu

.018 .55 120

Heating & Humidification

Evaporative Cooling

.017 .60 .016

-0.3 Btu

70

33

34

90% Relative Humidity

35

36

Humidification

110 .65 .015 .70

80% 30

100

65

28 27

26 25

55

50% -.06 Btu

23 22

Sensible Cooling

.80 .85

.012

.90

.010


.95

80

70

.009 -.08 Btu

21

.75

90

Sensible Heating

-.02 Btu

24

60% 60

20

18

Common Initial Reference

70%

Enthalpy deviation Btu per pound of dry air

29

.014

60

40%

18

17

.008 50

16

50 .007

11

13

40

20%

35

9

12

10

Chemical Dehydration or Dehumidification

Cooling & Dehumidification Dehumidification

14

12

15

30% 45

30

.006 40

.005 30 .004

40.1 Btu

.003

8

20 40.2 Btu

.002

60%

Saturation Temperature F

40.3 Btu

40%

10 40.4 Btu

.001

20% 40.5 Btu

0 60

55

65

70

80

75

85

90

100

95

105

0 110

14.0 cu ft

50

13.5 cu ft

45

13.0 cu ft

40

35

30

12.5 cu ft

20

25

14.5 cu ft per pound of dry air

Wet-Bulb, Dewpoint or

10%

80%

7

25

FIGURE 2-12 PSYCHROMETRIC CHART - PROCESSES Psychrometric charts base all the information given in content per pound (kilogram) of dry air. The standard equations are derived from these values and give a very close approximation of the actual calculation if all the conditions would have been worked out using the basic figures on the psychrometric chart. From any two given points on a psychrometric chart, the Btuh (watts) obtained for enthalpy is always equal to or greater than the Btuh obtained for sensible heat only. The reason for this is that the moisture contained in the air has heat content.

b. Locate the final condition on the chart as Point ?B". The wet bulb temperature is 62.7F and the en− thalpy is approximately 28.3 Btu per pound of dry air. c.

The decrease in enthalpy is: h = 41.6 − 28.3 = 13.3 Btu per pound

d.

Q (Total)

= 4.5 10,000 13.3 Q (Total) e.

Solution a. Locate the initial condition on the psychrometric chart, as Point ?A" in Figure 2−9. The corresponding wet bulb temperature is 78F, and the enthalpy is approximately 41.6 Btu per pound of dry air. 2.14

= 598,500 Btuh

Tonsofrefrigeration Btuh Btuhton 12, 000

Example 2.6 (I−P) Air at 95F DB and 48 percent RH enters a cooling coil at a rate of 10,000 cfm. If the air is cooled to a condi− tion of 75F DB and 50 percent RH, find the cooling load in Btuh, and in tons of refrigeration.

= 4.5 cfm h

598, 500 49.9tons 12, 000

Example 2.6 (SI) Air at 36C DB and 50 percent RH enters a coil at a rate of 5000 L/s. If the air is cooled to a condition of 24C and 50 percent RH, find the cooling load in watts and kilowatts (use Figure 2−8). Solution a. The enthalpy for 36C DB and 50 percent RH is approximately 85.3 kJ/kg.


AIR ENTERING 80F DB 67F WB 50% RH

AIR LEAVING 55F DB 54F WB 94% RH

FIGURE 2-13 COOLING AND DEHUMIDIFYING (I-P)

AIR LEAVING

AIR ENTERING

21C DB 40% RH 6.8C DP 13.1C WB 37.1 kJ/kg 6.2 g/kg

41C DB 38% RH 28.2C WB 24C DP 90 kJ/kg 18.9 g/kg

FIGURE 2-14 HEATING AND HUMIDIFICATION


2.15

b. The enthalpy for 24C and 50 percent RH is approximately 48.0 kJ/kg. c.

The decrease in enthalpy is: ∆h = 85.3 − 48.0 = 37.3 kJ/kg

d.

Q (Total) = 1.2 x L/s × h = 1.2 × 5000 × 37.3 Q (Total) = 223,800 (223.8 kW)

It is sometimes necessary to calculate the weight of the air. This is shown in the solution to Example 2.7 (I−P). The average person is not used to thinking of air as having weight. Air is a fluid and has weight just as wa− ter is a fluid and has weight. As a reminder, ?Standard air," at 0.075 pounds per cubic foot (1.204 kg/m3) is very much lighter than water at 62.4 pounds per cubic foot (1000 kg/m3). The specific volume of standard air is the reciprocal: 1 0.075 13.33 cubic feet per pound 1/0.075 = 13.33 cubic feet per pound (1/1.204 = 0.8305 m3/kg).

Solution 10, 000cuftmin 750.2poundsperminute 13.33cuftlb 750.2 60 minutes = 45,012 pounds per hour 45,012 13.3 (∆h) = 598,660 Btuh 598,660/12,000 = 49.9 tons The total heat content of air also can be taken from tables when the wet bulb temperature is known.

Example 2.7 (SI) Obtain the solution to Example 2.6 (SI) using the weight of the airflow volume.

Solution 5.0m 3s 6.02kgs 0.83m 3kg 6.02kgs 37.3kJkg 224.5kJs

Example 2.7 (I−P) 1W 1Js, so Obtain the solution to Example 2.6 (I−P) using the weight of the airflow volume.

2.16

224.5kJs 224.5kW


85.3

MIXED AIR

_ _

Q 3=10,000 l/s t 3=DB = ?3 WB3 = ?

C

_

Point B

_ _

56.3

49

_ Condition of mixture (Point C) 35_C DB

_

Point A

FIGURE 2-- 15 MIXING OF TWO AIRSTREAMS (SI)

2.2.5.4

Airstream Mixtures

Mixtures of two or more airstreams are a common requirement of the environmental system. The mixing of outside and return air on the entering side of the air cooling and heating equipment is the common method of introducing the outside ventilation air to the system. Outside air can be one of the greatest heating and cooling loads in more extreme climates.

 Assume that the airstreams are being mixed in a 50 percent ratio which causes the mixed point to be directly between points “A” and “B”.  Then assume that the damper moves to a position to take in less hot outside air (“B”). This will cause point “C” to move away from “B” (the line “B” to “C” gets longer when less air is taken in). If point “C” coincides with point “A,” 100 percent return air “A” will be used.

Figure 2-15 illustrates the mixing of an outside airstream and a return airstream, which usually is found in most air handling unit system applications.

 The mixed air temperature will be closest to the air temperature of the largest airstream.

When two airstreams are mixed and are plotted in graph form, the following steps should be used:

 Only the dry bulb air temperature can be obtained by using this method or by using Equation 2-11.

HVAC SYSTEMS Testing, Adjusting & Balancing  Third Edition

2.17

Equation 2--11 Xo T o + X r T r Tm = 100

b) From Figure 2-15:

Where:

27_C WB = 85.3 kJ/kg

Tm = Temperature of mixed air — _F (_C) Xo = Percentage of outdoor air — %

Using Equation 2-12:

To = Temperature of outdoor air—_F (_C)

X oH o + X rH r 100 20% × 85.3 + 80% × 49.0 Hm = 100 1706 + 3920 Hm = = 56.26 kJ∕kg 100 On a psychrometric chart, 56.26 kJ/kg = approximately 19.7_C wet bulb temperature.

Xr = Percentage of return air — % To = Temperature of return air—_F (_C) The wet bulb of the mixed airstream can be obtained by substituting the enthalpies of the two airstreams in Figure 2-15 in Equation 2-12 and calculating the enthalpy of the mixed airstream. From this value, the mixed airstream wet bulb temperature can be obtained from the psychrometric chart. Equation 2--12 X oH o + X rT r Hm = 100 Where: Hm = Mixed air enthalpy — Btu/lb (kJ/kg) Xo = Percentage of outdoor air — % Ho = Outdoor air enthalpy — Btu/lb (kJ/kg) Xr = Percentage of return air — % Ho = Return air enthalpy — Btu/lb (kJ/kg) Example 2.8 (SI) Calculate the dry bulb and wet bulb temperatures of the mixed airstream of Figure 2-15 using equations and by plotting in graph form on a psychrometric chart. Solution a)

Using Equation 2-11 for the dry bulb temperature: X oT o + X rT r 100 20% × 35˚C + 80% × 24˚C Tm = 100 700 + 1920 Tm = = 2620 = 26.2˚C (DB) 100 100 Tm =

2.18

17_C WB = 49.0 kJ/kg

Hm =

The psychrometric chart provides a quick method for calculating the mixed airstream conditions using only a scale as is shown in Figure 2-15. First, plot the two conditions (“A” and “B”) on the chart and draw a straight line between the two. Divide this distance in proportion to the mixed air quantities, and to scale, plot the mixed air point “C” so that it is closest to the conditions of the largest original quantity in the mixture. This is usually closest to the lowest dry bulb point since outside air quantities are usually less than 50 percent. (If they are 100 percent, no mixture determination is required.) The point “C” represents what mixture conditions should be if the air quantity proportions are correct. All of the properties of the mixture at point “C” are immediately available from the psychrometric chart (26.2_C DB and 19.7_C WB are two of them). One common error made by many novices, is the improper location of the mixed air point on the charts. Some reverse the ratio of the mixing streams, causing the mixed point shown in Figure 2-15 to occur near the top right hand point “B”. When two airstreams are mixed and are plotted in graph form, the following steps should be remembered: Assume that the air streams are being mixed a 50 percent ratio which causes the mixed point to be directly between points “A” and “B”. Then assume that a damper moves to a position to take in less hot outside air (at “B”). This will cause the point to move away from “B” (as the line “B” to “C” gets longer when less air is taken in). By the time it reaches “A”, 100 percent of the air of quality “A” will be used. It should be noted that the use of the psychrometric chart in the design to determine the properties of the


mixture does not establish the airflow quantity. This determination must be made independently. The de− signer must establish the total air quantity required from the sensible heat load and the outside air quantity from the design ventilation requirements. 2.2.5.5

Related Tables and Equations

Air (standard) density = 0.075 lb/ft3 (1.204 kg/m3 ) Water (standard) density = 62.4 lb/ft3 (1000 kg/m3) Specific volume is the reciprocal of density and is used to determine cubic feet of volume if the pounds of weight are known:

Chapter 6, Psychrometrics of the 2001 ASHRAE Fun− damentals Handbook has more detailed theory and data on this subject along with the necessary equations and psychrometric tables. Chapter 8, Thermal Comfort includes the comfort zone and effective temperature scale superimposed on a standard psychrometric chart. 2.3

FLUID MECHANICS

When the word fluid is mentioned, the average person thinks in terms of water. However, the full definition of fluid in the Glossary is ?gas, vapor, or liquid." In TAB work, the word fluid normally means air (from the atmosphere), water (or a heat transfer fluid), steam, refrigerants, and occasionally, a few other gases. This section will contain a detailed description of the behavior of air and water, as the properties of these two fluids affect most TAB work. Air and water generally have similar fluid properties except that the numerical values assigned to each property vary considerably. 2.3.1

Compressibility

For TAB purposes, water cannot be compressed. Air can be compressed and the volume of air can be pre− dicted by using Equations 2−7 to 2−9. 2.3.1.2

Water (Standard) specific volume = 0.016 ft3/lb(0.001 m3/kg) Standard Conditions for air as used above correspond to dry air to 70F (21C) and at an atmospheric pres− sure of 29.92 in. Hg. (101.325 kPa). For water, stan− dard conditions are 68F (20C) at the same baromet− ric pressure. 2.3.1.3

Weight, Density and Specific Volume Relationships

In TAB work, weight is measured in pounds (kilo− grams) and density in pounds per cubic foot (kg/m3); therefore, for standard conditions:

Specific Heat

The following specific heat values at standard condi− tions were used to develop Equations 2−5 (for water) and 2−6 (for air). Water: Specific Heat (Cp ) = 1.00 Btu/lbF (4190 J/kg C)

Fluid Properties

The basic categories of fluid properties are state, com− pressibility, viscosity, weight or density, volume or specific volume, volatility, specific heat and heat con− tent; but only those important to TAB technicians will be discussed. 2.3.1.1

Air (Standard) specific volume = 13.33 ft3/lb (0.831 m3/kg)

Air: Specific heat (Cp ) = 0.24 Btu/lbF (1000 J/kg C) Air Equation 2−6 only applies to sensible heat transfer (that which affects the dry bulb thermometer). 2.3.2

Fluid Statics

Static head is the pressure developed by the weight of the fluid at rest (not moving) in a system. The static head of air is insignificant and is ignored in TAB cal− culations, but not the static head of water. As Standard Atmospheric Pressure is measured at 14.696 psi (101.325 kPa), then Absolute Pressure is obtained by adding the 14.696 psi (101.325 kPa), atmospheric pressure to the gage pressure of the system static head.


2.19

To obtain the weight of one foot of water in psi and also absolute pressure: 62.4lbsqft 1footofwater 144sqinsqft = 0.433 psi gage pressure (psig) or 1 ft of water= 0.433 + 14.696 =15.129 psia (absolute pressure) Other I−P Unit pressure/weight/height relations are: 1 inch of mercury (Hg) = 13.6 inches of water gage (in.wg) 1 foot of mercury (Hg) = 5.89 psi 1 psi = 2.31 ft wg = 2.04 in.Hg 14.696 psi = 29.92 in.Hg = 33.9 feet of water gage (ft wg) In the SI system: 1 meter (water) = 9.807 kPa

1 millimeter (mercury) = 133.32 kPa From the above relationships, it can be seen that using water instead of mercury in a U−tube manometer for a pressure of 15 psi (103.4 kPa) would be quite cumber− some. However, it would be feasible to use when mea− suring the pressure of HVAC duct systems. Figure 2−16 shows the relationship between height and pressure of an open tank of water. The hydrostatic head at point B is 30 feet (9 m) and the fluid head at point B is 30 feet of water (9 m) which is equal to 13.0 psi gage (90 kPa) and to 27.7 psia (191.3 kPa a). 2.3.3 2.3.3.1

Fluid Dynamics Velocity

The term ?dynamics" is used to describe the condi− tions of motion of a fluid or gas in a system. The ?ve− locity" of the fluid is based on the cross−sectional area of the conduit (pipe or duct) through which it is flow− ing and the volume of fluid within the conduit. When there is no turbulence, the velocity varies within the conduit as shown in the diagram in Figure 2−17, ?Velocity Profile." This phenomenon, known as the velocity profile, is caused by the friction between the conduit walls and the fluid.

OPEN TANK MAINTAINS

ATMOSPHERIC PRESSURE

CONSTANT WATER LEVEL

14.7 psi (101.3 kPa)

GAGE 10ft(3m)

HYDROSTATIC HEADS

30ft(9m) 50ft(15m)

PRESSURES

ABSOLUTE PRESSURES

4.3 psi (29.7 kPa)

19.0 psi (131.1 kPa a)

13.0 psi (89.7 kPa)

27.7 psia (191.1 kPa a)

21.7 psi (149.7 kPa)

36.4 psia (251.2 kPa a)

30.3 psi (209.1 kPa)

45.0 psia (310.5 kPa a)

70ft(21m)

FIGURE 2-16 TANK STATIC HEAD

2.20


V min (Zero at Wall) V avg

V max

FIGURE 2-17 VELOCITY PROFILE

The only purpose of this device is to produce a pressure sufficient to overcome the resistance of the system to the flow of the fluid. The pressure produced is indi− cated by the pressure difference from the pump or fan inlet to the pump or fan discharge. This is exactly equal to the system resistance to flow and, in the case of wa− ter, the elevation differences for the amount of fluid being pumped. A measurement of the difference be− tween the inlet and discharge pressures of a pressure device of any kind is then a measurement of the system resistance at a particular flow rate. 2.3.4

When the flow is ?turbulent," the friction rate in− creases, heat transfer through the walls of an exchang− er increases, and usually, so does the system noise that is created by the fluid flow. Q AV

Equation 2-13

Where: Q = Fluid flow A = Area

Air

Water

cfm (L/s)

gpm (L/s)

ft2 (m2)

ft2 (m2)

V = Velocity fpm (m/s) fpm (m/s) *Correction constants are needed so that the units in the equation are compatible. 2.3.3.2

Friction

It has been indicated that the flow of the fluid is re− sisted by a well known paradox of nature called fric− tion. Friction is the natural resistance caused by a sub− stance with which it is in contact. One substance may be stationary and the other moving or both may be moving at different velocities. If it were possible to start the fluids flowing in a system and then eliminate friction losses of the conduit and the dynamic losses of the fittings, it would be possible to eliminate all power consuming pumping equipment in closed systems. The only purpose of the pumps is to overcome these losses which result from the flow they produce. 2.3.3.3

Pressure

Pressure is the force required to overcome the friction and dynamic losses of a system. This pressure is pro− duced by a pumping device which in HVAC systems may be a circulating pump, fan, or a gaseous refriger− ant compressor.

2.3.4.1

Air System Basics Duct Pressure Changes

The pressures in air systems are simpler than those in hydronic systems because the weight of air in the sys− tems is ignored in most calculations. The resistance to airflow, imposed by a duct system, is overcome by the fan, which supplies the energy (in the form of total pressure) to overcome this resistance and maintain the necessary airflow. Figure 2−18 illustrates an example of the typical pressure changes in a duct system with the total pressure and static pressure grade lines in ref− erence to the atmospheric pressure datum line. In air conditioning and ventilating work, the pressure differences are ordinarily so small that incompressible flow is assumed. Relationships are expressed for air at a standard density of 0.075 lb/ft3 (1.204 kg/m3) and corrections are necessary for significant differences in density due to altitude or temperature. Static pressure and velocity pressure are mutually convertible and can either increase or decrease in the direction of flow. To− tal pressure, however, always decreases in the direc− tion of airflow. At any cross−section, the total pressure (TP) is the sum of the static pressure (SP) and the velocity pressure (Vp ): TP SP Vp

Equation 2-14

For all constant−area straight duct sections, the change in static pressure losses are equivalent to the total pres− sure losses because the change in velocity pressure (Vp ) equals zero, as the velocity is constant. These pressure losses in straight duct sections are termed friction losses. Where the straight duct sections have smaller cross−sectional areas, such as duct sections BC and FG in Figure 2−18, the pressure lines fall more rap− idly than those of the larger area ducts (pressure losses increase as the square of the velocity). When the duct cross−sectional areas are reduced, such as at converging sections B (abrupt) and F (gradual),


2.21

A

B

C

D

E

F

G

AIR FLOW

ENTRY

H EXIT Diffuser

TP

SP

TP

Vp

ATMOSPHERIC PRESSURE SP TOTAL PRESSURE (TP)

VELOCITY PRESSURE (V ) p

STATIC PRESSURE (SP)

FIGURE 2-18 PRESSURE CHANGES DURING FLOW IN DUCTS both the velocity and velocity pressure increase in the direction of airflow and the absolute value of both the total pressure and static pressure decreases. The pres− sure losses at points B and F are dynamic losses. Dynamic losses are due to changes in direction or ve− locity of the air and occur at transitions, elbows, and duct obstructions, such as dampers, etc. Dynamic losses can be expressed as a loss coefficient (the constant which produces the dynamic pressure losses when multiplied by the velocity pressure) or by the equivalent length of straight duct which has the same loss magnitude. Increases in duct cross−sectional areas, such as at di− verging sections C (gradual) and G (abrupt), cause a decrease in velocity and velocity pressure, a continu− ing decrease in total pressure and an increase in static pressure caused by the conversion of velocity pressure to static pressure. This increase in static pressure is commonly known as static regain and is expressed in terms of either the upstream or downstream velocity pressure. 2.22

At the exit fitting, section H, the total pressure loss co− efficient may be greater than one upstream velocity pressure, equal to one velocity pressure, or less than one velocity pressure. The static pressure just upstream of the discharge fit− ting can be calculated by subtracting the upstream ve− locity pressure from the total pressure upstream. The entrance fitting at section A in Figure 2−18 also may have total pressure loss coefficients less than 1.0 or greater than 1.0. These coefficients are references to the downstream velocity pressure. Immediately downstream of the entrance, the total pressure is sim− ply the sum of the static pressure and velocity pressure. Note that on the suction side of the fan, the static pres− sure is negative with respect to the atmospheric pres− sure. However, velocity pressure is always a positive value. It is important to distinguish between static and total pressure. Static pressure is the blow−up pressure (like a balloon) which commonly has been used as the basis for duct system design. Total pressure determines how much energy must be supplied to the system to main−


tain airflow. Total pressure always decreases in the di− rection of airflow. But static pressure may decrease, then increase in the direction of airflow (as it does in Figure 2−18) and may go through several more in− creases and decreases in the course of the system. It can become negative (below atmospheric) on the dis− charge side of the fan, as demonstrated by points G and H. The distinction must be made between static pres− sure loss (sections BC or FG) and static pressure change as a result of conversion of velocity pressure (section C or G). The total resistance to airflow is noted by nTPsys in Figure 2−18. Since the prime mover is a vaneaxial fan, the inlet and outlet velocity pressures are equivalent; i.e. nTPsys = nSPsys . When the prime mover is a cen− trifugal fan, the inlet and outlet areas usually are not equal, thus the suction and discharge velocity pres− sures are not equal, and obviously nTPsys nSPsys . If one needs to know the static pressure requirements of a centrifugal fan, knowing the total pressure require− ments, the following relationship may be used: Equation 2-15 FanSP TP d TP s Vp d (orasSP TP Vp) FanSP SP d TP s where the subscripts d and s refer to the discharge and suction sections respectively of the fan.

Above 2000 feet (600 m) altitude, below 50F (10C), or above 90F (32C) temperature, the friction loss obtained from Tables A−1 and A−2 must be corrected for the air density. Table A−12 presents the correction factors for temperature and altitude. The actual air vol− ume is used to find the friction loss from Table A−1 and A−2 and this loss is multiplied by the correction factor or factors from Tables A−3 and A−4 to obtain the actual friction loss. 2.3.4.3

HVAC duct systems usually are sized first as round ductwork. Then, if rectangular ducts are desired, duct sizes are selected to provide flow rates equivalent to those of the round ducts originally selected. Tables A−5 and A−6 give the circular equivalents of rectangu− lar ducts for equal friction and flow rate. Note that the mean velocity in a rectangular duct will be less than in its circular equivalent. Rectangular duct sizes should not be calculated direct− ly from the actual duct cross−sectional area. Tables A−5 and A−6 should be used. If this is not done, the resulting duct sizes will be smaller, with a greater velocity and friction loss for the given airflow. 2.3.4.4

2.3.4.2

Circular Equivalent for Rectangular Ducts

Dynamic Losses

Friction Losses

Pressure drop in a straight duct is caused by surface friction, and varies with the air velocity, the duct size and length, and the interior surface roughness. Friction loss is most readily determined from Air Friction Charts, Tables A−1 and A−2 in the Appendix. They are based on standard air with a density of 0.075 lb/ft3 (1.204 kg/m3 ) flowing through average, clean, round, galvanized metal ducts (with an absolute roughness factor of 0.0003 ft. The values may be used without correction for temperatures between 50F (10C) and 90F (32C) and for altitudes up to 2000 feet (600 m) and for any relative humidity. Beyond those limits, corrections should be made for other than standard air densities, and when using other duct materials or flex− ible duct. Tables A−3 and A−4 are new tables and charts that pro− vide correction factors for ducts of materials other than hot−dipped galvanized sheet metal construction. The correction factor is multiplied by the friction loss ob− tained from Table A−1 and A−2 for each straight duct section.

Where turbulent flow is present, brought about by sud− den changes in the direction or magnitude of the veloc− ity of the air flowing, a greater loss in total pressures takes place than would occur in a steady flow through a similar length of straight duct having a uniform cross−section. The amount of this loss in excess of straight−duct friction is termed dynamic loss. 2.3.4.5

Duct Fitting Loss Coefficients

The duct fitting loss coefficient ?C" is dimensionless and represents the number of velocity heads lost at the duct transition or bend. Values of the loss coefficient for elbows and other duct elements have been deter− mined experimentally or computed and can be found in tables in the SMACNA HVAC Systems 9 Duct De− sign manual or the ASHRAE Fundamentals Hand− book. Tables A−4 and A−5 which show the relation of veloc− ity to velocity pressure for standard air, can be used to find the dynamic pressure loss for any duct element whose loss coefficient ?C" is known.


2.23

Fitting loss coefficient Velocity pressure—in.wg

Tee, 45 Entry, Rectangular Main and Branch Use the Vp of the upstream section Ac

Branch, Coefficient C

Qc

Vc

V / V b c

Qs Ab

Qb

Vs Ac = A s

As

Qb/ Qc 0.1

0.2

0.3

0.4

0.5

0.6

0.7

0.2

0.91

0.4

0.81

0.6

0.77

0.72

0.70

0.8

0.78

0.73

0.69

0.66

1.0

0.78

0.98

0.85

0.79

0.74

1.2

0.90

1.11

1.16

1.23

1.03

0.86

1.4

1.19

1.22

1.26

1.29

1.54

1.25

0.92

1.6

1.35

1.42

1.55

1.59

1.63

1.50

1.31

1.8

1.44

1.50

1.75

1.74

1.72

2.24

1.63

0.79

FIGURE 2-19 SAMPLE FITTING LOSS COEFFICIENT TABLE

TP C Vp

Equation 2-16

Equation 2-17 (SI) Q V 1000A Where:

Where:

V = Velocity (m/s) TP = Total pressure loss C in. wg (Pa) Q = Duct airflow (L/s) C = Fitting loss coefficient Vp = Velocity pressure C in. wg (Pa) The velocity pressure (Vp) used for rectangular duct fittings must be obtained from the velocity (V) ob− tained by using the following equation:

A = Duct cross−sectional area (m2) Where different areas are involved, letters with or without subscripts are used to denote the area at which the mean velocity is to be calculated, such as A for in− let area, A1, for outlet area and Ao for orifice area, etc. The equation for obtaining the velocity pressure (Vp) from the velocity is:

Equation 2-17 (IP)

Q V A

Equation 2-18 (I-P)

Vp V 4005

2

Equation 2-18 (SI) V p 0.602V 2

Where: V = Velocity (fpm) Q = Duct airflow (cfm)

Where: Vp = Velocity pressure C in. wg (Pa) V = Velocity C fpm (m/s)

A = Duct cross−sectional area (sq ft)

2.24

Commonly used fitting loss coefficient tables ex− tracted from the SMACNA HVAC Systems9Duct De− sign manual can be found in Tables A−17 and A−20.


Equation 2-19

Example 2.8 (I−P) Q2 rpm rpm2 1 Q1

A main duct has an airflow of 10,000 cfm at 2000 fpm. A 45 entry tap is used for the branch duct that requires 3,000 cfm at 1,600 fpm. Find the pressure loss to the branch of the fitting.

d2 rpm rpm2 1 d1

Using the data from Figure 2−19: Qb 3, 000 0.3 Qc 10, 000 Vb 1, 600 0.8 Vc 2, 000 Branch loss coefficient C = 0.69

2, 000

(0.5) 0.25 V 4, 005 2

2

p

TP C Vp 0.69 0.25 TP 0.1725in. wg Example 2.8 (SI) A main duct has an airflow of 5,000 L/s at 10 m/s. A 45 entry tap is used for the branch duct that requires 1500 L/s at 8 m/s. Find the pressure loss to the branch of the fitting. Solution Using the data from Figure 2−19: Qb 1500 0.3 Qc 5000 Vb 8 0.8 Vc 10 BranchlosscoefficientC 0.69 V p 0.602 (10) 2 60.2Pa TP C Vp 0.69 60.2Pa 41.5Pa

2.3.4.6

bp 2 rpm rpm2 1 bp 1

Solution

Fan Laws

The TAB technician can use fan curves or tables pub− lished by the fan manufacturer to determine the output of a fan under certain conditions. The following fan law equations also allow the TAB technician to calcu− late the necessary changes to be made to a system or a system component prior to the actual work.

Equation 2-20

P2 rpm rpm2 P1 1

2

Equation 2-21 3

Equation 2-22 2

Where: QPP=MAirflow C cfm(L/s) rpmP=Fan revolutions per minute P'P=Static or total pressure – in. wg (Pa) bpPM=Fan brake power – HP (W) dPP=Density C lb/ft3 (kg/m 3) The relationship between the fan laws, fan curves, and system curves will be discussed in Chapter 5 Fans. One of the most important things to remember is that the fan brake power increases as the cube of the fan rpm increase (or of the airflow increase by combining Equations 2−17 and 2−19). So when the TAB technician attempts to increase the airflow in a system without making other changes, the fan brake power (and the fan energy consumption) in− creases dramatically. 2.3.4.7

System Pressure

The total system pressure that the system fan must han− dle then is the sum of the friction losses of each straight duct section, the dynamic losses of each duct fitting or obstruction, and the pressure loss of each duct compo− nent such as coils, filters, dampers, etc. Examples can be found in Chapter 7 Air Systems. In a given duct system with a known airflow rate, and when the position of all dampers is stable, a specific total pressure can be measured. If the airflow is in− creased without any other changes, then Equation 2−21 can be used (this equation was derived from fan law Equations 2−17 and 2−18);

P2 Q2 P1 Q1

Equation 2-23 2

Where:


2.25

P Systempressure in.wg(Pa) Q SystemAirflow cfm(Ls)

Obviously, the 5 HP motor would be inadequate and a 10 HP motor would be marginal.

Example 2.9 (I−P Units) Example 2.10 (SI) A duct system is operating at 2.0 in.wg with an airflow of 10,000 cfm. If the airflow is increased to 13,000 cfm without any other change, determine the new duct sys− tem pressure.

Solution

000 QQ 2.0 13,

10, 000 2

P 2 P 1

The same system used in Example 2.9 has a 3.8 kilo− watt motor operating at 3.6 kilowatts. Find the fan brake power and standard motor size that would be re− quired if the airflow was increased to 6500 L/s.

2

2

Solution

1

2.0(1.3)2 3.38in.wg

7.91kilowatts 3.6 6500 5000

Q bp 2 bp1 2 Q1

Example 2.9 (SI) A duct system is operating at 500 Pa with an airflow of 5000 L/s. If the airflow is increased to 6500 L/s without any other change, determine the new duct sys− tem pressure.

845Pa 500 6500 5000

Q P 2 P 1 2 Q1

2

2

Example 2.10 (I−P Units) The same system used in Example 2.9 has a 5 HP mo− tor operating at 4.82 bhp. Find the bhp and standard motor size that would be required if the airflow was in− creased to 13,000 cfm. Solution

000 QQ 4.82 13,

10, 000 3

bp 2 bp1

2 1

4.82(1.3)3 10.59

2.26

3

3

3

The 3.8 kW motor is inadequate and a 7.5 kW motor would be marginal. 2.3.5 2.3.5.1

Hydronic System Basics Hydronic Pressure Losses

In air systems, the weight of air in the system is ignored by system designers and TAB technicians. In hydronic systems, it is the velocity head that is ignored because the values usually are insignificant. Otherwise, hy− dronic systems are subject to similar friction losses in the straight runs and dynamic losses in the fittings. Manufacturers normally supply pressure loss data for equipment used in piping systems. Pressure losses for hydronic systems are given in terms of equivalent feet of pipe, in pounds per square inch (psi), or in feet of water (ft wg) in I−P Units. In SI units, meters of water and Pascals or kilopascals are used.


Equation 2−22 is the hydronic equivalent of air systems Equation 2−21: Equation 2-24

P2 Q2 P1 Q1

2

towers, are subject to greater corrosion and therefore have higher pressure losses per 100 feet of pipe. Prop− erly sized pumps used with these systems generally will deliver a higher flow rate than required by a newly installed system because the anticipated corrosion has not occurred. However, after a few years, a partially closed balancing cock at the pump will have to be opened.

Where: 2.3.5.3 PP= Pressure difference C psi(Pa or kPa) QP= Flow rate C gpm(L/s)

Example 2−11 (I−P) A piping system has a flow rate of 100 gpm at a 20 ft wg head. Calculate the flow rate if the flow resistance is reduced to a 10 ft wg head.

Solution

2

P2 Q2 ; Q2 Q1 P1 Q1 Q 2 100

PP

10 71gpm 20

2

A piping system has a flow rate of 6.3 L/s at a 6 meter head. Calculate the flow rate if the flow resistance is reduced to a 3 meter head.

2

P2 Q2 ; Q2 Q1 P1 Q1 Q 2 6.3

PP

36 4.45Ls

Static head, discussed earlier, is the pressure due to the weight of the fluid above the point of measurement. In a closed system, the selection of the pump capacity is not affected as the static head is equal on both sides of the pump. However, the pump casing must be designed to handle the highest static head. Suction head is the height of fluid from the centerline of the pump on the suction side to the level of the fluid surface as is shown in Figure 2−20. The actual static head pressure loss that is added to the piping and sys− tem pressure loss in order to size the pump is the differ− ence between A minus B.

1

Example 2−11 (SI)

Solution

Heads

2 2

Suction lift is the height of fluid that a pump must lift on the suction side of the pump from the level of the fluid surface to the pump centerline as shown in Figure 2−20. This pressure loss value is added to any other sys− tem or pump pressure losses if additional piping or equipment is involved. 2.3.5.4

Pump Laws

The following equations are similar to (or the same as) the fan law equations. Again, they allow the TAB tech− nician to calculate changes that could occur in a given hydronic system when one or more of the conditions is altered. Most equations required for TAB work also can be found in the Appendix along with the SI equiva− lents. Equation 2-25

2.3.5.2

Friction Losses

Friction loss tables for hydronic systems vary in value depending on the condition of the piping system and the type of pipe or tubing used. Closed systems, where the fluid continuously recirculates, such as hot and chilled water systems for HVAC work, stay relatively clean and free from deposits that could roughen the in− terior surfaces. Open systems, such as domestic hot water systems and condenser water systems with normally open cooling

Q2 rpm rpm2 1 Q1 Equation 2-26 Q2 D 2 D1 Q1

H2 rpm rpm2 H1 1

H2 D2 H1 D1


Equation 2-27 2

Equation 2-28 2

2.27

STATIC PRESSURE

STATIC HEAD (DIFFERENCE = A - B) TO BE ADDED TO PUMP HEAD

A

SUCTION HEAD

B

CLOF PUMP TANK

PUMP

FIGURE 2-20 PUMP WITH STATIC HEAD AND SUCTION HEAD

bp 2 rpm rpm2 1 bp 1

bp 2 D2 D1 bp 1

2.28

Equation 2-29 3

Equation 2-30 3

Where: Q rpm D H bp

= Fluid flowCgpm (L/s) = Revolutions per minute = Impeller diameterCinches (mm) = HeadCfeet (meters) = Brake powerCHP (kW)


CLOF PUMP

PUMP

L SUCTION LIFT

TANK

FIGURE 2-21 PUMP WITH SUCTION LIFT


2.29


2.30


CHAPTER 3

ELECTRICAL EQUIPMENT AND CONTROLS

CHAPTER 3

ELECTRICAL EQUIPMENT AND CONTROLS

3.1

ELECTRICAL SYSTEMS

3.1.1

Basic Electricity

Equation 3-3 1R t 1R1 1R2 1R3 1R n (Parallel Circuits)

The electrical industry concerns itself with a broad range of subjects, many of which are not necessarily directly related to the HVAC industry. However, it is necessary to know basic electricity as it applies to that part of building construction which is related to me− chanical and electrical systems. Electric motors drive or power almost all HVAC equipment, including fans and pumps. Understanding the operation of, and the differences between, the many types of motors, the related controls, and the control circuits, is a necessity for the TAB technician. The TAB technician must be able to determine the brake horsepower that is being applied to air handling equipment and ensure that the motor is properly con− nected and protected. A few simple equations should be kept in mind when dealing with electricity. The first is a derivative of Ohm’s law: The current in a circuit is equal to the elec− tromotive−force activity in the circuit divided by the resistance in the circuit.

Where: R t TotalSystemResistance R n IndividualResistances

Equation 3−3 states mathematically that the parallel current flow will work similarly to hydronic flow, with the circuit with the highest resistance receiving the lowest flow. 3.1.2.2

Series Circuits

Resistances are added together for electrical circuits in series as in hydronic circuits. As more resistances are added, the flow becomes less and less. Equation 3-4 R t R1 R2 R3 R n (Series Circuits) Where: R t TotalResistance R n IndividualResistances

Equation 3-1 I E (or)E IR R P IE

Equation 3-2

Where: I Amps(A) E Volts(V) R Ohms(W) P Watts(W) 3.1.2 Electrical Resistances 3.1.2.1

Parallel Circuits

Parallel electrical circuits resemble HVAC terminal units piped in parallel circuits. Using a simple circuit with two units and one pump, it is known that the water flow will split in accordance with the resistance across each unit. If both resistances are the same, the flow will split 50−50.

The electrical diagram in Figure 3−1 is similar to a pip− ing circuit with a pump at ?E", two chillers in series at ?R1" and ?R2", seven terminal units piped in parallel, and a strainer, valve, etc., piped in series in the pump suction piping. This shows the similarity of electrical calculations to those for hydronic and air, where resist− ances in series are added, and those in parallel are com− bined. 3.2

ELECTRICAL SERVICES

3.2.1

Single Phase Circuit Voltages

A measured voltage may not be exactly one of the val− ues of voltages indicated in Figure 3−2. Voltages can vary, and in normal situations a variation of ±10 per− cent will not adversely affect equipment operation. The basic 115 volt two−wire circuit shown in part ?A" of Figure 3−2 is very common. There is a potential or pressure of 115 volts between the hot wire and the neu− tral or ground. Normal household use, such as a lamp, is representative of such a circuit. The 115 volt poten− tial in the hot wire will exist between the hot wire and the neutral, or between the hot wire and any other


3.1

Parallel

IT

R1

R2

(Etc.)

Series E Series

Series-Parallel

FIGURE 3-1 SERIES-P ARALLEL CIRCUIT

FUSE MAIN SWITCH

TO LOAD

115 V

GROUND

TO EQUIPMENT

GROUND LINE GROUNDS

A. TWO-WIRE CIRCUIT

GROUND LINE GROUND TO EQUIPMENT

MAIN FUSE HOT

MAIN SWITCH 115 V

115 V LOAD

115 V

115 V LOAD

NEUTRAL MAIN FUSE HOT CIRCUIT FUSES GROUND LINE

B. THREE-WIRE CIRCUIT

FIGURE 3-2 SINGLE-PHASE AC SERVICE

3.2


230 VOLT LOAD

+

Time

+

Time

+

Time

Current I

Voltage E

Voltage E

Voltage E

Current I

Current I

--

-Difference in Peaks Causing Power Loss Inductance Load Current Voltage

-Difference in Peaks Causing Power Loss Capacitance Load Current Voltage

Peaks Concurrent Providing Maximum Useful Power Inductance -- Capacitance Current Voltage (Power Factor Corrected)

FIGURE 3--3 CURRENT AND VOLTAGE-- TIME CURVES AND POWER FACTOR ground, such as a pipe or a person which might contact the wire.

The neutral or ground wire is another matter. The neutral normally has no voltage potential. Theoretically, if the neutral contacts a pipe or a person, nothing will happen. The neutral is connected to the ground. The term theoretically is used, because in actual field conditions, stray currents can find their way into the neutral and it then can be dangerous. A neutral should be treated with the same respect as a known hot wire.

Part “B” of Figure 3-2 shows another common single phase circuit. It is also a household circuit which serves items requiring greater power such as ranges, clothes driers and air conditioners. This circuit represents the type of three-wire service normally entering residences. Two of the three wires are hot wires and one is neutral. The voltage potential between either of the hot wires and the neutral is 115 volts. There are actually three circuits; two separate 115 volt circuits (from each of the hot wires to the neutral) and a 230 volt circuit (between the two hot wires). The neutral in a 230 volt circuit found in some appliance connections serves as a ground for safety, but it is not used as part of the power circuit. It is connected to the frame of the machine to carry off any stray currents or any short circuits resulting from failures. Ground or neutral wires are never switched or fused. The main advantage of the 230 volt, two hot wire circuit is that it allows each of the hot wires to carry half of the current flow. Therefore, twice the current will be handled by the same size wires.

3.2.2

Three Phase Circuit Voltages

The three-phase (3∞) concept is somewhat more difficult to understand. Figure 3-3 shows alternating current pulses of voltage and current changing with time. In the case of the single-phase, three-wire circuit, two different pulses are being sent down two different hot wires. After one starts, the second starts 1/120th of a second later. These pulses continue indefinitely at the same frequency and have the same phase relationship between the 2 wires. This can be thought of as +115 volts and -115 volts between the hot wires and the neutral wire. The pulses would average 115 volts between the hot wire and the neutral wire, would change to -115 volts and back to +115 volts and so on 120 times per second. The pulses going down the other wire would do the same, but because their timing is out of phase, an instantaneous look at the two 115 volt wires would show that the voltage in one wire would be +115 volts, and the other wire (because of its delay in phase) would be -115 volts. When voltage readings are taken with a voltmeter, there is no apparent way to tell the difference between 220 volt single phase circuits and 220 volt three-phase circuits. When measurements are taken, it is found that voltages do vary somewhat; that three-phase circuits are usually 220 volt, and that single-phase circuits are usually 230 volt. However, phasing cannot be determined by just voltage readings. Four-wire, three-phase circuits, as illustrated in Figure 3-6, produce 208 volts between phase wires. This arrangement is commonly used in commercial buildings, as the 120 volt loads can be divided equal between the three hot (phase) wires.


3.3

Three-Phase Motor (220 V)

L1 Main Switch Fuse

220 V

220 V

L2

220 V L3

FIGURE 3-4 220-VOL T THREE-WIRE DELTA THREE-PHASE CIRCUIT

Center Tap

Approx. 177 V

Fuse


L1 Transformer

Main Switch L2

110 V 220 V L2

Fuses

Single-Phase Loads (110 V) Neutral

FIGURE 3-5 220-VOL T DELTA THREE-PHASE CIRCUIT WITH 110-VOL T SINGLE-PHASE SUPPLY

Fuse


L1 Main Switch

208 V

208 V

L2

208 V L2

Fuses

Single-Phase Loads (120 V) Neutral

FIGURE 3-6 120/208-VOL T FOUR-WIRE WYE CIRCUIT

3.4


3.3

TRANSFORMERS

For the voltage reduction indicated for the transformer illustrated in Figure 3−7, the number of turns shown on the primary side should be twice the number of turns shown on the secondary side. Voltage is transformed by the transformer stepping down the voltage to one half the original voltage. By swapping the primary and secondary connections, this same transformer could step up the voltage from 440 volts to 880 volts. The ballasts in fluorescent lights in buildings step up the voltages from 115, 220, or 227 volts to voltages near 2500 volts, the required voltage to produce light in the tubes. The function of the center tap of a transformer also is illustrated in Figure 3−7. If a 220 volt difference exists between the legs of the secondary side, it is logical that a 110 volt difference would exist between one leg and a center tap. Most single−phase residential transform− ers have high voltages on the primary side, but the sec− ondary voltages use a ?center tap" (the ground) to fur− nish two 110 volt circuits along with the 220 volt power (Figure 3−2). This size transformer, which looks like a large can, usually is attached to a pole near the residence. It can supply power to several residences or buildings, or just to a single building. Larger transformers are ground mounted, and if out− side, are usually in a metal housing. These transform− ers can be either single−phase or three−phase. The ?tap" (or neutral) cannot be used with hot line ?L2" for 110 volt single−phase loads (Figure 3−5) from a three− wire, 220 volt circuit. However, all three hot line or phase wires may be used for 120 volt loads in four− wire, 208 volt circuits (Figure 3−6).

110 V Secondary 440 V Primary

(Center Tap) 220 V Secondary 110 V Secondary

FIGURE 3-7 TRANSFORMER WITH TAPPED SECONDARY 3.4

MOTORS

3.4.1

Types of Motors

Most motors used on HVAC system equipment are de− signed for alternating current. Small motors will use single−phase current, while the larger motors will use three−phase current. However, some rural areas must use larger motors on single−phase current, as three−

phase current is not available. There are many differ− ent motor speeds, but 1800 rpm and 3600 rpm are the most common. The actual speed of the motor will vary with the load imposed. Split phase, capacitor start, synchronous, induction, shaded poleCall are part of the many different types of motors that the TAB team will need to recognize. The characteristics of each is important for troubleshooting, as the wrong type of motor may be used. 3.4.2

Rotation

Motors rotating in the wrong direction are a common occurrence when a new system is started. The normal TAB procedures deal with this situation, as correct motor rotation is vital to the performance of the unit. The direction of the motor usually is changed in three− phase motors by switching any two of the three−phase power wires. In single−phase motors, the change of di− rection is accomplished by switching two of the inter− nal motor leads that connect to the motor line terminal lugs. CAUTION: Certain fans and most pumps will develop measurable pressures and some fluid flow when the rotation is incorrect. Rotation arrows can be found on many types of equipment. Correct rotation is obvious on some units. Flow and amperage readings also can be used to determine whether something is amiss. Whenever a piece of equipment does not perform as specified and the current flow is much lower than de− sign, rotation is one item to be checked. 3.4.3

Nameplate Data

Except for some small motors, an attached nameplate will supply the basic information that the TAB techni− cians needs: full load amps, rpm, horsepower or watts capacity, voltages, line phase, and cycles. Many motor nameplates contain starting load amps that are quite large compared to full load amps. Although starting load amps are not as important (and are not recorded), the amperage value must be used by the system design− er for electrical circuits, circuit breaker panels, and motor starting equipment. Information on special motors might have to be ob− tained from specification sheets or from the HVAC unit nameplate. Since voltage and amperage measure− ments are seldom the same as the nameplate values, the actual motor horsepower being produced can be es− timated. However, the no load amps reading is difficult to obtain from close−coupled equipment where the load cannot easily be detached from the motor. The motor must be running with all drives disconnected for this no load amps reading.


3.5

3.4.4

Operating Temperatures

Every electric motor generates heat as well as power. The more inefficient the motor, the greater the amount of heat produced. Unless this heat is dissipated, the temperature within the motor will rise until the insula− tion is destroyed. The amount of heat produced in the motor also de− pends upon the load. Therefore, most motors are rated on the basis of a certain temperature rise when running fully loaded. Open motors are generally designed to run at a temperature rise of 72F (40C) above the sur− rounding (ambient) air temperature. In other words, an open motor when running at full nameplate condi− tions, could be running at a temperature of 142F (61C) if the surrounding air is 70F (21C). Totally enclosed motors generally are rated at a 99F (55C) temperature rise. Under the latter conditions, a totally enclosed motor would run at 169F (76C). 3.4.5

Motor Performance

Figure 3−8 indicates ways in which various motor fac− tors are interrelated. The point at which the speed and the amperes cross corresponds to 55 percent of the maximum amps and over 60 percent of the maximum horsepower. At this point, the speed is between 97 and 98 percent of the maximum synchronous speed (which is not a great change) and the efficiency curve stays fairly flat close to 90 percent. The power factor also stays between 80 and 90 percent. Single−Phase Circuits: Equation 3-5 (I-P) I E P.F. Eff. bhp 746 Equation 3-5 (SI) I E P.F. Eff. kW 1000 Three−Phase Circuits: Equation 3-6 (I-P) I E P.F. Eff. 1.73 bhp 746

3.6

Equation 3-6 (SI) kW

I E P.F. Eff. 1.73 1000

Where: bhp'= Brake horsepower (I−P) kW' = Kilowatts (brake power) (SI) I' = Amps E' = Volts P.F.'= Power factor Eff.'= Efficiency In Equation 3−5 and 3−6, the power factor and efficien− cy values must be used to obtain the actual motor brake power. As these values usually are difficult to obtain, a reasonable estimate can be used. Referring to Figure 3−8, the normal range of both curves is between 80 and 90 percent. Therefore, 80 percent might be used for one value and 90 percent for the other value to obtain a brake power estimate. Brake power is calculated to verify that the proper size motor has been installed, i.e., that the installed motor is not overloaded and is operating within its service factor. It also is used to determine that the pump or fan is operating with the required efficiencies. The system designer usually has specified the total amount of pow− er or energy that may be consumed to perform a specif− ic function. For example, a pump is selected to circulate 120 gpm (7.6 L/s) of water at a 40 foot (12 m) head and consume not more than two horsepower (1.5 kW). Designers in the past may have used a three horsepower (2.3 kW) motor on the pump in order to have a safety factor, or to have extra capacity for future loads that may be planned for the system. With energy conservation con− siderations, the installation of a three horsepower (2.3 kW) motor just to have a safety factor should not be used unless a future load is planned. When the designer does not want to use more than the rated two horse− power (1.5 kW), it is essential to calculate the actual power consumption unless the amperage reading is well within the full load amperage rating of the motor. There are two equations that must give a less theoreti− cal, but more practical approach to the calculation of motor full load (F.L.) amperage and brake horsepower (kW).


Equation 3-7 ActualF.L.amps

Solution

F.L.amps * voltage * Actualvoltage

a.

Using Equation 3−7: 8.16 220V 210V 8.55amps

ActualF.L.amps

Equation 3-8 bhp(kW) HP * (kW *) (Motoroperatingamps) (Noloadamps 0.5)

b.

Using Equation 3−8:

(ActualF.L.amps) (Noloadamps 0.5) bph (kW)'= 3 HP (2.3 kW)

Use Equations 3−7 and 3−8 to obtain an accurate (but not exact) brake power by measuring motor amper− ages and voltages under no load and full load condi− tions.

(6.2) (4.7 0.5) (8.55) (4.7 0.5)

bph(kW) 3HP(2.3kW) *Nameplate ratings that supply the basic information.

(6.2) (2.35) (8.55) 2.35)

bph(kW) approx.1.86HP(1.43kW)

Example 3.1 Example 3.2 A fan has a 3 HP (2.3 kW), 220 volt, 3 phase motor that actually draws 6.2 amps at 210 volts. The full load am− perage shown on the nameplate is 8.16 amps and the ?no load" measurement is 4.7 amps. Determine the approximate fan brake power.

The following loads are paralleled across a 220 volt, single−phase, 60 Hz (hertz) source: a.

A 10,000 watt electric heater

90 100

80

Efficiency 90

% Power Factor

100

70 100 100

60

99

50

98

40

97

30 96 20

75

50

25

% Current (Amperes)

70

% Synchronous Speed

% Efficiency

80

95

10

0 0

10

20

30

40

50

60

70

80

90

100

% Motor Horsepower

FIGURE 3-8 TYPICAL PERFORMANCE OF STANDARD SQUIRREL CAGE INDUCTION MOTORS HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

3.7

b.

A 5 horsepower (3.8 kW) electric motor oper− ating at a P.F. of one, with a 90 percent effi− ciency and pulling full load

The reason for the variance in answers between the I−P and SI examples is that the 3.8 kW motor size is a nom− inal size, but 3.73 kW is the exact equivalent.

c.

A 220 to 110 volt transformer (assume 100 percent efficient) with fifty 100 watt incan− descent light bulbs in parallel with the secon− dary.

3.5

MOTOR CONTROLS

3.5.1

Safety Switches

Heater

Transformer 110 V

220 V

10,000 Watt

1&

Motor 5HP 50 @ 100 watts each

Determine the total current requirements on the 220 volt source. Solution a.

b.

P

= EI, I = P/E = 10,000 W/220 V

I

= 45.5 amps

Using Equation 3−5 (I−P):

I E P.F. Eff. 746 P P.F. Eff. bhp 746 bhp 746 5 746 4144watts P P.F. Eff. 1 0.9 bhp

Using Equation 3 − 5 (SI): I E P.F. Eff. 1000 P P.F. Eff. kW 1000 kW 1000 3.8 1000 4222watts P P.F. Eff. 1 0.9 I P 4144 18.8amps) E 220 I 4222 220 = 18.8 amps (I−P) kW

c.

For transformers with 100 percent efficiency, P(in) = P(out) P 50 100watts 5000watts I P 5000 22.7amps E 220

I (Total) = 45.5 + 18.8 + 22.7 = 87.0 amps (I−P) I (Total) = 45.5 + 19.2 + 22.7 = 87.4 amps (SI) 3.8

A simple on−off toggle switch, a safety switch, or an individual circuit breaker in an electrical power panel is not an overload protection device for a motor. Many ordinary looking toggle switches do contain overload protection for smaller single−phase motors. Many small motors do have built−in overload protec− tion, and do not need additional protection. The circuit breaker only provides overload protection for the wir− ing circuit, but not any connected motor(s). The electric current to a motor must be switched off and on to stop and start the motor (manually or auto− matically). The switching device is commonly called a motor starter. This is not to be confused with a safety switch, which is a device that must be placed in the off position before any work is done on a motor or electri− cal equipment. This prevents the motor from acciden− tally starting from remote control devices. 3.5.2

Motor Starters

There are a large number of different types of starters, each with various advantages and limitations. In most cases a specific type of starter is required by a particu− lar type of motor. For example, a full voltage magnetic starter usually is used with an induction motor. Re− duced voltage or reduced current starters, while more expensive than a magnetic starter, often must be used with larger horsepower motors to prevent disruption (by producing large drops in line voltage) of marginal− ly adequate power services. Many electrical utility companies have mandatory requirements for these starters above a certain horsepower (this varies with the type of equipment and voltage). The motor starter or the safety switch is the main source of access to motor terminal leads for measure− ment of voltage and amperage. The starter also can contain holding coils, auxiliary contacts, control trans− formers, and a push−button station or a hand−off−auto− matic selector switch. This last item is useful in trou− bleshooting. If the switch is turned to the hand position, the motor should run if each phase line is hot, unless there is trouble in the motor. This is because the hand portion of the switch bypasses the various con− trols in the circuit. The automatic portion of the switch is connected to the circuit containing auxiliary devices


such as thermostats, safety lockouts and other external switches used to control or turn off the motor. If the motor runs on hand, but not on automatic, contacts in one of the control or safety interlocks should be open.

The TAB technician can find many different voltages around starters and starter combinations (see Figure 3−9). If a remote room thermostat is added in series with the push button station of the ?A" unit, the 110 volt control circuit then may be required to be 24 volts.

It is not good practice to use line voltages (110 volt or higher) for control circuits, but to reduce costs, 240 volt control circuits are not uncommon.

The motor starter overload protection devices or heat− er coils should be sized from the actual motor name− plate data rather than from data from motor or starter manufacturer’s catalogs. Heater coils never should be oversized under any condition.

3.5.3

Push Button Stations

There are two basic types of push button stationsCthe maintained contact station and the momentary contact station. The important point to remember is that after an interruption of the current with the momentary con− tact station, the motor will not restart until the start but− ton is pushed. 3.6

VARIABLE FREQUENCY DRIVES

Variable speed or variable frequency drives (VFD) are becoming commonplace in new commercial and insti− tutional construction. On renovation projects, this electrical component is ?spliced" into the existing wir− ing between the electrical source and the fan or pump motor disconnect. Figure 3–10 shows an existing H & V unit that was retrofitted with a VFD wired into the motor disconnect. Drive manufacturers achieve variable motor speed op− eration by modulating the frequency of the electrical power being supplied to the motor during normal op− erations, as well as voltage adjustment during startup

440/3/60

MAINTAINED CONTACT PUSH BUTTON STATION

220/3/60 MAIN CONTACTS

L1

L2

L3

L1

L2

L3

T1

T2

T3

STOP AUXILIARY CONTACT

START T1

T2

HOLDING COILS

T3

HEATER COILS

110 V

440/110 V CONTROL TRANSFORMER

M A. CONTROLLING STARTER-MOT OR

110 V 220/110 V CONTROL TRANSFORMER

M B. CONTROLLED STARTER-MOT OR

FIGURE 3-9 INTERLOCKED STARTERS WITH CONTROL TRANSFORMERS HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

3.9

to approximately 40% of their full design speed with no noticeable effect on motor performance or life. However, unless specifically designed for variable fre− quency operation, most motors will develop a notice− able increase in noise level and temperature below this point. Continuous operation below 25 percent of full load speed is not recommended and this lower range should only be used to reduce high current surges and belt stress during fan startup or shutdown. Just as lowering the frequency below 60 cycles per sec− ond will lower the motor speed below nameplate RPM, you can also increase motor speed above name− plate rating by increasing the frequency above 60 cycles per second. All initial air balancing should be carried out with the VFD set for approximately 55 cycles per second (Hz), even if drive pulleys and belts must be changed to achieve the design peak air flow. This will provide added ?headroom" for fan capacity when the system experiences duct and coil static pres− sure loses from dirt buildup. 3.6.1 FIGURE 3-10 VFD ADDED TO EXISTING AIR HANDLING UNIT and shutdown. Although most 3−phase motors will op− erate satisfactorily on a VFD, low cost motors and some energy efficient motors may experience higher noise or heat levels, especially at very low speed op− eration. Many manufacturers now offer modified mo− tor designs specifically for VFD operation. Varying the speed of an AC motor is much more com− plex than DC motor speed control. Most traditional DC motors have a commutator and brushes to provide electrical power to the rotating armature coil. Varying the voltage to this coil using a resistor or a rheostat changes the motor’s speed fairly effectively. AC motors do not have a commutator or brushes since the constantly alternating electricity induces an oppos− ing electrical field in the armature coil much like the primary coil of a transformer inducing a voltage in the secondary coil. Each AC motor is designed for a spe− cific voltage and reducing the supply voltage below this value will cause the motor to quickly lose its load capacity and stall, causing overheating and eventual motor burnout. Keeping the supply voltage to an AC motor at its de− sign point while varying the frequency of this voltage results in a corresponding change of motor speed. Most AC motors can be operated continuously down 3.10

VFD Operation During TAB Work

Most VFD devices can be programmed with a maxi− mum and minimum motor speed, and ?ramp up" and ?ramp down" rates to reduce wear on drive belts and bearings. The days of a TAB technician replacing mo− tor pulleys to adjust fan speed is drawing to a close, and today’s TAB technician needs to become familiar with variable frequency motor drives and their setpoint pro− gramming. In most cases, drive and motor manufacturers do not recommend operating these systems for extended peri− ods of time below 40 percent motor speed, and all VFD programmed setpoints should be verified and recorded during the balancing process. 3.6.2

VFD Bypass

Almost all VFD devices include a hand−off−auto switch and a manual speed control dial that can be used in manual operation. If the VFD device is placed in manual mode during TAB work, be sure it is returned to the auto mode before completing this work. Also note that in manual mode it is possible to operate the fan or pump at full speed. Since a remote duct or piping pressure sensor may be connected to the VFD device to allow maintaining a fixed duct or water sup− ply pressure, manually operating the fan or pump while down stream dampers or valves may be closed could cause very high pressures to develop in ducts or piping.


CHAPTER 4

TEMPERATURE CONTROL

CHAPTER 4 4.1

AUTOMATIC TEMPERATURE CONTROL SYSTEMS

4.1.1

Introduction

Automatic temperature control of HVAC includes the control of temperature, humidity, and sometimes sys− tem or building pressures. The automatic temperature control (ATC) system constantly adjusts the HVAC systems to maintain design conditions within the occu− pied space. Because TAB work is concerned with the operation of the HVAC system, and because the control system constantly adjusts the HVAC system, it is imperative that the TAB technician understands the function and use of automatic temperature control systems. It must be remembered that motor controls covered in Chapter 3, are not automatic temperature controls, but the automatic temperature control system may moni− tor or operate motors and motor controllers. The use of the word control for both systems and devices some− times causes this basic difference to be overlooked re− lated to responsibility. 4.1.2

Types of ATC Systems

There are four basic types of controls for HVAC or en− vironmental systems:

electric pneumatic electronic self−contained

There are also combinations of the above types. 4.1.2.1

Electric Controls

Electric controls are those which are line voltage or less (generally 110 volts maximum). Reduced volt− ages are obtained from transformers, either locally or centrally situated. Many of these controls are simply on−off devices such as a high limit thermostat control− ling an exhaust fan. Generally, complex systems re− quire more sophistication than these controls can pro− duce. 4.1.2.2

Pneumatic Controls

Until the advent of micro−electronics, all HVAC con− trols of air handling systems, chillers, and boilers were pneumatic controls. These systems used compressed air to operate diaphragms and mechanical relays to

TEMPERATURE CONTROL position dampers and valves. Although fairly easy to visually observe the operation of each control device, changing the sequence of operation for any HVAC sys− tem was in many cases a plumbing ?nightmare," since everything was interconnected by air tubes. At the close of the 1990’s, most pneumatic logic controls have been replaced by direct digital controls (DDC), with the exception of very large dampers and valves which may still utilize pneumatic damper motors as large electronic motors are still relatively expensive. On most of today’s commercial and institutional pro− jects, the TAB technician may find that all HVAC con− trols are now based on micro−electronics. However, we are still including a review of pneumatic control basics in this chapter as these systems still exist and may be encountered during HVAC system renovation and expansion. 4.1.2.3

Electronic Controls

The term ?electronic controls" was first used to de− scribe newer control technology being installed to re− place some of the functions of the older pneumatic control devices. Most of these first generation elec− tronic controls did little more than monitor HVAC sys− tem operation and provide on/off control of fans, valves, and dampers. The control ?logic" was special− ized software operating on a large main frame central computer, communicating with field interface panels connected to temperature sensors and relays. Any pneumatic controlled dampers or valves were still op− erated by their original pneumatic controls, with E/P relays switched between fixed setpoints. These earlier electronic control systems were of little interest to the TAB technician other than requesting an unseen con− trol operator to start or stop a fan. During the 1990’s this older technology was replaced by DDC which no longer required the pneumatic con− trol devices to carry out the positioning of dampers, valves, and setpoint controllers. In addition, with the advent of micro−electronics, the operating program software is now contained within the remote field pan− els and the large central computer is no longer re− quired. Unlike the earlier electronic control systems, this new DDC technology does have a significant impact on the balancing contractor on all but the smallest projects. HVAC system manufacturers are finding that it is less expensive to use these easily programmed ?black box" control devices. There are no longer pneumatic control devices that must be constantly calibrated and ad− justed to maintain system reliability. The days of a TAB technician adjusting a pneumatic controller to reposition the damper on a VAV box may


4.1

be coming to an end, and today’s TAB technician may find a hand held computer indispensable on a balanc− ing job. What may have required above ceiling linkage adjust− ments while standing on a shaky ladder, can now be ac− complished by plugging a hand held device into the nearest electronic wall thermostat. The TAB techni− cian can instantly verify high and low air flows, adjust operating setpoints, and monitor room conditions as the HVAC system responds to these control input changes. It is very important for any SMACNA contractor en− tering today’s TAB field to have a good understanding of DDC basics. 4.1.2.4

Self-Contained Controls

Self−contained controls differ from the above types in that they do not use an external source of power, but develop their own power. Often used in automatic valves, a bellows or other sensing element has enough strength to move the valve. Because of strength and large mass involved in its construction, it is not capa− ble of providing as close control as other types of sys− tems. Applications include:

condenser water regulating valves on refrig− eration compressor units (city water). thermostatic expansion valves. steam control valves for heating domestic hot water. self−contained radiator control valves.

Other combinations are electro−hydraulic, commonly applied to valve operators, and electro−pneumatic sys− tems using electronic devices to sense temperature and pressure, and pneumatic devices to operate valves and dampers. This dual system combines advantages of electronic systems (sensitivity, wide range of adjust− ability) with simplicity of pneumatic operators. 4.1.3

Control Categories

Control systems also are divided into two categories: operating controls and safety or limit controls. 4.1.3.1

Operating Controls

Operating controls are used for the control of room conditions and system setpoints. The most common example of an operating control is a room thermostat. 4.2

4.1.3.2

Safety Controls

Safety or limit controls are used to provide safe equip− ment operation. Safety or limit controls must be set properly to avoid unsafe conditions such as pressures or temperatures that are too high or too low, and imple− ment emergency equipment shut off. Safety or limit controls may interrupt the operating controls at any given time to ensure safe system opera− tion. Examples of safety controls are freeze stats, fire stats, flow switches, smoke detectors, and refrigera− tion high−low pressure cutouts. 4.2

CONTROL LOOPS

No matter which type of control system is used, all control applications must involve a fundamental con− trol loop. A control loop consists of three components:

a controller (thermostat) a controlled device (valve, damper) a sensing device (transmitter, bi−metal strip)

For example, a sensing device (remote bulb) monitors the temperature of a supply air duct and sends a signal to the controller. The controller monitors the signal as sent by the sens− ing device, and reacts by either opening or closing a controlled device (valve or damper). As a result of the resulting change in system output, (such as hot water in a heating coil), the action of the controlled device creates a change in the sensing device which provide feedback to the controller that the system change took place. The operation of any control loop is continuous during normal operation of the HVAC system. 4.2.1

Controllers

Controllers (such as thermostats and humidistats) have two possible sets of control actions:

4.2.1.1

Modulating or two position Direct or reverse acting Modulating/Two Position

Modulating control (also called proportional control) is obtained when the control signal sent by the control− ler to the controlled device is constantly changing in small increments to gradually increase or decrease the capacity of a system component to suit the load condi− tions. Two position control, which also can be on−off,


only assumes two positions; fully open or fully closed. Two position controllers are normally used with equip− ment that only operates in an on−off application. Equipment such as gas valves or small air conditioning compressors would fall under this category. 4.2.1.2

Two−way valves are normally used for water or steam service. Three−way mixing and diverting valves are used only in hydronic piping (Figure 4−2). The valve constant or flow coefficient is used to calculate the flow and pressure drop of ATC valves in the wide open position.

Direct/Reverse Acting Equation 4-1 (I-P)

Modulating system control devices may increase or decrease the output (branch) control signal with changes in space conditions monitored by the sensing element. A direct acting controller will increase the output (branch) control signal as the controlled vari− able (temperature, humidity, pressure) increases. A re− verse acting controller increases the control signal as the controlled variable decreases. The action of the controller must be properly matched with the control device or the control loop will produce unexpected re− sults or those opposite of that desired.

DP

CQ

2

v

Where: DP Pressuredifferential(psi) Q Flowthroughvalve(gpm) C v FlowCoefficient

Equation 4-1 (SI)

DP

KQ

2

v

The position of a controlled device when de−energized is considered the normal position. Control devices such as valves or dampers are either normally open (N.O.) or normally closed (N.C.). Some electric de− vices also contain switches that are normally open or normally closed until moved to the opposite position by a controller.

Where: DP Pressuredifferential(kPa) Q Flowthroughvalve(Ls) K v FlowCoefficient

4.2.2

A control valve must be selected to control a flow of 20 gpm at a maximum 4 psi pressure drop. Calculate the Cv of the valve.

Controlled Devices

Controlled devices that affect the TAB technician the most are automatic control dampers and automatic control valves. Both affect flow and both can be two position or modulating.

Example 4−1 (I−P)

Solution 4.2.2.1

CQ ; DP CQ 2

ATC Valves

DP

v

Figure 4−1 illustrates the throttling characteristics of the different types of modulating ATC valves. Two position valves such as those used for automatic shut− off in seasonal change−over piping need no specific throttling characteristic, the major concern being tight shutoff. Gate

Butterfly

100%

v

Q C v 20 10 DP 4 Example 4.1 (SI) A control valve must be selected to control a flow of 1.3 L/s at a maximum 28 kPa pressure drop. Calculate the Kv of the valve.

Ideal Straight Line Characteristic Throttling Plug Travel 50%

Globe

Solution

KQ ; DP KQ 2

50%

100%

DP

% Flow

FIGURE 4-1 VALVE THROTTLING CHARACTERISTIC COMPARISON

K v

v

v

Q 1.3 0.25 D P 28


4.3

IN

OUT

OUT

IN

OUT

IN 3 Way Mixing

3 Way Diverting

FIGURE 4-2 ATC VALVE ARRANGEMENTS For proper control action, it is desirable for an ATC valve to be sized so the pressure drop cross the wide open valve at design flow rate will give an appreciable pressure drop. For example, in the case of a steam valve, it is considered good practice for the pressure drop at design flow to be approximately 50 percent of the absolute steam pressure available at the valve. 4.2.2.2

closed, produces a control characteristic that is unsat− isfactory. Opposed blade dampers do not eliminate the above problems, but they improve the control ability by clos− ing blades toward each other so that throttling begins sooner. Close and accurate control is improved but still limited. The linear operating characteristic is not as

ATC Dampers

Dampers used for automatic temperature control have either parallel or opposed blades as shown in Figure 4−3. Quality, tight fitting dampers with long lasting blade edge seals or the equivalent are necessary for ATC work. Parallel blade dampers are almost always used for two position or open−closed control. Opposed blade damp− ers are used for modulating control of airflow. Damp− ers present throttling problems similar to valves which is difficult to correct. Parallel blade dampers often have a throttling characteristic which is worse than gate valves. This deficiency is complicated by the pro− cedure of selecting damper sizes based on low air ve− locities across the dampers. The play in the dampers and damper linkages caused by flexibility and distor− tion, and the difficulty of seating the blades when 4.4

PARALLEL OPERATION

OPPOSED OPERATION

FIGURE 4-3 TYPICAL MULTIBLADE DAMPERS


nearly achieved as in the case of valves, but may be more closely approximated by sizing the dampers on higher velocities and by providing more sections and more rigidly constructed.

the sequence of control and the ATC diagrams would indicate which devices to inspect.

4.2.2.3

Most control system sensors and controllers are linear (which means straight line). Figure 4−1 indicates the error induced by non−linear control devices such as dampers and valves. Linear control in pneumatic sys− tems may translate to one degree of temperature change from one psi of air pressure change. One psi (kPa) or fluid pressure change on the discharge side of a pump also could result in a two or three psi (kPa) of control pressure change. These systems are linear as long as each increment of controlled variable produces the same increment of signal. The system would be non−linear if different amounts of signals emanate from a fixed increment of the controlled variable. For example, a system is non−linear if at 70F (21C), a one degree change produces a one psi (kPa) control signal; but at 90F (32C), a one degree change pro− duces a two psi (kPa) control signal.

Valve and Damper Linkages

The operation of automatic valves can often be re− versed in the field to suit the action of the controllers, although in some cases it may be necessary to change the operator. The action of automatic dampers can usu− ally be reversed by resetting the damper arm or other parts of the linkage. Sometimes it is necessary to limit the travel of dampers or valves in order to provide proper control. Most pneumatic and electric devices have operator stops that can be adjusted so that the valve or damper operator is permitted to complete a portion of its stroke. Electric operators have limit switches which can be positioned to electrically stop the operator at a desired position. The correct setting of stops and/or limit switches is important for the suc− cessful operation of a system, and personnel must un− derstand such adjustments, or equipment could be damaged. Although adjustments of ATC valves and dampers on larger systems are normally made by the Temperature Control Contractor, the TAB technician needs to un− derstand the factors required for proper valve and damper adjustment. 4.3

4.4

CONTROL RELATIONSHIPS

An actuator is considered linear if it has a signal range of ten psi (kPa) from fully open to fully closed. So a five volt or five psi (kPa) signal will cause a 50 percent travel. However, if the actuator device is used on a valve or damper, an actuator change of 50 percent will seldom change the fluid flow by the same 50 percent. From Figure 4−1, one can see that a 50 percent stem travel of a gate valve from wide open will have little effect on the fluid flow.

CONTROL DIAGRAMS Equation 4-2 (I-P)

In a typical job specification, there are general descrip− tions of various types of control applications , called the sequence of controls, which the automatic Temper− ature Control Contractor must translate into a set of drawings called control diagrams. These control dia− grams and related written sequence of controls for each HVAC system, are used by the ATC contractor for control system installation in coordination with the HVAC system contractor. The data found in these dia− grams is extremely important to the TAB technician and these diagrams frequently are the only description of how a complicated HVAC system will operate. Control system diagrams also can be used to assist the TAB technician in troubleshooting. For example, when the hand−off−automatic switch of a fan motor starter is in the automatic position, it is found that the fan will not run. But the fan will run when the switch is in the hand position. This indicates that some type of automatic temperature control device or safety switch is preventing the fan from running. A review of

C v Q

(2.3)½ Q (H)½

2.3 H

Equation 4-2 (SI) Q K v DP

Equation 4-3 Q21 H1 DP 1 2 H2 Q2 DP 2 Where: Cv Kv Q P H

= = = = =

Flow coefficient or valve constant (I−P) Flow coefficient or valve constant (SI) Fluid flow rateCgpm (L/s) Pressure differenceCpsi (kPa) Head loss or pressure dropCfeet (meters)


4.5

Equations 4−2 and 4−3 indicate that the pressure drop across the valve is proportional to the square of the fluid flow rate. This relationship is indicated by the general curve shown in Figure 4−1, which can apply to most systems, although the numbers may vary. The non−linearity of the controlling device is apparent with this curve. In order to minimize the resulting control inaccuracies, the controller and the controlled device must be carefully matched so that an average linearity is achieved. This cannot be done across the entire range of the device, therefore, the devices are matched for a normal operating range, which is a matter of judg− ment of the system designer or the ATC Contractor.

trol system is required. Outside air and exhaust air dampers usually are interlocked with the supply fan to open to a fixed minimum outside air position when the fan is started. A mixed air temperature sensor could then control the outside air, return air, and exhaust air dampers to maintain a set mixed air temperature. At a pre−set temperature or a high outside air humidity, the outside air damper often will be returned to a mini− mum position to decrease the cooling load of the out− side air. In a case of power or control system failure, the outside air damper usually closes automatically. A freeze stat also can stop the fan and close the outside air dampers.

4.5

4.6.2

ATC SYSTEM ADJUSTMENT

After completion of the physical installation, the ATC system components must be adjusted and calibrated so that they may operate individually and collectively to provide the specified environmental system control. The amount of adjustment and calibration will depend on the complexities of the ATC system. All calibration of ATC system instruments should have been done by the ATC installer prior to system balan− cing. However, there are some specific adjustments which should be done in conjunction with TAB per− sonnel during system adjustment and balancing. Fail− ure to provide this coordination may lead to the inabil− ity of the HVAC system to perform satisfactorily under load. Setting automatic dampers for proper air quantities, positioning hot and cold deck dampers, and maintain− ing valves open or closed to maintain design operating conditions are among the multitude of factors which affect systems operation and TAB work. After the installation has been completed, accepted, and the building occupied, problems can arise which may or may not be attributable to the TAB work. It is not unusual for accidental maladjustment of controls to produce symptoms which seem to point to improper HVAC system balancing. Outside air dampers that have slipped, or a reheat coil thermostat that malfunc− tions, are excellent examples. The ability to recognize the real source of the problem not only saves time but vindicates the TAB work. 4.6

TAB/ATC RELATIONSHIP

4.6.1

Related Problems

To properly balance and adjust any HVAC system, a thorough knowledge of the installed temperature con− 4.6

Controllers

A thermostat in the duct system often will control a heating or cooling coil valve, face and by−pass damp− ers or mixing dampers. A room thermostat can control a hot water, steam, chilled water or electric booster coil, and/or hot and cold mixing dampers. A humidis− tat can control a humidifier or a cooling coil for dehu− midification. Controls can be direct acting, reverse acting, modulating or two position, stepped, master, sub−master, series, or parallel. Controls can actuate dampers, valves, and relays; start, modulate or stop motors, fans, and other equipment; and be controlled by time clocks, time delay relays, static pressure con− trollers, air switches, flow switches, level controllers, fire and smoke detectors. They can be connected to alarm systems, and be controlled, readjusted, and be indicated from or at remote control panels. Finally, controls can be very simple or very complex. 4.6.3

Ventilation Air

Probably the most important effect on TAB work is the setting of the outside air, return air, and exhaust air dampers. After the fan speed (rpm) and airflow capac− ity have been checked out, set the outside air dampers for minimum outside air. Use thermometers or a ther− mocouple to measure outside air, return air and mixed air temperatures. Use the mixed air temperature Equa− tion 4−4 to determine the amount of outside air. Work− ing with the temperature control contractor, set the minimum outside air conditions. Mark the dampers for the minimum position and recheck the air flows. If an economizer cycle is used, next check using 100 percent outside air and again check air flows. Follow the procedure for 25, 50, 75 percent outside air. The mixing of the return air with outside air to give a known mixed air temperature can be determined by the following:


X oT o X rT r T m 100

Equation 4-4

Tm = Temperature of the mixture of return air− and outdoor air Xo = Percentage of outdoor air Xr = Percentage of return air To = Temperature of outdoor air, F (C) Tr = Temperature of return air, F (C) Being familiar with the interactions and functions of these control systems will go a long way in reducing on site system balancing time. The following equations are used for determining per− centages of outside air. For this work, more convenient forms of expressing Equation 4−4 are given in Equa− tion 4−5 and 4−6. (Tr Tm) X o 100 (T r T o)

(Tm To) X r 100 (Tr To)

Equation 4-5

Example 4.2 (SI) 24C return air is mixed with −4 outside air and the mixed air temperature is 13C. Find the percentage of outside air.

Solution (Tr Tm) (T r T o) (24°C 13°C) 100 [24°C ( 4°C)] 100 11°C 39.3% 28°C X o 100

4.7

CENTRALIZED CONTROL SYSTEMS

The concept of centralized controls or energy manage− ment systems (EMCS) briefly addressed in the begin− ning of this Chapter is applied to many buildings being constructed or modernized today.

Equation 4-6

Example 4−2 (I−P) 75F return air is mixed with 25F outside air and the mixed air temperature is 55F. Find the percentage of outside air.

Solution (Tr Tm) (T r T o) (75° 55°) 100 (75° 25°) 100 20° 40% 50° X o 100

FIGURE 4-4 DESKTOP COMPUTER DISPLAYING STATUS OF BUILDING HVAC SYSTEMS


4.7

Figure 4−4 shows a standard desktop computer being used to monitor and control all HVAC systems in a hos− pital. These software programs are becoming very easy to use and can display photos and diagrams of a given system, with all ?live" temperature and setpoint data displayed next to each system component. Almost all new buildings will utilize a computerized HVAC control system. Even the most basic dial time clock for start/stop control have been replaced by less expensive electronic time clocks that can adjust for seasonal length of daylight and changing outdoor tem− peratures. Large HVAC systems now have one or more field control panels that include programmable com− puterized memories containing all of the sequence of control logics and control algorithms for the systems and devices being controlled. These field panels include four types of control inputs and outputs, plus the ability to communicate local con− ditions and setpoints to other field panels or remote monitoring locations.

Digital Input Digital Output Analog Input Analog Output

4.7.1

4.7.2

In addition to these input and output field devices, a typical centralized computer control system consists of one or more field interface devices, and one or more programmable stand alone controllers as shown in Fig− ure 4−5. Each programmable stand alone controller contains a microcomputer and battery backed up memory con− taining all of the programming for all field devices and

SENSORS

Many field interface panels include a manual/auto switch for each control output. This allows local by− pass of the control system during system testing and balancing, but bypassing any controls should be autho− rized by the system operator and all switches returned to ?auto" mode when the TAB work is complete. Digital Input

A digital input is an on/off, open/closed, hi/low, or oth− er two position feedback signal input to the computer− ized control system from the HVAC equipment being monitored.

Four basic EMCS signals:

HVAC systems it controls. After this software has been created on a desktop computer, it is ?downloaded" to each field controller. Since the actual sequence of con− trols and operating schedules reside in the field stand alone controller, the central control computer is only needed when operating schedules or system setpoints need to be changed by the operator, or to display sys− tem alarms sent from the field controllers. Since the programmable stand alone controllers are a micro− computer device, a field interface device provides re− lays and analog to digital transducers which allows the tiny electronic circuits to control larger current field devices like motor starters and damper operators.

Digital Output

A digital output is an on/off, open/close, hi/low, or oth− er two position control signal command output from the computerized control system to the HVAC system being controlled. 4.7.3

Analog Input

An analog input is a variable feedback signal to the computerized control system from the HVAC system being monitored. It could indicate the position of a

DESKTOP COMPUTER FIELD INTERFACE DEVICES & RELAYS

PROGRAMMABLE STAND ALONE CONTROLLER(S)

ACTUATORS

FIGURE 4-5 FUNCTIONAL BLOCK DIAGRAM A CENTRALIZED COMPUTER CONTROL SYSTEM 4.8


valve or damper, speed of a fan or pump, or a room temperature or humidity. 4.7.4

Analog Output

An analog output is a variable command signal output from the computerized control system to the HVAC system being controlled. This signal could be adjust− ing a 0 to 20 psi pressure to a valve or damper motor, or a 0 to 10 volt signal to a variable speed fan motor drive, or other variable signal to match the device be− ing controlled. 4.7.5

EMCS Communications

In addition to the above control inputs and outputs, al− most all computerized building control systems have the ability to communicate with other field panels or central monitoring displays and alarm printers.

vision of the Centralized Control installer or ATC con− tractor. Also, readings obtained from centralized sys− tems can be used by the TAB technician to balance the HVAC system being controlled. 4.7.7

How an EMCS Helps TAB Work

A person entering the TAB field may be wondering what all this has to do with the testing and balancing field. Prior to micro−electronics and DDC, all HVAC sys− tems were controlled by pneumatic devices. Figure 4−6 shows a typical pneumatic control cabinet for a large air handling unit. Notice the high concentration of ?spaghetti" tubing which are used to interconnect all of the pneumatic control devices including pneu− matic relays, receiver/controllers, and various pneu− matic logic controls.

Originally, this communication was by direct hard wire or dial up phone connection, but the trend today is towards less manufacturer specific and more open communication ?protocols," allowing any computer on the same computer ?network" having the proper software and access codes to view and/or change any HVAC system on the same network. This system of in− terconnect also reduces the need for multiple sensors. For example, one outside air sensor can now have its present temperature reading accessed by all air han− dling units and their controls for outside temperature reset instead of a separate sensor for each. 4.7.6

EMCS Points List

In addition to the control contractor providing a writ− ten sequence of controls and related control diagrams, the control documentation for a computerized automa− tion control system should also include a ?points list." This list is actually a table or chart, which at a glance indicates each physical piece of equipment being con− trolled down the ?y" axis, and the different types of software programs utilized across the top ?x" axis. The type of points, analog input (AI) analog output (AO), digital input (DI), and digital output (DO) is also indi− cated. By placing an ?x" or ?dot" where each column and row intersect, it is easy to see the interaction of the physical world with the software programming. Reviewing this points list is helpful to the TAB technician to under− stand how a system is being operated. The TAB technicians should not attempt to adjust or change control settings except when under the super−

FIGURE 4-6 HVAC CONTROLS PANEL WITH ORIGINAL PNEUMATIC CONTROLS.


4.9

Unfortunately, many times this calibration has not been completed prior to the TAB work. For this reason, if you have access to the automation system display during the balancing work, be sure to record air and water flows being displayed at the time of your own measurements and advise the automation system con− troller. This is especially critical on variable air volume sys− tems since out of calibration automation flow stations can cause an air handling unit that was just balanced to produce much higher or lower flow rates than in− tended. Figure 4−8 shows a TAB technician using a small por− table laptop computer to adjust the minimum and max− imum air flows on an above ceiling VAV box. Note how the computer is ?plugged" directly into a jack that is provided under each electronic wall thermostat con− nected to a DDC system. The computer screen is displaying actual VAV box dis− charge air cfm, discharge air temperature, percent damper position, reheat coil discharge temperature, branch duct supply temperature, duct static pressure, maximum discharge cfm setpoint, and minimum dis− charge cfm setpoint. All of these values can be easily read and adjusted if necessary during system testing

FIGURE 4-7 THE SAME HVAC CONTROL PANEL AFTER UPGRADING TO DIRECT DIGITAL CONTROL (DDC).

Figure 4−7 shows the same control cabinet after all of the pneumatic controls and control tubing were re− placed with a DDC system. Now changing discharge air temperature from the unit or adjusting the outside air damper is as simple as moving the ?mouse" across the control screen in the building manager’s office. Many building automation systems include air flow and water flow measuring stations in addition to the many temperature and humidity sensors. Before being tempted to use these easy to read values, keep in mind that these control input devices require calibration in the software program which may not have been com− pleted. On new construction projects, many control contractors will install all of these field devices using factory ?default" settings. This allows faster system startup and the default valves are acceptable for initial startup operation. 4.10

FIGURE 4-8 PORTABLE COMPUTER PLUGGED INTO ELECTRONIC WALL THERMOSTAT DURING SYSTEM BALANCING.


and balancing without the need to climb a ladder or re− move ceiling tiles and access covers. Since each control manufacturer may have their own custom software and plug−in thermostat to computer cables, most TAB technicians have developed a good working relationship with the control system installers

who may be able to provide these programs and cables at little or no cost. The ability of a TAB technician to use these portable control devices can reduce the time the control contractor needs to be on site during the TAB work, and in return, the TAB technician does not need to schedule his site times to meet the availability of the control contractor.


4.11


4.12


CHAPTER 5

FANS

CHAPTER 5 5.1

FAN CHARACTERISTICS

5.1.1

Introduction

This chapter will apply the basic airflow fundamentals discussed in Chapter 2 HVAC Fundamentals to fans. As stated earlier, each type of system needs a pump to overcome the friction and dynamic losses of the sys− tem. This device can be either a centrifugal pump, a fan, a compressor, a turbine, or some other sophisti− cated device. Therefore, each device must be studied by the TAB technician so that not only all of its unique characteristics are known, but that the device has been applied properly within the system, and that the system has been designed to circulate fluid in the most eco− nomical manner and to provide maximum comfort. With an in−depth understanding of HVAC fans and their relationship to HVAC systems, it becomes easy for the TAB technician to apply proper balancing pro− cedures in the correct sequence when on the job. Centrifugal Fans

Three basic types of fans are used in HVAC systems, the centrifugal fan or blower, the axial flow fan, and special designs using fans or blowers in different hou− sings. The airflow within the centrifugal fan is sub− stantially radial through the wheel, while the airflow through the axial flow fan is parallel to the fan shaft. The components of centrifugal fans are identified in Figure 5−1. The three variations of the centrifugal fan used in HVAC work are forward curved, backward in− clined, and airfoil. 5.1.2.1

*SCROLL SIDE SCROLL PIECE SIDE SHEET SIDE PLATE

*OUTLET DISCHARGE

*BACKPLATE HUB DISK HUBPLATE *BLADES FINS INLET CONE INLET RING INLET BELL INLET FLARE INLET NOZZLE VENTURI

*SCROLL CASING HOUSING *IMPELLER WHEEL SCROLL HOUSING *RIM VOLUTE MOTOR SHROUD WHEEL RING WHEEL CONE *SUPPORTS RETAINING RING STIFFENERS INLET RIM *INLET COLLARWHEEL RIM INLET SLEEVEFLANGE INLET BAND INLET PLATE * PREFERRED NOMENCLATURE

PEDESTAL

FIGURE 5-1 CENTRIFUGAL FAN COMPONENTS shape of its performance curve, which allows the pos− sibility of overloading the motor, if system static pres− sure decreases. It also is not suitable for material han− dling because it has an inherently weak structure. Therefore, FC fans are generally not capable of the high speeds necessary for developing higher static pressures. STATIC PRESSURE CURVE STATIC EFFICIENCY CURVE BHP CURVE

100

Forward Curved (FC) Fans

The FC centrifugal fan turns at a relatively slow speed and generally is used for producing high airflow vol− umes at low static pressures. The FC fan will surge, but the magnitude is less than for other types. The static pressure proportion of the total pressure discharge is 20 percent while the velocity pressure is 80 percent. Typical operating range of this type of fan is from 30 percent to 80 percent wide open volume (see Figure 5−2). The maximum static efficiency of 60−80 percent generally occurs slightly to the right of peak static pressure. The horsepower curve has an increasing slope and therefore is referred to as an overloading type fan.

70 SE. SP AND BHP

5.1.2

FANS

0

30

80

100

FIGURE 5-2 CHARACTERISTIC CURVES FOR FC FANS 5.1.2.2

Advantages of the FC fan are its low cost, and the slow speed which minimizes shaft and bearing size, and its wide operating range. Disadvantages include the

0

Backward Inclined (BI) Fans

Backward inclined fans travel at about twice the speed of the FC fan. The normal selection range of the BI fan


5.1

is approximately 40−85 percent of wide open airflow volume (see Figure 5−3). Maximum static efficiency of about 80 percent generally occurs close to the edge of its normal operating range. Generally, using a larger fan will allow greater efficiency for a given selection. The magnitude of surge for a BI fan is greater than for the FC fan.

STATIC PRESSURE CURVE STATIC EFFICIENCY CURVE BHP CURVE

100


SE. SP AND BHP

86

SE., SP AND BHP

100

80 0

50

CFM

85

100

FIGURE 5-4 CHARACTERISTIC CURVES FOR AIR FOIL are shown in Figure 5−4. For a specific application, the airfoil fan has the highest rpm of the three centrifugal fans. 5.1.3 0

0

40

CFM

85

100

FIGURE 5-3 CHARACTERISTIC CURVES FOR BI FANS Advantages of the BI fan include its higher efficiency and non−overloading horsepower curve. The horse− power curve generally reaches a maximum in the middle of the normal operating range, thus overload− ing is normally not a problem. Inherently stronger de− sign makes it suitable for the higher static pressure op− eration of 70 percent of the total pressure measured at the fan discharge. This leaves the measured velocity pressure at only 30 percent. Disadvantages include the higher speed, which re− quires larger shaft and bearing sizes and places more importance on proper wheel balance; and unstable op− eration, which occurs as block−tight static pressure is approached. 5.1.2.3

Airfoil Fans

A refinement of the flat bladed BI fan is a fan that uses airfoil shaped blades. This improves the static effi− ciency to about 86 percent and reduces noise level slightly. The magnitude of surge also increases with the airfoil blades. Characteristic curves for airfoil fans 5.2

0

Axial Fans

Components of axial fans are illustrated in Figure 5−5. HVAC axial fans may be divided into three groups, propeller, tubeaxial, and vaneaxial.

GUIDE VANE

INLET CONE OR INLET BELL WHEEL ROTOR IMPELLER

MOTOR

BLADE

HOUSING CASING

HUB

NOSE COVER PLATE SPINNER

FIGURE 5-5 AXIAL FAN COMPONENTS 5.1.3.1

Propeller Fans

HVAC propeller fans normally are not connected to duct systems. They are well suited for handling high volumes of air at very low or no static pressures and low efficiencies (see Figure 5−6). 5.1.3.2

Tubeaxial and Vaneaxial Fans

Tubeaxial and vaneaxial fans are simply propeller fans mounted in a cylinder and are similar except for vane−




100

SE.SP ANDBHP

SE. SP AND BHP

100

50

80

0 0

0

CFM

65

0

FIGURE 5-6 CHARACTERISTIC CURVES FOR PROPELLER FANS type straighteners on the vaneaxial. These vanes re− move much of the swirl from the air and improve the efficiency. A vaneaxial fan is more efficient than a tu− beaxial fan and can reach higher pressures. Note that with axial fans the brake horsepower (BHP) is maxi− mum at the blocktight static pressure (see Figure 5−7). Tubeaxial fans and vaneaxial fans generally are used for handling large volumes of air at low static pres− sures. Advantages of tubeaxial fans and vaneaxial fans in− clude the reduced size and weight, and the straight− through airflow which frequently eliminates elbows in the ductwork. The maximum static efficiency of an in− dustrial vaneaxial fan is approximately 65 percent. The operating range for axial fans is from 65 percent to 90 percent. Disadvantages of axial fans include high noise levels and efficiencies lower than those of centrifugal fans. Special Designs

There are variations of both centrifugal and axial fans that are designated special design fans. These include tubular centrifugal fans and power roof ventilators.

65

90 100

FIGURE 5-7 CHARACTERISTIC CURVES FOR VANEAXIAL FANS (HIGH PERFORMANCE) 5.1.4.1

5.1.4

CFM

100

Tubular Centrifugal Fans

Tubular centrifugal fans, illustrated in Figure 5−8, gen− erally consist of a single width airfoil wheel arranged in a cylinder to discharge air radially against the inside of the cylinder. Air is then deflected parallel with the fan shaft to provide straight−through flow. Vanes are used to recover static pressure and to straighten air flow.

SW CENTRIFUGAL FAN WHEEL STREAMLINE INLET

AIR OUT AIR IN

STRAIGHTENING VANES

FIGURE 5-8 TUBULAR CENTRIFUGAL FAN Characteristic curves are shown in Figure 5−9. The selection range is generally about the same as the scroll type BI or airfoil bladed wheelC50 to 85 per− cent of wide open volume. However, because there is


5.3

limitations of the wheels, bearings, and housing of fans. Under the most recent class standards, there are three classifications, as shown in Figures 5−10 and 5−11.


Note the line of demarcation between Class I and Class II construction.

100

SE. SP AND BHP

Example 5.1 (I−P) A fan operates at 9056 cfm, 1478 rpm, requiring 5.08 BHP at 1.0 in. wg SP. The airflow must be increased to 10,188 cfm to handle an additional load. Find the new SP and BHP.

70

15 14 13 1/2” @ 3780 13 RATINGS MAY BE PUBLISHED IN THIS UPPER RANGE

12

0

50

85

100

CFM

FIGURE 5-9 CHARACTERISTIC CURVES FOR TUBULAR CENTRIFUGAL FANS no housing of the turbulent air flow path through the fan, static efficiency is reduced to a maximum of about 72 percent and noise level is increased.

(SP)INCHES OF WATER

0

10

STATIC PRESSURE

11

6

MINIMUM PERFORMANCE CLASS III

9 8 1/2” @ 3000 8

CLASS III SELECTION ZONE

7 6 3/4” @ 5260

5” @ 2300

4

2 1/2” @ 3200

1000

2000

Fan Classes

3000

4000

5000

6000

7000

OUTLET VELOCITY (OV) FEET PER MINUTE

FIGURE 5-10 FAN CLASS STANDARDS (I-P) (SW BI FANS) 3750 3500 3375 Pa @ 18.9 3250 RATINGS MAY BE PUBLISHED IN THIS UPPER RANGE

3000 2750

STATIC PRESSURE (SP)-P ASCALS (Pa)

5.2.1

RATINGS MAY BE PUBLISHED IN THIS LOWER RANGE

1

TYPICAL CLASS II CHARACTERISTIC CURVE

2500

FAN CONSTRUCTION

4 1/2” @ 4175

CLASS I SELECTION ZONE

2

Power roof ventilators allow the air to discharge in a full circle from the impeller, which may be either cen− trifugal or axial with similar characteristics. A large advantage is that they provide positive exhaust ven− tilation over gravity ventilators.

5.2

CLASS II SELECTION ZONE

MINIMUM PERFORMANCE CLASS I

Power Roof Ventilators

Disadvantages include lower available static pressures than centrifugal fans and loss of the discharge velocity pressure component that is recovered.

MINIMUM PERFORMANCE CLASS II

5

3

Frequently, the straight−through flow results in signifi− cant space savings. This is the main advantage of tubu− lar centrifugal fans. 5.1.4.2

TYPICAL CLASS II CHARACTERISTIC CURVE

MINIMUM PERFORMANCE CLASS III

2250 2125 Pa @ 15.0 2000

CLASS III SELECTION ZONE

1750 1563 Pa @ 26.3 1500 MINIMUM PERFORMANCE CLASS II

1250 Pa @ 11.5 1250 1000

MINIMUM PERFORMANCE CLASS I

750

1063 Pa @ 20.9

CLASS II SELECTION ZONE

625 Pa @ 16.0

When using a fan rating table published by the fan manufacturer, if fan speeds and static pressures in− crease above certain given conditions, the class of the fan changes. Class again refers to an Air Movement and Control Association, Inc. (AMCA) standard which has been developed to regulate actual structural 5.4

500

CLASS I SELECTION ZONE

RATINGS MAY BE PUBLISHED IN THIS LOWER RANGE

250

5

10

15

20

25

30

OUTLET VELOCITY-METRES PER SECOND (m/s)

FIGURE 5-11 FAN CLASS STANDARDS (SI) (SW BI FANS)


35

SW—Single Width SI—Single Inlet

DW—Double Width DI—Double Inlet

Arrangements 1, 3, 7 and 8 are also available with bearings mounted on pedestals or base set independent of the fan housing. For designation of rotation and discharge (see Figure 5–17) For motor position, belt or chain drive (see Figure 5–16)

ARR. 2 SWSI For belt drive or direct connection. Impeller overhung Bearings in bracket supported by fan housing.

ARR. 3 SWSI For belt drive or direct connection. One bearing on each side and supported by fan housing.

ARR. 1 SWSI For belt drive or direct connection. Impeller overhung. Two bearings on base.

ARR. 3 DWDI For belt drive or direct connection. One bearing on each side and supported by fan housing.

ARR. 4 SWSI For belt drive. Impeller overhung on prime mover shaft. No bearings on fan. Prime mover base mounted or integrally directly connected.

ARR. 7 SWSI For belt drive or direct connection. Arrangement 3 plus base for prime mover.

ARR. 7 DWDI For belt drive or direct connection. Arrangement 3 plus base for prime mover.

ARR. 8 SWSI For belt drive or direct connection. Arrangement 1 plus extended base for prime mover.

ARR. 9 SWSI For belt drive. Impeller overhung, two bearings with prime mover outside base.

ARR. 10 SWSI For belt drive. Impeller overhung, two bearings with prime mover inside base.

FIGURE 5-12 DRIVE ARRANGEMENTS FOR CENTRIFUGAL FANS HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

5.5


5.6


VEL

OUT VEL

Press

CFM

FPM

H20

2264 2547

800 900

2830 3113

VOL


0.125

SP

0.250

SP

0.375

SP

0.500

SP

0.625

SP

0.750

SP

0.875

SP

1.000

SP

1.250

SP

1.500

SP

1.750

SP

2.000

SP

RPM

BHP

RPM

BHP

RPM

BHP

RPM

RPM

RPM

BHP

RPM

BHP

RPM

BHP

RPM

BHP

RPM

BHP

BHP

BHP

RPM

BHP

RPM

BHP

0.04 0.05

398 434

0.10 0.13

456 487

0.15 0.19

507 536

0.21 0.25

567 578

0.26 0.30

608 624

0.32 0.37

656 669

0.40 0.44

703 712

0.47 0.51

747 755

0.55 0.60

835

0.78

1000 1100

0.06 0.08

472 510

0.17 0.21

519 552

0.23 0.27

565 595

0.29 0.34

608 636

0.36 0.42

645 675

0.42 0.49

686 708

0.49 0.56

727 745

0.57 0.63

767 782

0.65 0.71

843 855

0.83 0.89

916 924

1.03 1.10

991

1.31

3396

1200

0.09

549

0.26

587

0.33

627

0.40

666

0.48

702

0.56

738

0.64

768

0.71

802

0.79

870

0.97

936

1.17

999

1.39

1062

1.63

3679

1300

0.11

589

0.32

624

0.39

661

0.47

697

0.55

731

0.64

765

0.73

798

0.81

825

0.89

888

1.07

950

1.26

1012

1.48

1070

1.71

3962

1400

0.12

629

0.39

662

0.46

695

0.54

729

0.63

762

0.72

794

0.81

826

0.91

856

1.01

909

1.17

967

1.37

1026

1.58

1083

1.82

4245

1500

0.14

668

0.46

700

0.54

730

0.62

762

0.72

794

0.81

825

0.91

854

1.01

884

1.12

936

1.30

989

1.50

1043

1.71

1097

1.94

4528

1600

0.16

709

0.55

739

0.63

767

0.72

796

0.81

827

0.91

856

1.02

884

1.13

912

1.23

967

1.46

1013

1.64

1063

1.86

1114

2.09

4811

1700

0.18

749

0.65

778

0.74

805

0.83

832

0.92

860

1.03

888

1.14

915

1.25

942

1.36

994

1.59

1044

1.82

1087

2.01

1134

2.25

5094

1800

0.20

790

0.75

818

0.85

843

0.95

868

1.05

894

1.15

921

1.26

948

1.38

973

1.50

1023

1.74

1073

1.99

1115

2.21

1157

2.43

5377

1900

0.23

830

0.88

857

0.98

882

1.08

906

1.19

930

1.29

955

1.40

980

1.53

1005

1.65

1053

1.90

1100

2.16

1146

2.42

1185

2.64

5660

2000

0.25

872

1.01

897

1.12

921

1.23

944

1.33

966

1.44

989

1.56

1014

1.68

1038

1.81

1084

2.08

1129

2.34

1173

2.61

1217

2.89

5943

2100

0.27

913

1.16

937

1.27

960

1.39

982

1.50

1004

1.61

1025

1.73

1048

1.85

1071

1.99

1116

2.26

1160

2.54

1202

2.82

1245

3.12

6226

2200

0.30

954

1.32

977

1.44

999

1.56

1021

1.68

1042

1.80

1062

1.91

1083

2.04

1104

2.17

1148

2.46

1191

2.75

1231

3.04

1272

3.34

6509

2300

0.33

995

1.50

1017

1.62

1039

1.75

1059

1.87

1080

1.99

1100

2.12

1119

2.24

1139

2.38

1181

2.57

1222

2.97

1262

3.28

1301

3.58

6792

2400

0.36

1037

1.70

1067

1.82

1079

1.95

1099

2.08

1118

2.21

1137

2.34

1156

2.47

1175

2.60

1215

2.90

1255

3.20

1293

3.52

1331

3.84

7358

2600

0.42

1120

2.13

1139

2.26

1159

2.40

1178

2.55

1196

2.68

1214

2.82

1231

2.97

1248

3.10

1284

3.40

1321

3.72

1358

4.06

1393

4.40

7924 8490

2800 3000

0.49 0.56

1204 1287

2.64 3.23

1221 1303

2.78 3.38

1239 1320

2.93 3.53

1257 1337

3.08 3.70

1274 1353

3.23 3.86

1291 1370

3.38 4.02

1308 1385

3.53 4.18

1324 1401

3.69 4.34

1356 1431

3.99 4.67

1389 1461

4.32 5.00

1424 1492

4.67 5.35

1458 1525

5.03 5.73

9056 9622 10754

3200 3400 3600 3800

0.64 0.72 0.81 0.90

1371 1455 1539 1623

3.90 4.66 5.51 6.46

1386 1469 1552 1636

4.05 4.82 5.68 6.64

1401 1483 1566 1648

4.21 4.99 5.86 6.82

1417 1498 1579 1661

4.39 5.16 6.04 7.01

1433 1513 1594 1674

4.56 5.35 6.24 7.21

1448 1528 1608 1688

4.74 5.54 6.43 7.42

1464 1542 1621 1701

4.91 5.72 6.63 7.63

1478 1556 1636 1714

5.08 5.91 6.82 7.84

1507 1583 1661 1740

5.43 6.27 7.20 8.25

1535 1611 1687 1764

5.77 6.64 7.59 8.65

1663 1637 1713 1768

6.13 7.00 7.99 9.06

1593 1664 1737 1813

6.51 7.39 8.37 9.48

11320

4000

1.00

1707

7.52

1719

7.70

1731

7.89

1743

8.09

1755

8.29

1769

8.52

1781

8.74

1794

8.95

1818

9.39

1841

9.80

1865

10.24

1888

10.68

10188

IN

Pressure class limits:

Class I II

Maximum RPM 1550 2140

Table 5-1 Typical Fan Rating Table 5.7

MOTOR LEFT

VIEW FACING DISCHARGE

FIGURE 5-13 ARRANGEMENT 1 IN-LINE FANS for motors too large for fan casing. Arrangement 4 (Figure 5−14)Cdirect drive with wheel overhung on motor shaft.

Solution New static pressure

1.0 10188 9056 2

cfm2 P 2 P1 cfm1

2

1.27in.8 8 wgSP New brake horsepower: BP 2 BP1

cfm

5.08 10188 cfm 9056 3

3

2 1

BP 2 7.23BHP A review of Table 5−1 not only confirms the calcula− tions, but also indicates that the change in capacity moved the fan into a different pressure classification which could result in a failure of the fan wheel and/or bearings. CAUTION)Always check with the published rat− ings of fan equipment to make sure that revised op− erating conditions do not require a different class fan. Often this type of change also could change the pressure classification of part or all of a duct system to higher duct construction and sealing require− ments. 5.2.2

Fan Nomenclature

5.2.2.1

Drive Arrangements

AMCA has developed standard fan drive arrange− ments (shown in Figure 5−12) for various bearing and drive locations. Axial or in−line fans are designated in much the same way as standard centrifugal fans. Stan− dard arrangements for in−line are: Arrangement 1 (Figure 5−13)Cbelt drive with motor mounted independent of fan casingCtypically used 5.8

Arrangement 9 (Figure 5−15)Cbelt drive with motor located on periphery of casing in one of eight standard locations designated by the letters beginning with A at the top and proceeding clockwise at eight equal inter− vals through the letter H when viewing the fan from the discharge. Vertical units are designated as either upblast or down− blast and generally are available only in Arrangements 4 and 9. 5.2.2.2

Motor Arrangement Locations

Motor location is specified at W, X, Y, and Z as shown in Figure 5−16. This motor location always is deter− mined by facing the fan drive sheave. It is independent of the discharge or rotation. 5.2.2.3

Rotation

Rotation is determined by the direction the fan wheel will be turning for proper operation as viewed from the drive side of the fan. Rotation is designated as clock− wise (CW) or counter clockwise (CCW). 5.2.2.4

Non-Sparking Construction

For applications where sparks generated in the air stream could be dangerous, AMCA provides three non−sparking construction classifications based on the degree of assurance desired. For all classes, bearings must be out of the air stream, the fan must be grounded, and non−sparking belts are required. The three classes are: 1.

AMCA A. Requires all components in the airstream be made of nonferrous material.


VIEW FACING DISCHARGE

FIGURE 5-14 ARRANGEMENT 4 IN-LINE FANS

A B

H

G

C

D

F MOTOR SHOWN IN POSITION A

E VIEW FACING DISCHARGE

FIGURE 5-15 ARRANGEMENT 9 IN-LINE FANS 2.

AMCA B. This requires all components in the airstream be made of nonferrous material. Housing can be steel.

3.

AMCA C. Nonferrous wear ring is required on the inlet cone so that, if the impeller shifts, it will rub the nonferrous material.

Generally, AMCA A is the most expensive and AMCA C is the least expensive. 5.2.3

Fan Motors and Drives

5.2.3.1

General

Most fans are driven at constant speed by constant speed motors, and they generally deliver a constant air quantity. The motors range from small single phase fractional horsepower motors to large polyphase mo− tors. Motors generally are connected to the driven fan by means of a V belt drive which not only transmits power but allows the synchronous speed of the motor such as 1200, 1800, 3600 rpm, to be converted to the lower fan speed. Some small fans have motors directly connected. Some V belt drives have an adjustable speed range by providing a variable drive sheave on which the pitch diameter can be manually adjusted to allow for minor speed variations. Variable air volume (VAV) systems are now commonly used, and some reduce system air− flow by using variable speed motors or drives.


5.9

d.

Ratios should not exceed 8:1.

e.

Belt speed, preferably should not exceed 5,000 fpm (25 m/s), or be less than 1,000 fpm (5 m/s). Best practice is about 4,000 fpm (20 m/s).

f.

Sheaves should be dynamically balanced when used for speeds in excess of 5,000 fpm (25 m/s) rim speed.

FAN MOTOR

Z

W Y

X DRIVE

Equation 5-1 rpm(fan) Pitchdiam.motorpulley rpm(motor) Pitchdiam.fanpulley

FIGURE 5-16 CENTRIFUGAL FAN MOTOR LOCATIONS

5.2.3.3

Drive Installations

When installing or reviewing fan drives, these points should be particularly watched: Of major importance to the TAB technician, is that the V belt drives must be properly aligned before testing, and the belt tension adjusted properly. Too little belt tension results in belt slippage and excessive belt wear. Too much belt tension can cause excessive bearing loading, causing motor bearing or fan bearing failure. One further caution to the TAB technician is that the motor must have sufficient starting torque to over− come the inertia of the fan wheel and drive package. Most HVAC supply air systems do not have this pro− blem. However, in return air or exhaust air systems where design airflow volumes may be high and fan to− tal pressures low, check to assure that the installed mo− tor has sufficient starting torque to accelerate the fan to its design speed. 5.2.3.2

b.

c.

5.10

Be sure that shafts are parallel and sheaves are in proper alignment. Check again after a few hours of operation.

b.

Do not drive sheaves on or off shafts. Wipe shaft, key, and bore clean with oil. Tighten screws carefully. Recheck and retighten after a few hours of operation.

c.

Belts should never be forced over sheaves.

d.

In mounting belts be sure the slack in each belt is on the same side of the drive. This should be the slack side of the drive.

e.

Belt tension should be reasonable. When in operation, the tight side of the belts should be in a straight line from sheave to sheave and with a slight bow on the slack side. All drives should be inspected periodically to be sure belts are under proper tension and are not slipping.

f.

When making replacements of multiple belts on a drive, be sure to replace the entire set with a new set of matched belts.

Drive Design

Regardless of whether drives consist of stock or spe− cial items, there are certain primary conditions to con− sider with respect to the design of satisfactory drives. The conditions most commonly encountered are: a.

a.

Drives should be installed with provisions for center distance adjustment. This is essential, as all belts stretch.

5.3

FAN AIRFLOW AND PRESSURES

5.3.1

Fan Air Volume

Centers should not exceed 2½ to 3 times the sum of the sheave diameters nor be less than the diameter of the larger sheave.

The airflow volume (cfm or L/s) produced by a fan in a given system is independent of the air density.

Arc of contact on the smaller sheave should not be less than 120.

cfm (L/s)Ccubic feet per minute (liters per second) of air handled by a fan at any air density.


Counter Clockwise Top Horizontal

Clockwise Top Horizontal

Clockwise Bottom Horizontal

Counter Clockwise Bottom Horizontal

Clockwise

Counter Clockwise

Counter Clockwise

Clockwise

Up Blast

Up Blast

Down Blast

Down Blast

Counter Clockwise

Clockwise

Clockwise

Counter Clockwise

Top Angular Down

Top Angular Down

Bottom Angular Up

Bottom Angular Up

Clockwise

Counter Clockwise

Counter Clockwise

Clockwise

Bottom Angular Down

Bottom Angular Down

Top Angular Up

Top Angular Up

FIGURE 5-17 DIRECTION OF ROTATION AND DISCHARGE


5.11

scfm (sL/s)Ccubic feet per minute (liters per second) of standard air (0.075 lb/ft3 or 1.2041 kg/m3 density) handled by a fan. 5.3.2

Where it is possible to take field measurements, care must be taken to measure fan total pressure at the fan inlet duct rather than fan static pressure.

Fan Total Pressure (TP)

Fan total pressure is the difference between the total pressure at the fan outlet and the total pressure at the fan inlet. The fan total pressure is a measure of the total mechanical energy added to the air or gas by the fan. This generally can be measured accurately only (as il− lustrated in Figure 5−18) in a test laboratory.

IMPACT TUBE

FAN

FAN

STATIC TUBE

IMPACT TUBE

AIR FLOW

AIR FLOW

SP

FIGURE 5-19 FAN STATIC PRESSURE (SP)

5.3.4 IMPACT TUBE

TP

Fan velocity pressure (Figure 5−20) is the pressure cor− responding to the fan outlet velocity pressure. It is the kinetic energy per unit volume of flowing air.

FIGURE 5-18 FAN TOTAL PRESSURE (TP)

5.3.3

Fan Static Pressure (SP)

Fan static pressure (Figure 5−19) is the fan total pres− sure less the fan velocity pressure. It can be calculated by subtracting the total pressure at the fan inlet from the static pressure at the fan outlet. This is a source of some confusion within the industry, but, by definition: Fan SP = Fan TP (outlet) TP (inlet) Vp (out− let) Also, TP (outlet) SP (outlet) = Vp (outlet)

5.3.5

Fan SP = SP (outlet) TP (inlet)

Fan Outlet Velocity

Fan outlet velocity is the theoretical velocity of the air as it leaves the fan outlet, and is calculated by dividing the air volume in cfm (L/s) by the fan outlet area in square feet (m2). However, all fans have a non−uni− form outlet velocity; that is, the velocity varies over the cross−section of the fan outlet. Therefore fan outlet velocity as calculated above is only a theoretical value that could occur at a point downstream from the fan. All velocity (velocity pressure) readings, including to− tal pressure and static pressure should be taken down− stream in a straight duct connected to the fan discharge where the flow is more uniform. A large portion of the discharge airflow occurs at the side of the fan outlet farthest from the fan shaft. Veloc− ity readings taken at the side of the duct nearest the shaft, may indicate air appearing to flow from the duct back into the fan. 5.3.6

and substituting:

5.12

Fan Velocity Pressure (Vp)

Fan Brake Power

Fan brake power is the actual power required to drive the fan. It is greater than a theoretical air power be−


Example 5.2 (I−P) Find the tip speed of a 30 inch diameter fan wheel ro− tating at 954 rpm.

PITOT TUBE TOTAL PRESSURE

VELOCITY PRESSURE

Solution TipSpeed p D RPM p 30 954 12 12 Tip Speed = 7493 (fpm)

STATIC PRESSURE

Example 5.2 (SI) VELOCITY PRESSURE TOTAL PRESSURE = STATIC PRESSURE

Find the tip speed of a 750 mm diameter fan wheel ro− tating at 954 rpm.

FIGURE 5-20 FAN VELOCITY PRESSURE (VP) Solution cause it includes loss due to turbulence and other inef− ficiencies in the fan, plus bearing losses. Fan brake power is an important value to the TAB technician be− cause it is the power furnished by the fan motor. 5.3.7

Tip8 8 Speed p D RPM p 0.75 954 60 60 Tip Speed = 37.46 (m/s)

TIP SPEED

Also called peripheral velocity, tip speed equals the circumference of the fan wheel times the rpm of the fan and is expressed in feet per minute (meters per second) (Figure 5−21)

RPM

D

Equation 5-2 (I-P) TipSpeed p d RPM 12in.ft. Where: TipSpeed Feetperminute D Wheeldiameter inches RPM Revolutionsperminute Equation 5-2 (SI) p d RPM TipSpeed 60secmin Where: TipSpeed Metersperminute D Wheeldiameter meters RPM Revolutionsperminute

FIGURE 5-21 TIP SPEED

5.4

FAN/SYSTEM CURVE RELATIONSHIP

5.4.1

System Curve

Duct system resistance is the sum of all pressure losses through filters, coils, dampers, and ductwork. The sys− tem curve or system resistance curve (Figure 5−22) is a plot of the pressure that is required to move air through the system. For fixed systems, that is, with no changes in damper settings, etc., system resistance


5.13

varies as the square of the airflow. The system curve for any system is represented by a single curve. For exam− ple, consider a system handling 1,000 cfm (500 L/s) with a static pressure (SP) resistance of 1 in. wg (250 Pa).

System Curve BHP Curve (W) Fan Curve

Operating Point

3

(750)

AND

(1000)

SP

4

(500)

BHP

Static Pressure - In wg (Pa)

5

2 1 (250) 0

1000 (500) AIRFLOW - CFM (l/s)

2000 (1000)

CFM (L/S)

FIGURE 5-22 SYSTEM RESISTANCE CURVE

FIGURE 5-23 OPERATING POINT If the airflow is doubled, the SP resistance will in− crease by that ratio squared (4) to 4 in. wg (1000 Pa). This system curve changes, however, as filters load with dirt, coils start condensing moisture, or when bal− ancing dampers are moved to a new position.

reading across the fan and concluding that if it is at or above design requirements, the airflow is also at or above design requirements. 5.4.3

5.4.2

The system operating point (Figure 5−23), a point at which the fan and system will simultaneously perform, is determined by the intersection of the system curve with the fan performance curve for each designated speed (rpm). Every fan operates only along its perfor− mance curve. If the designed system SP resistance is not the same as the SP resistance in the installed sys− tem, the operating point will move along the fan curve and the SP and volume delivered will not be as calcu− lated. In Figure 5−24 the actual duct system has more pres− sure drop then predicted by the system designer. Thus, airflow is reduced because the SP increased. The shape of the horsepower curve typically would result in a re− duction in fan power. Typically, the fan rpm would then be increased, and more fan power would be need− ed to achieve the desired airflow. In many cases, when there is a difference between actual and calculated fan output, the difference is due to a change in system re− sistance rather than to any shortcomings of the fan or motor. Frequently, the mistake is made of taking the SP 5.14

Fan Law Relationships

System Operating Point Fan law equations 2−17 and 2−18 from Chapter 2 ap− plying a change only in fan rpm (with the system re− maining unchanged), are graphically shown in Figure 5−25. Use Equation 2−17 to obtain rpm 2: rpm 2 rpm1

Q2 Q1

Then use Equation 2−18 to obtain the new static pres− sure:

rpm P 2 P1 rpm 2 1

2

5.4.4

Density

5.4.4.1

When Volume is Constant

The resistance of an HVAC duct system is dependent on the density of the air flowing through the system. Air density at standard conditions of 0.075 pounds per cubic feet (1.2041 kilograms per cubic meter) is used for rating fans in the HVAC industry. A fan is a constant volume machine and will produce the same


volume of airflow regardless of the air density being handled (see Figure 5−26). The fan SP and fan power, however, will vary directly as the air density increases or decreases.

SP. The fan is drawing 9.22 Bhp from a 10 HP motor. If the fan is run with the oven off (70F ambient), cal− culate the new SP and Bhp. (Air density at 250F = 0.0563 lb/ft3 ).

Equation 5-3 SP 2 d 2 SP 1 d1

Solution Equation 5-4

FP 2 d 2 FP 1 d1

Using Equations 5−3 and 5−4: d2 2.6 0.075 d1 0.0563 SP 2 3.46in.w.g. d FP 2 FP1 2 9.22 0.075 0.0563 d1 FP 2 12.28Bhp With only a 10 HP motor, a 23 percent motor overload occurs. SP 2 SP1

Where (airflow1 = airflow2 ): SP Staticpressure in.wg(Pa) 3 d Density lbft (kgm3) FP Fanpower Bhp(W) In other words, the heavier or more dense the air, the greater the fan power or SP will be.

SP @ RPM

2

System Curve

Fan Curve Actual System Curve Design System Curve

SP

2

New Operation Point Change

SP

SP Increase

SP @ RPM 1 SP

1

AIirflow Reduction CFM (L/S)

FIGURE 5-24 VARIATIONS FROM DESIGN AIR SHORTAGE

Example 5.3 (I−P) A 15,000 cfm fan is delivering 250F air from an oven through an air−to−air heat exchanger against 2.6 in. wg

Airflow 1

Airflow 2

FIGURE 5-25 FAN LAW - RPM CHANGE

Example 5.3 (SI) A 7500 L/s fan is delivering 125C air from an oven through an air−to−air heat exchanger against 650 Pa SP. The fan is drawing 6.88 kW from a 7.5 kW motor. If the fan is run with the oven off (20C ambient) calcu− late the new SP and kW. (Air density at 125C = 0.891 kg/m 3).


5.15

Where: (SP1 = SP2 ):

Solution

Q Airflow cfm(Ls) d Density lbft3(kgm3) RPM FanSpeed FP Fanpower Bhp(W)

Using Equations 6−2 and 6−3: d SP 2 SP1 2 650 1.2041 0.891 d1 SP 2 878Pa d FP 2 FP1 2 6.88 1.2041 d1 0.891 FP 2 9.30kW With only a 7.5 kW motor, a 24 percent motor overload occurs. 5.4.4.2

When Static Pressure is Constant

SP @ d 1

SP1

If the system SP remains constant, the airflow volume, fan speed and fan power will vary inversely as the square root of the density (see Figure 5−27).

Q1 Q2

SP @ d

Chg. 2

Equation 5-5

SP 2

d2 d1

SYSTEM d 1

RPM 1 RPM 2

Equation 5-6 SYSTEM d 2

d2 d1

Airflow1= Airflow 2

FIGURE 5-26 EFFECT OF DENSITY CHANGE (CONSTANT VOLUME) FP 1 FP 2

d2 d1

Equation 5-7 5.4.4.3

Constant Mass Flow

With a constant mass flow rate in a system that remains constant without any changes and using the same fan with a variable drive, the airflow rate, RPM and SP will vary inversely with the air density. The fan brake power will vary inversely with the square of the densi− ty (see Figure 5−28). Equation 5-8 Q1 d 2 Q2 d1 Equation 5-9 RPM 1 d 2 RPM 2 d1 Equation 5-10 SP 1 d 2 SP 2 d1

FP 1 d 2 FP 2 d1

5.16

Equation 5-11 2


Static Pressure

Example 5−4 SI SP@d1

SP@d 2

A fan is required to handle 7500 L/s at 500 Pa, 75C, 950 rpm, 750 meters altitude and 4.2 kW. Find the fan airflow, SP, RPM and fan brake power that must be se− lected from the fan table.

Solution

Chg.

SP1 -SP2

Using an air density correction factor table found in Appendix A, the correction factor to standard condi− tions is 0.78 or: d 2(Actual) 0.78 d 1(Standard) Airflow (d1< d 2)

Using Equations 5−8 to 5−11:

FIGURE 5-27 EFFECT OF DENSITY CHANGE (CONSTANT STATIC PRESSURE)

d2 d1 Q 1 7500 0.78 5850Ls SP 1(Std) 500 0.78 390Pa RPM 1(Std.) RPM 2(Act.) d 2d 1 RPM 2 9500.78 1218rpm FP1(Std.) FP 2(Act.) (d2d 1)2 FP 2 4.2(0.78)2 6.90kW FAN CAPACITY RATINGS Q 1(Std) Q2(Act.)

Where: Airflow cfm(Ls) Q d Density lbft 3(kgm3) RPM Fanspeed SP Staticpressure in.w.g.(Pa) FP Fanpower Bhp(W) 5.5 Example 5.4 (I−P) 5.5.1 A fan is required to handle 15,000 cfm at 2 in. wg, 150F, 950 RPM, 2000 feet altitude and 5.6 Bhp. Find the fan cfm, SP, Bhp and RPM that must be selected from the fan table.

Solution Using an air density correction factor table found in Appendix A, the correction factor to Standard condi− tions is 0.81 or: d 2(Actual) 0.81 d 1(Standard)

Using Equations 5−8 to 5−11:

Fan Testing

Most fan manufacturers rate the performance of their products from tests made in accordance with ANSI/ AMCA Standard 210, Laboratory Methods of Testing Fans for Rating. The purpose of Standard 210 is to es− tablish uniform methods for laboratory testing of fans and other air moving devices to determine perfor− mance in terms of flow rate, pressure, power, air densi− ty, speed of rotation and efficiency, for rating or guar− antee purposes. Two basic methods of measuring airflow are included, the Pitot tube and the long radius flow nozzle. These are incorporated into a number of different setups or figures. In general, a fan is tested on the setup which most closely simulates the way in which it will be installed in an HVAC system. Centrif− ugal, tubeaxial and vaneaxial fans are usually tested with only an outlet duct. Figure 5−29 is a reproduction of a test setup from AMCA Standard 210. Note that this particular setup includes a long straight duct connected to the outlet of the fan. A straightener is located upstream of the Pitot


5.17

PL.1

PL.2

PL.3 L 2,3 10 D3 MIN. +0.25

8.5 D 3

-0.00

D3 MIN. D3 5 D3

+0.25 -0.00

D3 t d3

D3

FAN

PITOT TUBE STRAIGHTENER

TRAVERSE

TRANSFORMATION PIECE

THROTTLING DEVICE

FIGURE 5-28 AMCA FAN TEST - PITOT TUBE tube traverse to remove swirl and rotational compo− nents from the airflow and to ensure that the flow at the plane of measurement is as near to uniform as possible.

A manufacturer may test a fan with or without an outlet duct or inlet duct. Catalog ratings should state whether ducts were used during the rating tests. If the fans are not to be applied with similar duct configurations as used in the test setup, an allowance should be made for the difference in the resulting performance.

Static Pressure

SYSTEM @ d 1

SP2 @ d AND RPM 2

5.5.2

SP2 SP @ d 1 AND RPM 1

Chg.

Airflow1

System Effect

For years, many HVAC system designers, system in− stallers, fan company sales engineers and testing, ad− justing, and balancing (TAB) contractors have found that system total pressure measurements and airflow capacities were considerably less than the fan horse− power and rpm curves indicated.

SP1

Airflow2

FIGURE 5-29 EFFECT OF DENSITY CHANGE (CONSTANT MASS FLOW)

The angle of the transition between the test duct and the fan outlet is limited to ensure that uniform flow will be maintained. A steep transition, or abrupt change of cross−section would cause turbulence and eddies, and lead to non−uniform flow. 5.18

Uniform flow conditions ensure consistency and re− producibility of test results and permits the fan to de− velop its maximum performance. In any installation where uniform flow conditions do not exist, the fan’s performance will be reduced.

This derating of duct system fans is called system ef− fect, and it is very important that this phenomenon be taken into account by all concerned with HVAC sys− tems if they are to operate as designed. System effect diminishes a fan’s performance because of the interaction of the fan and the connected duct sys− tem; and system effect factors are used to compensate for the fan’s decreased performance. In general, sys− tem effect factors are approximations obtained from many research studies. Some studies have been pub− lished previously by individual fan manufacturers, and


many represent the consensus of engineers with con− siderable experience in fan applications. Fans of different types (and even fans of the same type from different manufacturers) will not necessarily react with a duct system in exactly the same way. Therefore, it is necessary to use judgment, based on ac− tual experience, in applying the system effect factors. 5.5.2.1

Fan Selection

Figure 5−30 illustrates deficient fan/system perfor− mance caused by system effect. HVAC system pres− sure losses have been determined and the fan selected to operate at Point 1 (system curve A). However, no al− lowance has been made for the effect of poor duct con− nections to the fan. To compensate, a system effect fac− tor must be added to the calculated system pressure losses to determine a new system curve that is then used to select the fan. The point of intersection between the fan performance curve and this new ?phantom" system curve B is Point 4. Therefore, the actual system flow volume is defi− cient by the difference from Point 1 to Point 4. To achieve the design airflow volume, a system effect factor equal to the pressure difference between Point 1 and Point 2 should be added to the calculated system pressure losses. The fan should be selected to operate at Point 2 where the new corrected rpm curve crosses phantom system curve B. A higher fan brake horse− power will also be required.

corrected fan, the airflow volume and static pressure will be established as point 1, because that is where the system actually is operating. The system is not operat− ing on the phantom system curve, which was used only to select the derated capacity fan. System effect cannot be measured in the field, but only calculated after a visual inspection is made of the fan/duct system con− nections. Because system effect is velocity related, the differ− ence between Points 1 and 2 is greater than the differ− ence between Points 3 and 4. The system effect factor includes only the effect of the system configuration on the fan’s performance. All duct fitting pressure losses are calculated as part of the HVAC system pressure losses and are part of system curve A. 5.5.2.2

Figure 5−31 shows the changes in velocity profiles from the fan outlet to where a uniform velocity profile has developed in the duct. The distance of this point from the fan is called the effective duct length. To ob− tain 100 percent of the energy recovery or static regain, duct fittings or abrupt changes in duct configuration should not be used within that space. In other words, any changes to the discharge duct con− figuration within the effective duct length (which dif− fers from the duct configuration used when the fan was tested and rated) may cause the fan to perform less effi− ciently. 5.5.2.3

However, when a testing and balancing technician measures the actual HVAC system conditions with the

PHANTOM CURVE B WITH SYSTEM EFFECT CURVE A CALCULATED DUCT SYSTEM WITH NO ALLOWANCE FOR SYSTEM EFFECT 2

SYSTEM EFFECT LOSS AT DESIGN VOLUME

4

DESIGN PRESSURE

1 3

SYSTEM EFFECT AT ACTUAL FLOW VOLUME

SELECTED FAN CURVE FAN CATALOG PRESSURE-VOLUME CURVE

DEFICIENT PERFORMANCE DESIGN VOLUME

FIGURE 5-30 EFFECTS OF SYSTEM EFFECT

Fan Outlets

Fan Inlets

Power roof exhausters are tested and rated when mounted on a roof curb through which the exhaust air duct passes, so system effect is not a problem. The problem occurs with HVAC centrifugal and axial flow fans that are tested without any inlet obstructions or in− let duct connections. For rated performance, the air must enter the fan uni− formly over the inlet area in an axial direction without pre−rotation. Non−uniform flow into the inlet is the most common cause of reduced fan performance. Such inlet conditions are not equivalent to a simple in− crease in the system resistance, so they cannot be treated as a percentage decrease in the fan airflow and pressure output. A poor inlet condition results in an en− tirely new fan performance. Many system effect curves for various round and rec− tangular elbows may be found in AMCA Publication 201−90, Fans and Systems or the SMACNA HVAC


5.19

BLAST AREA CENTRIFUGAL CUTOFF FAN

OUTLET AREA

DISCHARGE DUCT

LENGTH OF DUCT R

100% EFFECTIVE DUCT LENGTH

a. Round Elbow

FIGURE 5-31 FAN OUTLET EFFECTIVE DUCT LENGTH

System9Duct Design manual. When a suitable (often sizeable) length of duct is used between the fan inlet and return air duct elbow, system effect may be avoi− ded. These improvements help maintain uniform flow into the fan inlet and thereby approach the flow condi− tions of the laboratory test setup. Most often where space is at a premium, the inlet duct will be mounted directly to the fan inlet, as shown in Figure 5−32 b. The reduction in capacity and pressure for this type of inlet condition is impossible to tabulate. The many possible variations in width and depth of the duct influence the reduction in performance to varying degrees. Therefore, this inlet should be avoided. Fans and Systems and the HVAC Systems9Duct De− sign manual states that capacity losses as high as 45 percent have been observed in poorly designed inlets, such as those shown in Figure 5−32 b. Field fabricated or factory designed inlet boxes (see Figure 5−32 c) may often eliminate or substantially reduce system effect at fan inlets. Inlet elbows at axial fans may cause an instability in fan operation in addition to system effect that could re− sult in serious damage to the fan. It is strongly advised that inlet elbows be installed at least three duct diame− ters away from any axial fan inlet.

5.20

b. Rectangular Duct

c. Inlet Box

FIGURE 5-32 NON-UNIFORM FLOW CONDITIONS INTO FAN INLET 5.5.2.4

Field Measurements

Recent research has determined that accurate duct ve− locity measurements cannot be made until a near uni− form velocity profile has developed. This point may vary from 3 to 20 duct diameters downstream from the object causing the turbulence. So, any accurate mea− surements on the discharge side of a fan must be well away from the point where system effect occurs. On the fan inlet, non−uniform airflow, spin in the air− flow or a duct condition that produces a vortex create the problem. When one observes the various poor duct fan inlet conditions normally installed, accurate mea− surements are impossible. However, on the fan inlet side, the system effect loss usually occurs within the entry to the fan wheel, so it is not field−measurable. Finally, system effect is a real and often occurring pro− blem. It may be avoided by using better fan/system duct connections where space permits. it also may be avoided if the installing contractor would order HVAC fans and equipment with the proper inlet and discharge connection configurations so that elbows would not have to immediately change airflow direction. Just re− member that: fan capacity reductions due to system ef− fect cannot be measured in the field by TAB techni− cians, system effect losses are approximate, and that system effect factors must not be confused with duct fitting loss coefficients.


CHAPTER 6

AIR DISTRIBUTION AND DEVICES

CHAPTER 6

AIR DISTRIBUTION AND DEVICES

6.1

AIR TERMINAL BOXES

6.1.2.1

6.1.1

Introduction

Constant flow rate controllers may be of the pneumatic or electric volume regulator type. They typically re− quire internal differential pressure sensing, selector devices, and pneumatic or electric motors for opera− tion.

An air terminal box or terminal unit is a device that controls the volume of conditioned air introduced into a space or zone from the HVAC air duct system. The air terminal box manually or automatically fulfills one or more of the following functions. 6.1.1.1

Pressure

The air terminal box may control the pressure of the discharge airflow. 6.1.1.2

Airflow Rate

The air terminal box may control the rate and velocity of the discharge airflow. 6.1.1.3

Temperature

The air terminal box may mix airstreams of different temperatures or humidities, or include a coil to add additional heating or cooling capacity. 6.1.1.4

Air Blending

6.1.2.2

Constant Airflow

Variable Air Volume

Variable air volume (VAV) controllers incorporate a means to reset the constant volume regulation auto− matically to a different control point within the range of the control device in response to an outside signal, such as from a thermostat. Boxes with this feature are pressure independent and may be used with reheat components. Variable flow rate may also be obtained by using a modulating damper ahead of a constant vol− ume regulator. This arrangement typically allows for variations in flow between high and low limits or be− tween a high limit and shutoff. These boxes are pres− sure dependent and volume limiting in function. Pneu− matic variable volume may be either pressure independent, volume limiting, or pressure dependent, according to the equipment selected. 6.1.3

Box Power Sources

A terminal box commonly integrates a sound chamber to reduce noise generated by the manual damper or flow controller reducing the inlet air velocity or pres− sure. The sound attenuation chamber is typically lined with thermal and sound insulating material and is equipped with baffles. Special sound attenuation in the air discharge ducts usually is not required in smaller boxes.

Terminal boxes can be further categorized as being system powered, wherein the assembly derives all of the energy necessary for its operation from the supply air within the distribution system, or as externally powered, wherein the assembly derives part or all of the energy necessary for its operation from a pneumat− ic or electric outside source. In addition, assemblies are self−contained (when they are furnished with all necessary controls for their operation, including actua− tors, regulators, motors, and thermostats), as opposed to non−self−contained assemblies (where part or all of the necessary controls for operation are furnished by someone other than the terminal box manufacturer). In this latter case, the controls may be mounted on the as− sembly by the assembly manufacturer or may be mounted by others after delivery of the equipment.

6.1.2

6.1.4

Types of Air Terminal Boxes

6.1.4.1

Reheat Terminal Boxes

The air terminal box may mix primary air at high ve− locity and/or secondary air from the conditioned space. 6.1.1.5

Sound Attenuation

Categories

Terminal boxes are typically categorized according to the function of their airflow volume controllers, which are generally either constant or variable air volume (VAV) devices. They are further categorized as being pressure dependent, where the airflow rate through the assembly varies in response to changes in system pres− sure, or as pressure independent where the airflow rate through the device does not vary in response to changes in system pressure.

Reheat terminal boxes add sensible heat to the supply air. They cannot be used in many applications unless the added heat is recovered heat due to energy con− servation codes. Water or steam coils or electric resist− ance heaters can be located within or attached directly to the air discharge end of the box. These boxes typi− cally are single duct, and operation can be either


6.1

constant or variable volume. However, if they are VAV, they must maintain some minimum airflow to accomplish the reheat function. 6.1.4.2

Dual Duct Terminal Boxes

Duct terminal boxes receive warm and cold air from separate air supply ducts in accordance with space re− quirements. Pneumatic and electric volume regulated boxes often have individual modulating dampers and operators to regulate the amount of warm and cool air. When a single modulating damper operator regulates the amount of both warm and cold air, a separate pres− sure reducing damper or volume controller (either pneumatic or mechanical) is needed in the box to re− duce pressure and limit airflow. Specially designed baffles may be required within the unit or at the box discharge to mix varying amounts of warm and cold air and/or to provide uniform flow downstream. Dual duct boxes can be equipped with constant flow rate or vari− able flow rate devices to be either pressure indepen− dent or pressure dependent to provide a number of vol− ume and temperature control functions. 6.1.4.3

Ceiling Induction Boxes

The ceiling induction box provides either primary air or a mixture of primary air and relatively warm air to the conditioned space. It accomplishes this function by permitting the primary air to induce air from the ceil− ing plenum or via inducted return air from conditioned space. A single duct supplies primary air at a tempera− ture cool enough to satisfy all zone cooling loads. The ceiling return air inducted into the primary air is at a higher temperature than the room because heat from recessed lighting fixtures enters the plenum directly. The induction box contains damper assemblies con− trolled by an actuator in response to a thermostat to control the amount of cool primary air and warm in− duced air. As reduction in cooling is required, the pri− mary air flow rate is gradually reduced and the induced air rate is generally increased. Reheat coils can be used in the primary air supply and/or in the induction open− ing to meet occasional interior and perimeter load re− quirements. 6.1.4.4

Fan Powered Boxes

Fan powered boxes differ from the above induction boxes in that they are equipped with a blower. This blower, generally driven by a fractional horsepower motor, draws air from the conditioned space (secon− dary air) to be mixed with the cool primary air. The ad− vantage of fan assisted boxes over basic VAV boxes is 6.2

that for a small energy expenditure to the terminal fan, constant air circulation can be maintained in the space. Fan assisted boxes operate at a lower primary air static pressure than air induction boxes, and perimeter zones can be heated without operating the primary fan during unoccupied periods. Warm air from the ceiling return can be used for low to medium heating loads depend− ing on construction of the building envelope. As the load increases, heating coils in the perimeter boxes can be activated to heat the recirculated plenum air to the necessary level. Fan assisted boxes can be divided into two categories: constant volume and bypass−type units. Constant volume, fan assisted boxes (Figure 6−1) are used when constant air circulation is desired in the space. The unit has two inletsCone for cool primary air from the central fan system and one for the secon− dary air. All air delivered to the space passes through the blower. The blower operates continuously when− ever the primary air fan is on and can be cycled to de− liver heat, as required, when the primary fan is off. As the cooling load decreases, a damper throttles the amount of primary air delivered to the blower. The blower makes up for this reduction of primary air by drawing air in the space or ceiling plenum through the return air opening.

DISCHARGE AIR FAN

(CONSTANT VOLUME)

PRIMARY AIR

DAMPER

RETURN AIR (PLENUM)

FIGURE 6-1 CONSTANT VOLUME FAN-POWERED BOX

In the bypass−type fan assisted box (Figure 6−2), the cool primary air bypasses the blower portion of the unit and is delivered directly to the space. The blower section draws in plenum air only and is mounted in par− allel with the primary air damper. A back draft damper prevents primary air from flowing in to the blower sec− tion when the blower is not energized. The blower in these units generally is energized after the damper in the primary air is partially or completely throttled. Some electronically controlled units gradually in− crease the fan speed as the primary air damper is throttled to maintain constant airflow, while permit− ting the fan to shut off when it is in the full cooling mode.


6.2

VARIABLE AIR VOLUME (VAV) TERMINAL BOXES

6.2.1

Introduction

A VAV system controls the dry bulb temperature with− in a space by varying the volume of supply air rather than the supply air temperature. VAV systems can be applied to interior or perimeter zones, with common or separate fan systems, common or separate air tempera− ture control, and with or without auxiliary heating de− vices. As the VAV boxes on a given supply duct system begin to reduce their air flow, duct pressure controls sense this duct pressure increase and the supply fan air flow is reduced accordingly.

PRIMARY AIR

DAMPER DISCHARGE AIR

6.2.3

Combination Pressure Dependent-Independent Boxes

These combination boxes regulate maximum volume, but the airflow volume below the maximum flow rate varies with the inlet pressure variation. Generally, air− flow will oscillate when pressure varies. They are less expensive than pressure independent units and can be used where pressure independence is required only at maximum volume, where system pressure variations are relatively minor, and where some degree of hunt− ing is tolerable. 6.2.4

Pressure Dependent Boxes

Pressure dependent boxes do not regulate the airflow volume, but position the volume regulating device in response to the thermostat. They are the least expen− sive and should only be used where there is no need for limit control and the system pressure is stable. 6.2.5

Bypass (Dumping) Boxes

(VARIABLE VOLUME) RETURN AIR (PLENUM)

FAN

FIGURE 6-2 BYPASS-TYPE FANPOWERED BOX

A space thermostat can control flow by varying a damper, or a volume regulating device in the duct, or a pressure reducing terminal box. Depending on the complexity of the air distribution system, and consid− erations, VAV may or may not be combined with fan or system static pressure controls. The fan system is designed to handle the largest simul− taneous block load, not the sum of the individual peaks. As each zone peaks at a different time of day, it borrows the extra air from off−peak zones. This transfer of air from low−load to high−load zones occurs only in a true VAV system. 6.2.2

Pressure Independent Boxes

Pressure independent boxes regulate the airflow vol− ume in response to the thermostat’s call for heating or cooling. The required airflow is maintained regardless of fluctuation of the VAV unit inlet or system pressure. These units can be field or factory adjusted for maxi− mum and minimum cfm (L/s) settings. They will oper− ate at inlet static pressures as low as 0.2 in. of water (50 Pa) at maximum system design volumes.

VAV room supply is accomplished in constant volume systems by returning excess supply air into the return ceiling plenum or return air duct, thus bypassing the room. However, this reduction of system volume is not energy efficient. Use generally is restricted to small systems where a simple method of temperature control is desired, initial cost is modest, and energy conserva− tion is deemed unimportant. 6.3

OTHER AIRFLOW DEVICES

6.3.1

Pressure Reducing Valves

Pressure reducing valves or air valves each consist of a series of gang operated vane sections mounted within a rigid casing and gasketed to reduce as much air leak− age as possible between the valve and duct. They usu− ally are installed between a high pressure trunk duct and a lower pressure branch duct. Pressure is reduced by partially closing the valve, which results in a high pressure drop through the valve. This action generates noise, which must be attenuated in the low pressure discharge duct. The length and type of duct lining depend on the amount and frequency of noise to be attenuated. Volume control is obtained by adjustment of the valve manually, mechanically, or automatically. Automatic adjustment is achieved by a pneumatic or electric con− trol motor actuated by a pressure regulator or a thermo− stat. Pressure reducing valves are generally equal in size to the low pressure branch duct connected to the valve


6.3

discharge. This arrangement provides minimum pressure drop with valves opened fully. 6.3.2

thermostatically controlled. The opening varies in approximate proportion to the air volume to maintain discharge throw pattern stability, even with low air quantities. Since these units are pressure dependent, constant pressure regulators are usually required in the duct system. Noise is a particular concern when selecting outlets.

Supply Outlet Throttling Units

The area of the throat or the discharge opening of these supply outlets, which are usually linear diffusers, is

FRAME: 2’ or 1 1_w" ¢ 1 1_w" ¢ 1_i" STRUCTURAL OR FORMED CHANNEL

FRAME 3_i" OR 1_w" DIA.

SHAFTS

18 GA MIN. BLADES

ANGLE STOP 1_w" ¢ 1_w" BAR OPTIONAL

SHAFT EXTENSION

SECTION

FIG. A OPPOSED ACTION

CHANNEL FRAME

PIN & BRONZE BUSHING

CONNECTING BAR FRAME

NOTICE 48" MAX. WIDTH FRAME

STOP SHAFT EXTENSION FIG. B PARALLEL ACTION

SEE TEXT ON VOLUME DAMPERS

SECTION FIG. C

FIGURE 6--3 MULTIBLADE VOLUME DAMPERS

6.4


Volume Dampers

6.3.3.1

Introduction

Volume dampers are primary elements in the duct sys− tem. They are used for controlling airflow rates by introducing a resistance to airflow in the system. In higher pressure systems, the damper is referred to as a pressure reducing valve. Volume control or balancing dampers should be installed in each branch of zone duct. Single leaf dampers which are part of a manufactured air grille are not acceptable for system balancing. Opposed blade dampers which are part of a manufactured air grille can be used if there is not enough room for a regular damper and if sufficient space is provided behind the grille face for proper operation of the damper. Other− wise, a balancing damper should be installed in the branch register termination at a location where it is ac− cessible from the grille or diffuser opening, or a quad− rant damper should be used. Volume dampers installed in branch ducts where the total estimated static pressure is less than 0.5 in. wg (125 Pa) can be single leaf type. Volume dampers installed in ductwork where the total estimated system static pressure exceeds 0.5 in. wg (125 Pa) should be manufactured in accordance with Figure 6−3. 6.3.3.2

Multiblade Dampers

Figure 6−3 shows two types of multiple blade dampers: parallel blade and opposed blade. The terms parallel and opposed refer to the movement of the adjacent bla− des. In the parallel blade damper, all of the blades move in parallel. The opposed blade damper has a linkage which causes the adjacent blades to move in opposite directions. Partial closing of a damper increases the resistance of the duct system to airflow. The reduction in airflow with closure of the damper may or may not be propor− tional to the amount of adjustment of the damper. That is, closing the damper half way does not necessarily mean that the air volume will be reduced to fifty per− cent of that volume which flows through the damper when it is wide open. The relation between the position of the damper and the percent of air that flows through the damper with respect to the airflow through the wide open damper is termed the flow characteristic. Typical flow characteristic curves for parallel blade and opposed blade dampers are shown in Figures 6−4 and 6−5. In Figure 6−4 the flow characteristic curves for

the parallel blade damper show that as the damper is closing, the flow reduction may be proportional to the closing of the damper as is shown by curve J, or partial closure of the damper may have little effect on the flow as is shown by curve A. 100 90 PERCENT OF MAXIMUM FLOW

6.3.3

80 A B

70

C D

60

E F G

50

H J

40

K

30 20 10 0 10

20

30 40 50 60 70 DAMPER POSITION, DEGREES OPEN

80

90

FIGURE 6-4 FLOW CHARACTERISTICS FOR A PARALLEL OPERATING DAMPER The manner in which the damper performs in any duct system is determined by how complicated the system is. If the system is very simple and the damper makes up a major part of the resistance, then any movement of the damper will change the resistance of the entire system and good control of the airflow will result. If the damper resistance is very small in relation to that of the entire system, a poor flow characteristic such as curve ?A" in Figures 6−4 and 6−5 will result. Typical ratios of damper to system resistance are shown in Table 6−1 for each flow characteristic curve. The set of curves for the opposed blade damper (Figure 6−5) shows that for a given ratio of damper to system resistance, a better flow characteristic usually results than with the parallel blade damper (Figure 6−4). As the opposed blade damper is closed, it introduces more resistance to airflow for a given position than a parallel blade damper. When balancing systems, it should be realized that the flow characteristics of a damper are not constant and will vary from one system to another. The actual effect of closing the damper can only be determined by mea− surements in the particular system unless the system designer has taken into account the damper flow char− acteristics in his system design. It is important that the TAB technician understand the airflow patterns of multiblade dampers. The parallel blade damper has a tendency to throw the air toward one side of the duct. This uneven pattern may adverse− ly affect coil or fan performance, or airflow into


6.5

branch ducts if the damper is located just upstream of any system component. 100 PERCENT OF MAXIMUM FLOW

90 80 70

Where dampers should have tight shutoff when closed, the linkage between blades must be properly adjusted. Damper motor linkage also must be properly adjusted. Cold deck and hot deck dampers, as used in multi−zone units, must close tightly, as must face and by−pass dampers used in some air handling units.

A

60

B C

50

D E F

40

G H

30 20 10 0 10

20

30 40 50 60 70 DAMPER POSITION, DEGREES OPEN

80

90

FIGURE 6-5 FLOW CHARACTERISTICS FOR AN OPPOSED OPERATING DAMPER

These flow patterns should be noted when it is neces− sary to measure airflow in a duct near a damper. Where possible, make any measurements upstream rather than downstream of a damper. 6.3.3.3

Quadrants and Linkages

When dampers are located within ducts and are manu− ally controlled, they are usually secured in place with locking linkage or quadrant such as those shown in Figure 6−6. Varying in strength and locking ability, they should be of suitable size for the damper with which they are used. When adjusting a damper, the regulator or quadrant must be tightened securely to en− sure that the damper remains as set.

Parallel−leaf dampers Open damper resistance, percent of system resistance 0.5− 1.0 1.0− 1.5 1.5− 2.5 2.5− 3.5 3.5− 5.5 5.5− 9.0 9.0−15.0 15.0−20.0

Do not always accept the position of the regulator pointer as indicating the actual position of the damper blade. When in doubt, inspect the end of the damper rod at the face of the regulator. A groove, usually cut by a hacksaw, will indicate that the damper blade runs in the same direction as the cut.

Opposed−leaf dampers

Flow character− istic curve

Open damper resistance, percent of system resistance

Flow character− istic curve

A B C D E F G H

0.3− 0.5 0.5− 0.8 0.8− 1.5 1.5− 2.5 2.5− 5.5 5.5−13.5 13.5−25.5 25.5−37.5

A B C D E F G H

6.4

AIR DISTRIBUTION BASICS

6.4.1

Introduction

Air distribution criteria will vary considerably in com− mercial and institutional buildings as well as zone tem− perature and humidity levels. People sitting with little activity require closer tolerances than those actively moving about. Spillover from open refrigerated dis− play equipment in super markets causes frequent com− plaints from customers. An understanding of the prin− ciples of room air distribution helps in the selection, design, control, and operation of HVAC duct systems. The real evaluation of air distribution in a space, how− ever, is if most occupants are comfortable. The object of good air distribution in HVAC systems is to create the proper combination of temperature, hu− midity, and air motion, in the occupied zone of the con− ditioned room from the floor to 6 feet (2 m) above floor level. To obtain comfort conditions within this zone, standard limits have been established as acceptable ef− fective draft temperature. This term includes air tem− perature, air motion, relative humidity, and the physio− logical effects on the human body. Any variation from accepted standards of one of these elements causes dis− comfort to occupants. Lack of uniform conditions within the space or excessive fluctuation of conditions in the same part of the space may produce less than ac− ceptable conditions. Although the percentage of room occupants who ob− ject to certain conditions may change over the years, more recent research has shown that a person tolerates higher air flow velocities and lower temperatures at ankle level than at neck level. Because of this, condi− tions in the zone extending from approximately 30 to 60 inches (0.75 to 1.5 m) above the floor are more criti− cal than conditions nearer the floor.

6.4.2 Table 6-1 Typical Ratios of Damper to System Resistance for Flow Characteristic Curve 6.6

Air Velocity And Air Entrainment

For comfortable air distribution, room air velocities within the occupied zone (floor to 6 feet [2 m] above


ROD. WASHER, LOCK NUT OR HINGE

MIN.-12 " FLOW

SET SCREW 1 ROD TO 24" DEPTH 2 RODS 25" TO 60" 3 RODS 61" & OVER

SPLITTER DAMPER

BEARING OPTION

ROD CONTINUOUS ON 2" WG CLASS AND ON ALL DAMPERS OVER 12" DIA.

NUT

ARM FIG D ELEVATION TWO BLADE ARRANGEMENT

FIG C ROUND DAMPER

DUCT

CIRCULAR DUCTS

3 8

DUCT

" QUADRANT

DUCT

QUADRANT 1" 2

" PIN

ROD-PIN

12" MAX

3 8

1" 2

22 Ga. BLADE. 1" 8

16 Ga. BLADE 1" 8

CLEARANCE ALL AROUND UP TO 18" FIG A

CLEARANCE ALL AROUND 19" TO 48" FIG B

NOTE: OVER 12" HIGH USE MULTIPLE BLADES. SEE FIG 14-3

D

DUCT DEPTH

STIFFEN AS REQUIRED

D

1" 2

FIG A OR B SIDE ELEVATION

RECTANGULAR DUCTS

FIGURE 6-6 VOLUME DAMPERS HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

6.7

floor level) should be in a range of 20 fpm to 70 fpm (0.1 to 0.35 m/s) with 50 fpm (0.25 m/s) normally be− ing used. Stagnant air areas should be avoided as tem− perature in these areas may not be acceptable to the oc− cupants. Room air velocities less than 50 fpm (0.25 m/s) are ac− ceptable; however, even higher velocities may be ac− ceptable to some occupants. ASHRAE Standard 55−1981 recommends elevated air speeds at elevated air temperatures. No minimum air speeds are recom− mended for comfort, although air speeds below 20 fpm (0.1 m/s) are usually imperceptible. The velocity of the air (primary air) emerging from the supply outlet induces air movement within the room area (secondary air). This process of entrainment or capturing of secondary air into the primary air is an es− sential part of air distribution to create total air move− ment within the room, thereby eliminating stagnant air areas and reducing temperature differences to accept− able levels before the air enters the occupied zone. Air entrainment will also tend to overcome natural con− vection and radiation effects within the room. 6.4.3

Surface Effect

Air entrainment takes place only along one surface of the outlet discharge jet when the outlet discharges air directly parallel and adjacent to a wall or ceiling. The

surface effect (Coanda effect) is illustrated in Figure 6−7. Since turbulent jet airflow from a grille or diffuser is dynamically unstable, it may veer rapidly back and forth. When the jet airflow veers close to a parallel and adjacent wall or ceiling, the surface interrupts the flow path on that side as shown in Figure 6−7B. The result is that no more secondary air is flowing on that side to replace the air being entrained with the jet airflow. This causes a lowering of the pressure on that side of the outlet device, creating a low pressure bubble that causes the jet airflow to become stable and remain at− tached to the adjacent surface throughout the length of the throw. The surface effect counteracts the drop of horizontally projected cool airstreams. Ceiling diffusers exhibit surface effect to a high degree because a circular air pattern blankets the entire ceil− ing area surrounding each outlet. Slot diffusers, which discharge the airstream across the ceiling, exhibit sur− face effect only if they are long enough to blanket the ceiling area. Grilles exhibit varying degrees of surface effect, depending on the spread of the particular air pattern. In many installations, the outlets must be mounted on an exposed duct and discharge the airstream into free space. In this type of installation, the airstream en− trains air on both its upper and lower surfaces; as a re− sult, a higher rate of entrainment is obtained and the throw is shortened by approximately 33 percent. Air− flow per unit area for these types of outlets can, there−

SEPARATION BUBBLE CEILING

CEILING JET FLOW

WALL

JET FLOW

ENTRAINED AIRFLOW (SECONDARY AIR)

FLOOR (A)

ENTRAINED AIRFLOW (SECONDARY AIR)

FLOOR (B)

FIGURE 6-7 SURFACE (COANDA) EFFECT

6.8


fore, be increased. Because there is no surface effect from ceiling diffusers installed on the bottom of ex− posed ducts, the air drops rapidly to the floor. There− fore, temperature differentials in air conditioning sys− tems must be restricted to a range of 15F to 20F (8C to 1C). Airstreams from slot diffusers and grilles show a marked tendency to drop because of the lack of surface effect. 6.4.4

Smudging

Smudging may be a problem with ceiling and slot dif− fusers. Dirt particles held in suspension in the secon− dary (room) air are subjected to turbulence at the outlet face. This turbulence, along with surface effect, is pri− marily responsible for smudging. Smudging can be ex− pected in areas of high pedestrian traffic (lobbies, stores, etc.) When ceiling diffusers are installed on smooth ceilings (such as plaster or metal pan), smudg− ing is usually in the form of a narrow band of discolor− ation around the diffuser. Anti−smudge rings may re− duce this type of smudging. On highly textured ceiling surfaces (such as rough plaster and sprayed−on com− position), smudging often occurs over a more exten− sive area. 6.4.5

Sound Levels

The sound level of an outlet is a function of the air dis− charge velocity and the transmission of HVAC equip− ment noise, which is a function of the size of the outlet. Higher frequency sounds can be the result of excessive outlet velocity, but may also be generated in the duct by the moving airstream. Lower pitched sounds are generally the result of mechanical equipment noise transmitted through the duct system and outlet. The cause of higher frequency sounds can be pin− pointed as outlet or equipment sounds by removing the outlet during operation. A reduction in sound level in− dicates that the outlet is causing noise. If the sound lev− el remains essentially unchanged, the system is at fault. Chapter 46 Sound and Vibration Control in the 1999 ASHRAE HVAC Applications Handbook has more information on design criteria, acoustic treat− ment, and selection procedures. 6.4.6

Effect Of Blades

Blades affect grille performance if their depth is at least equal to the distance between the blades. If the blade ratio is less than one, effective control of the air− stream discharged from the grille by means of the blades is impossible. Increasing the blade ratio above

two has little or no effect, so blade ratios should be be− tween one and two. A grille discharging air uniformly forward (blades in straight position) has a spread of 14 to 24, depend− ing on the type of outlet, duct approach, and discharge velocity. Turning the blades influences the direction and throw of the discharged airstream. A grille with diverging blades (vertical blades with uniformly increasing angular deflection from the cent− erline to a maximum at each end of 45) has a spread of about 60, and reduces the throw considerably. With increasing divergence, the quantity of air dis− charged by a grille for a given upstream total pressure decreases. A grille with converging blades (vertical blades with uniformly decreasing angular deflection from the centerline) has a slightly higher throw than a grille with straight blades, but the spread is approximately the same for both settings. The airstream converges slightly for a short distance in front of the outlet and then spreads more rapidly than air discharged from a grille with straight blades. In addition to vertical blades that normally spread the air horizontally, horizontal blades may spread the air vertically. However, spreading the air vertically risks hitting beams or other obstructions or blowing primary air at excessive velocities into the occupied zone. On the other hand, vertical deflection may increase adher− ence to the ceiling and reduce the drop. In spaces with exposed beams, the outlets should be lo− cated below the bottom of the lowest beam level, pre− ferably low enough to employ an upward or arched air path. The air path should be arched sufficiently to miss the beams and prevent the primary or induced air− stream from striking furniture and obstacles and pro− ducing objectionable drafts. 6.5

ROOM AIR DISTRIBUTION

6.5.1

Natural Airflow

The natural air convection currents flowing down the glass during heating, and up the glass during cooling as shown in Figure 6−8, are a major influence on the air distribution in the perimeter zones of a building. During heating, these currents carry cool air down to the floor level causing a stratification of air in layers of increasing temperatures from the floor to the cei− ling. The severity of the temperature gradient depends


6.9

0

75F

95F

75F 80F

HEATING

COOLING

FIGURE 6-8 SOME ELEMENTS AFFECTING BODY HEAT LOSS on outdoor temperature, construction, and air distribu− tion. It is easily understood that warm supply air introduced at the base of the wall would tend to coun− teract these currents and reduce or eliminate stratifica− tion. Optimum air distribution in perimeter zones re− quires perimeter introduction of air or supplementary radiation at the perimeter. During cooling, currents carry warm air up the wall to ceiling level. Stratification then forms from the ceiling down. To eliminate stratification, cool air should be projected into this region near the ceiling. To do this most effectively, supply air outlets should be located high in the wall or in the ceiling. 6.5.2

Supply Air Outlet Performance

6.5.2.1

Outlet Throw

Extensive studies of supply outlet performance have shown that air discharge throw from free round open− ings, grilles, perforated panels and ceiling diffusers are related to the average velocity at the face of the supply outlet or opening. An air jet discharged from a free opening has four zones of expansion and the centerline velocity of the jet in any zone is related to the initial velocity as shown in Figure 6−9. Regardless of the type of outlet, the air stream will tend to assume a circular shape in free space. The important point is that the performance of any supply outlet is related to the initial velocity and initial area as shown in Figure 6−9.

the initial volume of the jet at any distance from the point of origin depends mainly on the ratio of the initial velocity (Vo) to the terminal velocity (Vx). For exam− ple, doubling the initial velocity for the same terminal velocity doubles the induction ratio and also the throw. In zone 4 where the terminal velocity is relatively low and specifically for terminal velocities of 50 fpm (0.25 m/s), the throw should be reduced 20 percent. The buoyant forces with non−isothermal jets cause the air jet to rise during heating and drop during cooling. These conditions result in shorter throws when the throw is reduced to a terminal velocity less than 150 fpm (0.75 m/s) The discussion of throw and drop has been limited to free space applications. If the air discharge jet is pro− jected parallel to and within a few inches of a surface, the jet performance will be affected by the surface, which limits the induction on the surface side of the jet. This creates a low pressure region between the jet and the surface which draws the jet toward the surface. In fact, this effect will prevail if the angle of discharge be− tween the jet and the surface is less than 40. Surface effect will draw the jet from a sidewall outlet to the ceiling if the outlet is within one foot (0.3 m) of the ceiling. The jet from a floor outlet will be drawn to the wall and the jet from a ceiling outlet will be drawn to the ceiling. Surface effect increases the throw for all types of outlets and decreases the drop for horizontally projected air streams. 6.5.2.2

Beyond the second zone, the jet is a mixture of supply and room air. The air jet expands because of induction of room air. The ratio of the total volume of the jet, to 6.10

VAV Applications

Air distribution is very important in VAV applications. Consideration must be given to distribution and to


1

3

2

4

X

PRIMARY

HIGH VELOCITY USUALLY ON OR NEAR ROOM SURFACES

PRIMARY AND INDUCED ROOM AIR

AIR

HIGH TEMP-HEATING LOW TEMP-COOLING

ROOM AIR GENTLE MOVEMENT

GREATEST POSSIBLE SOURCE OF DRAFTS FREE JET—ROOM AIR INDUCED ON ALL SIDES JET NEAR A SURFACE—HIGH VELOCITY AIR HUGS SURFACE AND INDUCES AIR ON ONLY ROOM SIDE OF JET

ZONE

SUPPLY VELOCITY,VO JET CENTERLINE VELOCITY,VX

1

V

2

V

3

V

4

V

X

X

=

V

X

»

V

O

X

»

V

O

O

/

X

/

X

APPROACHES ROOM VELOCITY

FIGURE 6-9 FOUR ZONES IN JET EXPANSION sound levels at maximum and minimum airflow. If the combined sound level of the terminal unit and diffuser at maximum flow is at least 3 dB below the room ambi− ent sound level, variations will not be noticed. In gen− eral, several important considerations are listed for variable volume system air distribution using outlets for horizontal discharge patterns as follows: a.

b.

An outlet with a low throw coefficient should be used. A small throw coefficient gives a smaller absolute change in the throw values with variation in volume and thus tends to minimize the change in air motion within the occupied space due to change in airstream pattern. Outlets should be chosen for small quantities of air. In this manner, absolute values of throw will vary a minimum with the variation in flow rate for the outlet. If the system ap− plication requires modular outlet arrange− ments for occupancy flexibility, as with dif− fusers in combination with light troffers or with ceiling suspensions, no increase in the

number of outlets is necessary to satisfy this requirement. For under window air distribution, vertical throw out− lets with nonspreading pattern should be used. To pre− vent cool air dropping back into the occupied space at minimum flow conditions, the outlet discharge veloc− ity should be 500 fpm (2.5 m/s) minimum. The throw coefficient should be higher to project air up to the cei− ling. With these exceptions, the preceding items also apply to under window distribution. 6.5.3

Supply Outlets

Outlets are selected for each specific room, based on air quantity required, distance available for throw or radius of diffusion, structural characteristics and ar− chitectural concepts being considered. Table 6−2 is based on experience and typical ratings of various out− lets. It may be used as a guide to the outlets applicable for use with various room air loadings. Special condi− tions, such as ceiling heights greater than the normal 8 to 12 feet (2.4 to 3.5 m) and exposed duct mounting,


6.11

FIGURE 6-10 TYPICAL SUPPLY OUTLETS

as well as product modifications and unusual condi− tions of room occupancy, can modify this table.

6.5.3.1

The adjustable blade grille is the most common type of grille used as a supply outlet. The single deflection blades install behind and at right angles to the face bla− des. This grille controls the airstream in both the hori− zontal and vertical planes.

Grille Slot Perforated Panel Ceiling Diffuser Perforated Ceiling

Air Loading, cfm/ft2 (L/s per m2) of Floor Space

Approx. Max. Air Changes @Hour for 10’ (3 m) Ceiling

0.6 to 1.2 (3 to 6) 0.8 to 2.0 (4 to 10) 0.9 to 3.0 (5 to 15) 0.9 to 5.0 (5 to 25) 1.0 to 10.0 (5 to 50)

7 12 18 30 60

Table 6-2 Guide to Use of Various Outlets 6.12

Fixed Blade Grilles

The fixed blade grille is similar to the single deflection grille, except that the blades are not adjustable; the blades may be straight or set at an angle. The angle at which the air is discharged from this grille depends on the type of deflection blades.

Adjustable Blade Grilles

Type of Outlet

6.5.3.2

6.5.3.3

Stamped Grilles

The stamped grille is stamped from a single sheet of metal to form a pattern of small openings through which air can pass. Various designs are used, varying from square or rectangular holes to intricate ornamen− tal designs. 6.5.3.4

Variable Area Grilles

The variable area grille is similar to the adjustable double deflection grille but can vary the discharge area to achieve an air volume change (variable air volume outlet) at constant pressure, so that the variation in throw is minimized for a given change in supply air volume.


6.5.3.5

Slot Diffusers

6.5.3.9

Perforated Face Ceiling Diffusers

A slot diffuser is an elongated outlet consisting of a single or multiple number of slots. It is usually installed in long continuous lengths. Outlets with di− mensional aspect ratios of 25 to 1 or greater and a max− imum height of approximately 3 inches (75 mm) gen− erally meet the performance criteria for slot diffusers.

Perforated metal diffusers meet architectural demands for air outlets that blend into perforated ceilings. Each has a perforated metal face with an open area of 10 to 50 percent which determines its capacity. Units are usually equipped with a deflection device to attain multi−pattern horizontal air discharge.

6.5.3.6

6.5.3.10 Variable Area Ceiling Diffusers

Air-Light Diffusers

Air−light slot diffusers have a single slot discharge in nominal 2, 3, and 4 foot (0.6, 0.9, and 1.2 m) lengths and are available for use in conjunction with recessed fluorescent light troffers. A diffuser mates with a light fixture and is entirely concealed from the room. It dis− charges air through suitable openings in the fixture and is available with fixed or adjustable air discharge pat− terns, air distribution plenum, inlet dampers for bal− ancing, and inlet collars suitable for flexible duct con− nections. Light fixtures adapted for slot diffusers are available in styles to fit common ceiling constructions. Various slot diffuser and light fixture manufacturers may furnish products compatible with one another’s equipment. 6.5.3.7

Multi-Passage Ceiling Diffusers

Multi−passage ceiling diffusers consist of a series of flaring rings or louvers, which form a series of concen− tric air passages. They may be round, square, or rectan− gular. For easy installation, these diffusers are usually made in two parts; an outer shell with duct collar and an easily removable inner assembly. 6.5.3.8

Flush And Stepped-Down Diffusers

Flush and stepped−down diffusers also are available. In the flush unit, all rings or louvers project a plane sur− face, whereas in the stepped−down unit, they project beyond the surface of the outer shell. Common variations of this diffuser type are the adjust− able pattern diffuser and the multi−pattern diffuser. In the adjustable pattern diffuser, the air discharge pat− tern may be changed from a horizontal to a vertical or downblow pattern. Special construction of the diffuser or separate deflection devices allow adjustment. Mul− tipattern diffusers are square or rectangular and have special louvers to discharge the air in one or more di− rections. Other outlets available as standard equipment are half round diffusers, supply and return diffusers, and light fixture air diffuser combinations.

Variable area ceiling diffusers may be round, square, or linear and have parallel or concentric passages or a perforated face. In addition, they feature a means of ef− fectively varying the discharge area to achieve an air volume change (VAV outlet) at constant pressure, so that the variation in throw is minimized for a given change in supply air volume. 6.5.4

Under Floor Distribution

Raised floor plenums for air distribution are primarily used in computer rooms or research facilities having high concentrations of heat generating electronic equipment requiring very clean and cool supply air. Distributing conditioned air under a raised floor al− lows supply diffusers to be located directly under elec− tronic equipment cabinets having high sensible heat generation. To meet these high equipment cooling needs with ceiling supply diffusers would require dis− charge air flows that would be very uncomfortable for most room occupants, and may still not provide ade− quate cooling air flow into these equipment cabinets. In most applications, these spaces are served by a dedi− cated packaged air conditioning unit also located in the space, which discharges supply air directly down and into the under floor cable access plenum. Since the primary purpose of this system is to provide adequate cooling of very expensive electronic equip− ment, it is important for the TAB technician to obtain direction from those responsible for this equipment before balancing these floor registers. In some cases, equipment nameplate data may indicate a design air flow temperature and flow rate, if this information is not provided by the design drawings. It should also be understood that most raised floors are constructed of manufactured removable metal panel sections to allow easy access to the high concentration of wiring and easy addition or removal of equipment. For this reason, the air distribution must remain flex− ible and easy to modify.


6.13

6.5.5

Return Air Inlet Performance

6.5.5.1

Location

Return air and exhaust air inlets should be located to suit architectural design requirements including ap− pearance and compatibility with supply outlets and ductwork. Generally, inlets are installed to return room air of the greatest temperature differential that collects in the stagnant areas. The location of return and ex− haust inlets does not significantly affect air motion. The location of return and exhaust inlets will not com− pensate for ineffective supply air distribution. A return air inlet should not be located directly in the primary airstream from supply outlets. To do so will short circuit the supply air back into the return air sys− tem without allowing it to mix with the room air. The TAB technician should remember that the supply air maintains the conditions within a space by mixing and dilution. Any removal of excess warm or cold air which is allowed to stratify before mixing within the space will permit lower temperature differentials or lower airflow rates. Removal of excess warm or cold air is accomplished with hoods in certain industrial processes. It can also be done by selecting the supply outlet performance to promote the formation of the stagnant zone directly from the local heat gain or loss. The return intake or exhaust would then be located in the stagnant zone. 6.5.5.2

Noise

In addition to the location of the return intake as dis− cussed above, the intake should be sized to return the proper amount of air to the HVAC unit with minimum static pressure requirements and noise levels. In gener− al, most commercial return grilles have a free area of between 45 and 55 percent, because they are designed so that one cannot see through them. With this type of grille, the velocity should not exceed approximately 500 fpm (2.5 m/s) to have reasonable pressure drop re− quirements and a reasonable sound level (see Table 6−3). In general, return air inlets should be sized on available pressure requirements and sound data, rather than relying on indicated free area values. The problem of return inlet noise is the same as that for supply outlets. In computing resultant room noise lev− els from operation of an air conditioning system, the return inlet must be included as a part of the total grille area. The major difference between supply outlets and return inlets is the frequent installation of the later at ear level. When they are so located, the return inlet ve− 6.14

locity should not exceed 75 percent of maximum per− missible outlet velocity.

Inlet Location Above occupied zone Within occupied zone Not near seats Within occupied zone Near seats Door or wall louvers Undercut doors

Velocity Over Gross Inlet Area−fpm (m/s) 800 Up (4.0 Up) 600−800 (3.0−4.0) 400−600 (2.0−3.0) 200−300 (1.0−1.5) 200−300 (1.0−1.5)

Table 6-3 Recommended Return Air Inlet Face Velocities 6.5.6

Return Air Inlets

6.5.6.1

Adjustable Blade Grilles

The same adjustable blade grilles used for air supply are used to match the deflection setting of the blades with that of the supply outlets. 6.5.6.2

Fixed Blade Grilles

The same fixed blade grilles described in the supply air section are used. This grille is the most common return air inlet. Blades are straight or set at a certain angle, the latter being preferred when appearance is important. 6.5.6.3

V-Blade Grille

The V−blade grille is made with blades in the shape of inverted v’s stacked within the grille frame, this grille has the advantage of being sight proof; it can be viewed from any angle without detracting from appea− rance. Door grilles are usually v−blade grilles. The air− flow capacity of the grille decreases as visibility through the grille decreases. 6.5.6.4

Light Proof Grille

The light proof grille is used to transfer air to or from darkrooms. The blades of this type of grille form a lab− yrinth and are painted black. The blades may take the form of several sets of v−blades or be of some special interlocking louver design to provide the required lab− yrinth. 6.5.6.5

Stamped Grilles

Stamped grilles frequently are used as return air and exhaust air inlets, particularly in more demanding areas like rest rooms and utility areas. 6.5.6.6

Ceiling And Slot Diffusers

Supply air ceiling diffusers also may be used as return air and exhaust air inlets.


Neck VelocityCfpm (m/s)

400 (2.0)

500 (2.5)

600 (3.0)

700 (3.5)

800 (4.0)

1000 (5.0)

Round Diffuser Half Round Diffuser Half Round Diffuser, Flush Square Diffuser Square Diffuser, Adjustable Rectangular Diffuser Curved Blade Diffuser Perforated Diffuser High Capacity Diffuser Slimline Diffuser, 2 Way* Extruded Fineline Diffuser 0.25 in.(6.4 mm) Bar Spacing* Linear Slot Diffuser*

0.024 (6.0) 0.035 (8.7) 0.046 (11.5) 0.021 (5.2) 0.036 (9.0) 0.043 (10.7) 0.056 (13.9) 0.037 (9.2) — 0.010 (2.5)

0.039 (9.7) 0.054 (13.4) 0.074 (18.4) 0.033 (8.2) 0.057 (14.2) 0.066 (16.4) 0.090 (22.4) 0.058 (14.4) C 0.015 (3.7)

0.056 (13.9) 0.080 (19.9) 0.106 (26.4) 0.048 (12.0) 0.080 (19.9) 0.096 (23.9) 0.131 (32.6) 0.083 (20.7) C 0.022 (5.5)

0.075 (18.7) 0.107 (26.6) 0.143 (35.6) 0.064 (15.9) 0.112 (27.9) 0.131 (32.6) 0.175 (43.6) C 0.050 (12.5) 0.028 (7.0)

0.096 (23.9) 0.141 (35.1) 0.184 (45.8) 0.083 (20.7) 0.144 (35.9) 0.170 (42.3) 0.225 (56.0) 0.148 (36.9) 0.060 (14.9) 0.040 (10.0)

0.152 (37.8) 0.219 (54.5) 0.290 (72.2) 0.130 (32.4) 0.226 (56.3) C 0.355 (88.4) 0.230 (57.3) 0.100 (24.9) 0.063 (15.7)

0.011 (2.7) 0.051 (12.7)

0.015 (3.7) 0.079 (19.7)

0.024 (6.0) 0.110 (27.4)

0.030 (7.5) 0.150 (37.4)

0.044 (11.0) 0.200 (49.8)

0.069 (17.2) C

Extractor

0.004 (1.0)

0.006 (1.5)

0.010 (2.5)

0.013 (3.2)

0.017 (4.2)

0.023 (5.7)

*Velocity Through Face Open Area

Table 6-4 Air Outlets and Diffusers Total Pressure Loss Average—in. wg (Pa)

VelocityCfpm (m/s)

300 (1.5)

400 (2.0)

500 (2.5)

600 (3.0)

800 (4.0)

1000 (5.0)

0 Deflection

0.010 (2.5)

0.017 (4.2)

0.028 (7.0)

0.038 (9.5)

0.069 (17.2)

0.107 (26.6)

22½ Deflection

0.011 (2.7)

0.019 (4.7)

0.031 (7.7)

0.043 (10.7)

0.078 (19.4)

0.120 (29.9)

45 Deflection

0.016 (4.0)

0.029 (7.2)

0.047 (11.7)

0.064 (15.9)

0.117 (29.1)

0.181 (45.1)

Table 6-5 Supply Registers Total Pressure Loss Average—in. wg (Pa)

Velocity—fpm (m/s)

300 (1.5)

400 (2.0)

500 (3.0)

600 (3.0)

800 (4.0)

900 (4.5)

0.033 (8.2)

0.060 (14.9)

0.092 (22.9)

0.068 (16.9)

0.122 (30.4)

0.187 (46.6)

0.134 (33.4)

0.238 (59.3)

0.302 (75.2)

0.272 (67.7)

0.483 (120.7)

0.055 (13.7)

0.098 (24.4)

0.614 (152.9)

0.152 (37.8)

0.222 (55.3)

0.390 (97.1)

0.496 (123.5)

0.025 (6.2)

0.060 (14.9)

0.080 (19.9)

0.100 (24.9)

0.180 (44.8)

0.230 (52.3)

0.012 (3.0)

0.020 (5.0)

0.032 (8.0)

0.046 (11.5)

0.080 (19.9)

0.102 (25.4)

0.033 (8.2)

0.055 (13.7)

0.088 (21.9)

0.126 (31.4)

0.220 (54.8)

0.275 (68.5)

Register, 0Deflection

0.054 (13.4)

0.090 (22.4)

0.144 (35.9)

0.207 (51.5)

0.360 (89.6)

C

Register, 30 Deflection

0.042 (10.5)

0.070 (17.4)

0.112 (27.9)

0.161 (40.1)

0.280 (69.7)

0.350 (87.2)

Rectangular Diffuser 12 24 in. (300 300 mm) 24 24 in. (600 600 mm) 12 21 in. (300 525 mm) Perforated Return Diffuser (Neck Velocity)

Register, 45 Deflection Register, Perforated Face

Table 6-6 Return Registers Total Pressure Loss Average—in. wg (Pa) HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

6.15


6.16


CHAPTER 7

AIR SYSTEMS

CHAPTER 7

AIR SYSTEMS

7.1

INTRODUCTION

7.1.2

7.1.1

Categories

In general, air systems offer the following advantages:

This chapter covers design and application of air sys− tems used in single and multiple zoning applications. An air system is defined as a system that provides total sensible and latent cooling in the cold air supplied by the system. No additional cooling is required at the ter− minal units. Heating may be accomplished by the same airstream, either from the central system or at the terminal devices. In some applications, heating is ac− complished by a separate air, water, steam, or electric heating system. The term zone implies the provision or the need for separate thermostatic control, while the term room implies a partitioned area which may or may not require separate control. Air systems may be classified into two basic catego− ries: single−path systems and dual−path systems. 7.1.1.1

Single-Path Systems

Single−path systems are those which contain the main heating and cooling coils in a series flow air path, using common duct distribution system at a common air temperature to feed all terminal apparatus.

7.1.2.1

a.

single duct, single zone, constant volume,

b.

single duct, variable air volume (VAV),

c.

single duct, VAV induction, and

d.

single duct zoned reheat.

7.1.1.2

Dual-Path Systems

Dual−path systems are those which contain the main heating and cooling coils in a parallel flow, or series− parallel flow air path, using either: (1) a separate cold and warm air duct distribution system, which is blended at terminal (dual duct systems); or (2) a sepa− rate supply duct to each zone, with blending of warm and cold air at the main supply fan. Dual−path systems may be: a.

dual duct (including dual duct, VAV), and

b.

multi−zone.

Consolidation

Air systems permit centralized location of major equipment, they consolidate operation and mainte− nance in unoccupied areas, and permit maximum choice of filtration systems, odor and noise control, and high quality, durable equipment. There is com− plete absence of drain piping, electrical equipment power wiring, and filters in the conditioned space. 7.1.2.2

Outdoor Air Cooling

The greatest advantage of air systems is the number of free cooling season hours that may be had when out− door air can be used for cooling in lieu of mechanical refrigeration. Economizer control systems usually are more trouble−free than enthalpy control systems. 7.1.2.3

Flexibility

Air systems allow a wide choice of zonability, flexibil− ity, and humidity control under all operating condi− tions, with simultaneous availability of heating and cooling during off season periods. 7.1.2.4

Single−path systems may be:

Air System Advantages

Heat Recovery

Air systems are readily adapted to heat recovery de− vices. 7.1.2.5

Design Freedom

Air systems allow full design freedom for optimum air distribution in air motion, draft control, and extenuat− ing local requirements. 7.1.2.6

Makeup Air

Air systems are best suited to applications requiring abnormal exhaust makeup. 7.1.2.7

Adaptable

Air systems are easily adaptable to automatic seasonal changeover and winter humidification. 7.1.3

Air System Disadvantages

Air systems have the following disadvantages: 7.1.3.1

Duct Space Requirements

Additional duct clearance requirements can penalize floor space for duct risers and fan rooms, and building


7.1

height for ceiling clearances. (Particularly true of low velocity systems). 7.1.3.2

Longer Fan Hours

In those systems which use air (not radiation) for pe− rimeter heating, longer fan−operating hours are re− quired to take care of unoccupied period heating (in low temperature locales). 7.1.3.3

System Balancing

In those systems which have no built−in zone self−bal− ancing devices, air balancing is difficult and may have to be done several times when a common air system serves areas which are not rented simultaneously. 7.1.3.4

Temporary Heat

Air heating perimeter systems may not be available for use during building construction as rapidly as perime− ter hydronic systems. 7.1.3.5

Terminal Devices

Accessibility to terminal devices demands close coop− eration between architectural, mechanical, and struc− tural designers.

ment stores, small individual shops in a shopping cen− ter, individual classrooms of schools, computer rooms, etc. A rooftop unit, for example, complete with refrig− eration system serving an individual space, would be considered a single zone system. The refrigeration sys− tem, however, could be remote and serve several single zone units in a larger installation. A schematic of a more sophisticated single zone cen− tral system is shown in Figure 7−2. The return air fan may be used if 100 percent outdoor air is used for cool− ing purposes, and may be eliminated if air is relieved from the space with very little pressure loss through the relief system. However, objectionable pressuriza− tion of conditioned spaces should be avoided to allow entrance doors to open or close normally. Control of the single zone system can be affected by varying the quantity of cooling medium, providing re− heat, face and bypass dampers or a combination of these. Single duct systems with reheat satisfy varia− tions in load by providing independent sources of heat− ing and cooling. When a humidifier is used, humidity control may be completely responsive to space needs. Single duct systems without reheat offer cooling flexi− bility, but cannot control summer humidity indepen− dent of temperature requirements. 7.2.2

7.1.3.6

Reheat Prohibition

Energy inefficiency of reheat type systems may pro− hibit use. 7.2

TYPES OF AIR SYSTEMS

7.2.1

Single Zone Systems

The simplest form of the air system is a single condi− tioner serving a single temperature control zone (see Figure 7−1). The unit may be installed within or remote from the space it serves and may operate with, or with− out distributing ductwork. Ideally, this can provide a system which is completely responsive to the needs of the space. Well designed systems can maintain tem− perature and humidity closely and efficiently. They can be shut down when desired, without affecting the operation of adjacent areas. A single zone system responds to only one set of space conditions. Its use is limited to situations where varia− tions occur almost uniformly throughout the zone or where the load is stable; but when multiple units are installed, they can handle a variety of conditions effi− ciently. Single zone systems are used in small depart− 7.2

Variable Air Volume (VAV) Systems

Control of dry bulb temperatures within a space re− quires that a balance be established between the space load and the air supplied to offset the load. The design− er may choose between varying the supply air temper− ature (constant volume) or varying the airflow volume (variable air volume) as the space load changes. To control part load volume reduction, supply air temper− atures and air volumes may be controlled simulta− neously. VAV systems (Figure 7−3), may be applied to interior or perimeter zones, with common or separate fan sys− tems, common or separate air temperature control, and with or without auxiliary heating devices. The VAV concept may apply to volume variation in the main system total airstream and to the zones of control. Variation of flow under control of a space thermostat may be accomplished by positioning a simple damper or a volume regulating device in a duct, in a VAV ter− minal box, or at a diffuser or grille. Depending on the complexity of the air distribution system, first cost considerations, the lowest throttling ratio expected at part load, and the complexibility of the initial and part load balancing problems, VAV may or may not be


POSSIBLE PREHEAT COIL

OUTDOOR AIR INTAKE

FILTERS

OUTDOOR DAMPER

COOLING COIL

HEATING COIL

SUPPLY AIR FAN RETURN AIR DAMPER

RELIEF AIR LOUVER

USE OF BAROMETRIC RELIEF AIR LOUVERS NOT RECOMMENDED

SUPPLY AIR SYSTEM

RELIEF AIR DAMPER RETURN AIR SYSTEM

FIGURE 7-1 SINGLE DUCT SYSTEM

EXHAUST AIR OR RELIEF AIR LOUVER & DAMPER RETURN AIR FAN

RETURN AIR DUCT REHEAT COIL IF BYPASS IS USED

RETURN AIR DAMPER BYPASS DAMPER SPRAYS

MIN. O.A. DAMPER MAX. O.A. DAMPER

SUPPLY AIR DUCT

FACE AND BYPASS DAMPER

SUPPLY AIR FAN OUTDOOR AIR LOUVER

MIXED AIR PLENUM

FILTERS PREHEAT COIL

COOLING COIL

REHEAT COIL IF BYPASS IS NOT USED

SPRAY PUMP

FIGURE 7-2 TYPICAL EQUIPMENT FOR SINGLE ZONE DUCT SYSTEM


7.3

OUTDOOR AIR INTAKE


FILTERS

HEATING COIL

COOLING COIL

PRIMARY AIR DUCT S.P. CONTROLLER

SUPPLY FAN WITH RETURN S.P. CONTROL AIR DAMPER

OUTDOOR DAMPER

VAV TERMINAL UNITS

EXHAUST AIR LOUVER

EXHAUST AIR DAMPER

OPTIONALRETURN AIR FAN RETURN AIR SYSTEM

T

T

FIGURE 7-3 VARIABLE AIR VOLUME (VAV) SYSTEM

combined with fan volume or system static pressure controls. It is possible to permit system airflow volume varia− tions without fan volume variation by using a simple fan bypass. It is possible to vary zone air volume only, while keeping fan and system volume substantially constant, by dumping excess air into a return air ceil− ing plenum or directly into the return air duct system. These methods of system control do not provide the fan horsepower savings usually associated with VAV systems. 7.2.3

Terminal Reheat Systems

The terminal reheat system (Figure 7−4) is a modifica− tion of the single zone system. It permits zone or space control for areas of unequal loading, provides heating or cooling of perimeter areas with different exposures, and promotes process or comfort applications where close control of space conditions is desired. As the word reheat implies, the application of heat is a secondary process being applied to either precondi− tioned primary air or recirculated room air. Under present energy codes, the use of Na reheat system is dis− couraged or prohibited unless recovered heat is used. A single low pressure reheat system is produced when a heating coil is inserted into the duct system down− stream of the cooling coil(s). The more sophisticated 7.4

systems use higher pressure duct designs and pressure reduction devices to permit system balancing at the re− heat zone. The medium for heating may be hot water, steam, or electricity. A big advantage of the reheat sys− tem is that it has the capability of maintaining very close control of space humidity. The system is generally applied to hospitals, laborato− ries, or spaces where wide load variations are expec− ted. Terminal units are designed to permit heating of primary air, or secondary air inducted from the condi− tioned space, located either under the window or in the duct system overhead. Conditioned air is supplied from a central unit at a fixed cold air temperature de− signed to offset the maximum cooling load in the space(s). The control thermostat simply calls for heat as the cooling load in the space drops below maxi− mum. 7.2.4

Induction Reheat Systems

The induction reheat system is shown schematically in Figure 7−5. Full cooling capacity is provided in the pri− mary higher pressure airstream and supplied by the central equipment to the terminal. Zone control is ac− complished by heating the secondary or induced air− stream. This type of terminal is used when it is desir− able to introduce supply air to the space at a higher temperature, or permit higher space air movement without increasing the quantity of primary air over the amount of air required for cooling.


POSSIBLE OUTDOOR PREHEAT AIR INTAKE COIL

COOLING COIL

RETURN AIR DAMPER

EXHAUST AIR EXHAUST AIR DAMPER LOUVER

POSSIBLE RETURN AIR FAN

SUPPLY AIR SYSTEM

SUPPLY AIR FAN REHEAT COIL NO. 1

RETURN AIR SYSTEM

REHEAT COIL NO. 2

T

T SUPPLY DUCT TO ZONE 1

SUPPLY DUCT TO ZONE 2

FIGURE 7-4 TERMINAL REHEAT SYSTEM The primary airN is discharged from nozzles arranged to induce room air into the induction unit approximate− ly four times the volume of the primary air. The in− duced air is cooled or heated by a secondary water coil. The water coil may be supplied by a 2−pipe system where either chilled water or heated water is available, but not simultaneously; by a 3−pipe system where sep− arate supplies of hot or chilled water are continuously available and, after passing through the unit, are mixed into a common return; or by a 4−pipe system, where a supply and return of hot water and chilled water are both continuously available. Induction units generally are located under the win− dow to offset winter downdrafts. Overhead installa− tions are limited, since ductwork connections carrying induced air have limited static pressure available, thereby decreasing induction air volume and unit ca− pacity. When installed under the window, this unit has the advantage of providing gravity heating during off− hour operation, permitting shutdown of the air system. The primary supply air fan operates at high pressures requiring high horsepower input. When balancing, at− tempt to reduce the primary air volume and pressure to the minimum required to operate the induction ter− minal units under full load conditions. Induction unit nozzles may be worn through many years of cleaning and operation, resulting in increased primary air quantity at lower air velocities with lower induced air volumes. Check each induction unit and either repair the nozzles or replace them before at− tempting any balancing work on the system.

7.2.5

Variable Air Volume (VAV) Reheat Systems

The VAV concept, when applied to reheat systems, permits flow reduction as a first step in control, there− by suspending the application of heat until flow condi− tions reach a predetermined minimum. By proper application of VAV, the reheat system may be designed to permit initial and operating cost sav− ings. With air volume selected for maximum instanta− neous peak loads rather than the sum of all peaks, the total system air volume is reduced. Also, any addition− al system diversity, such as areas with intermittent loads (conference rooms, office equipment rooms, etc.) may be included in the total volume reduction. When air volume reduction is used as a first step in control, reheat is not applied until the minimum vol− ume is reached. This procedure reduces system operat− ing costs appreciably for summer and intermediate weather. 7.2.6

Dual Duct Systems

7.2.6.1

Low Velocity Systems

The low velocity dual duct system distributes condi− tioned air through two parallel ducts. One duct carries cold air and the other warm air, allowing heating or cooling at all times. In each conditioned space or zone, automatic control dampers, responsive to a room ther− mostat, mix the warm and cold air in proper propor− tions to satisfy the prevailing heat load of the space. The return air fan shown in Figure 7−6 may be elimi− nated on small installations if provisions are made to


7.5

OUTDOOR AIR INTAKE


FILTERS

HEATING COIL

COOLING COIL

HIGH VELOCITY PRIMARY AIR SYSTEM

T

RETURN SUPPLY FAN AIR DAMPER

OUTDOOR AIR DAMPER

T

INDUCED AIR POSSIBLE EXHAUST AIR EXHAUST AIR DAMPER RETURN LOUVER AIR FAN

SECONDARY WATER

INDUCTION

FIGURE 7-5 INDUCTION REHEAT SYSTEM relieve excess outdoor air from the conditioned spaces. They generally are required for economizer cooling cycles and in systems with substantial return air ductwork. 7.2.6.2

High Velocity Systems

High velocity dual duct systems operate in the same manner as the low velocity systems except that the supply fan runs at a much higher pressure and each zone requires a mixing box with sound attenuation. A large amount of energy is required to operate the fan at high pressure. When balancing, a close analysis of the pressure drops within the duct system should be made and the fan pressure reduced to the minimum re− quired to operate the mixing boxes. 7.2.6.3

Energy Savings Ideas

In conditions when there is no cooling load, install controls to close off the cold air duct; de−energize chillers and cold water pumps and operate as a single duct system, rescheduling the warmer air duct temper− ature according to heating loads only. Under conditions where there is no heating load, install controls to close off the warm air ducts; shut off hot water, steam, or electricity to the warm duct and operate the system with the cold duct air only; resched− uling supply air temperature according to cooling loads. 7.6

Replace obsolete or defective mixing boxes to elimi− nate leakage of hot or cold air when the respective damper is closed. Provide volume control for the supply air fan and re− duce capacity preferably by speed reduction when both the hot deck and cold deck air quantities can be reduced to meet peak loads. Reducing the heat loss and heat gain provides an opportunity to reduce the amount of air circulated. When there is more than one air handling unit in a dual air system, modify duct work, if possible, so that each unit supplies a separate zone to provide an opportunity to reduce hot and cold duct temperatures according to shifting loads. Change dual duct systems to VAV systems when ener− gy analysis is favorable and the payback in energy saved is sufficiently attractive by adding VAV boxes and fan control. 7.2.7

Multi-Zone Systems

The multi−zone system (Figure 7−7) is applicable for serving a relatively small number of zones from a single, central air handling unit. The requirements of the different zones are met by mixing cold and warm air through zone dampers at the central air handler in response to zone thermostats. The mixed conditioned air is distributed throughout the building by a system of single zone ducts. Either packaged HVAC units complete with all components or field fabricated HVAC components may be used. Return air is usually handled in a conventional manner.


EXHAUST AIR LOUVER

EXHAUST AIR DAMPER

RETURN AIR FAN (OPTIONAL)

SUPPLY AIR

RETURN AIR DAMPER

HEATING COIL

FAN

Ô Ô Ô

OUTDOOR AIR INTAKE

FILTERS

OUTDOOR AIR DAMPER

COOLING COIL RETURN AIR SYSTEM


T

ZONE 2

T

ZONE 1

MIXING BOXES SUPPLY DUCT

SUPPLY DUCT

SYSTEM (COLD)

SYSTEM (HOT)

FIGURE 7-6 DUAL DUCT HIGH VELOCITY SYSTEM Multi−zone systems are somewhat similar to dual duct systems. They can provide a smaller building with some of the advantages of dual duct systems at a lower first cost with a wide variety of packaged HVAC units, but are limited to handling smaller projects by multi− ple runs of single zone ducts. Most packaged HVAC units lack the control sophistication for comfort and operating economy that can be built into dual duct sys− tems.

Multi−zone systems may handle more than one room with a single duct. Multizone packaged HVAC equip− ment is usually limited to about 12 zones, while built− up systems may have as many as can be physically in− corporated into the layout.

7.2.7.1

VAV Terminal Devices

VAV may be applied to multi−zone systems with pack− aged or built−up systems having the necessary zone volume regulation and fan controls. However, it is sel− dom applied in this manner for entire distribution sys− tems except for TV studios and other critical noise lev− el applications. More often, a few select rooms in a zone may incorporate VAV terminal devices, when off−peak requirements permit this approach and air balancing considerations indicate there will be no problems from omission of fan control or static pres− sure regulation. 7.2.7.2

Zone Coils

Some multi−zone units have individual heating and cooling coils for each zone supply duct. These units


7.7

use less energy then units with common coils as the supply air is heated or cooled only to the degree re− quired to meet the zone load. 7.2.7.3

Install controls or adjust existing controls to give the minimum hot deck temperature and maximum cold deck temperature consistent with the loads of critical zones.

TAB Arrange the controls so that when all of the hot duct dampers are partially closed, the hot deck tempera− tures will progressively reduce until one or more zone dampers are partially closed; the cold duct tempera− ture will progressively increase until one or more of the zone dampers are fully opened.

Before starting testing and balancing work, analyze multi−zone systems carefully and treat each zone as a single zone system. Adjust air volumes and tempera− ture accordingly. 7.2.7.4

Energy Savings Ideas Install controls to shut off the fan and all heating con− trol valves during unoccupied periods in the cooling season, and shut off the cooling valve during unoccu− pied periods in the heating season.

Hot and cold deck dampers are often of poor quality and allow considerable air leakage even where fully closed. Modify these dampers to avoid leakage.

EXHAUST AIR LOUVER

EXHAUST AIR DAMPER

POSSIBLE RETURN AIR FAN RETURN AIR SYSTEM ZONE 4

SUPPLY AIR FAN

OUTDOOR AIR INTAKE OUTDOOR AIR DAMPER

FILTERS COOLING COIL

HEATING COIL ZONE MIXING DAMPERS

RETURN AIR DAMPER

ZONE 3 ZONE 2 ZONE 1 SUPPLY AIR SYSTEMS

POSSIBLE PRE-HEAT COIL

FIGURE 7-7 MULTI-ZONE SYSTEM

7.8


7.3

AIR SYSTEM DESIGN

7.3.1

Introduction

Knowing the basics of good air system duct design will allow the TAB technician to determine if HVAC sys− tems can be balanced properly, in addition to helping solve routine problems while balancing. This section contains highlights only of duct design basics found in the SMACNA HVAC Systems9Duct Design manual. The tables and charts found in the Appendix of this manual are not as extensive as those found in the duct design manual, but they should be adequate for all TAB work. Air systems may be designed at higher or lower pres− sures. Higher friction rates and system pressures are required with higher velocity systems to reduce duct sizes and save space. For some lower velocity systems, higher pressures may be desirable for ease of balanc− ing and for flow control regulators that have a substan− tial pressure drop. On any of the variable air flow systems, there will be an infinite set of operating conditions which will create rates of airflow in the ducts entirely different from those used in the design. Every effort should be made when designing the ductwork to reduce the total fan power needed. This will assure a quieter system, reduced duct leakages, and in most cases maximum operating economy for the system owner. High velocity duct systems have been used mainly be− cause of space limitations created by architectural and structural practices. On the other hand where space is not at a premium, the use of high velocities and pres− sures are not economical. Some installations have areas of great space restriction where duct velocities must be higher. As soon as these points are passed and space becomes less critical, ve− locity rates should be sharply dropped, then gradually reduced toward the end of the duct system. 7.3.2

Equal Friction Design Method

The equal friction method of duct sizing probably has been the most universally used means of sizing low pressure supply air, return air and exhaust air duct sys− tems. It also is being adapted by many HVAC system designers for use in medium pressure systems. It nor− mally has not been used for sizing high pressure sys− tem. This design method automatically reduces air ve− locities in the direction of the airflow, so that by using a reasonable initial airflow velocity the chances of

introducing airflow generated noise are reduced or eliminated. When noise is an important consideration, the system velocity may be readily checked at any point during the design. Then there is the opportunity to reduce ve− locity created noise by increasing duct size or adding sound attenuation materials (such as duct lining). The major disadvantages of the equal friction method are there are no natural provisions for equalizing pres− sure drops in the branches (except in the few cases of a symmetrical layout); and there are no means of pro− viding the same static pressure behind each supply or return terminal device. Consequently, balancing can be difficult, even with a considerable amount of damp− ering in short duct runs. However, the equal friction method can be modified by designing portions of the longest run with different friction rates from those used for the shorter runs (or branches from the long run). Duct static regain (or loss) due to airflow velocity changes is included in the duct fitting pressure losses calculated using the duct fitting loss coefficient tables found in the Appendix. Otherwise, the omission of sys− tem static regain, when using older tables, could cause the calculated system fan static pressure to be greater than actual field conditions, particularly in larger, more complicated systems. Equal friction does not mean that total friction remains constant throughout the system. It means that a specific friction loss or stat− ic pressure loss per 100 feet (per meter) of duct is se− lected before the ductwork is laid out, and that this loss per 100 feet (per meter) is used constantly throughout the design. The SMACNA Duct Design System Calcu− lator makes this design method easy to use. 7.3.3

Supply Air Duct Sizing Procedures

To size the main supply air duct leaving the fan, the usual procedure is to select an initial velocity from Figures A−1 and A−2 found in the Appendix of this Manual. This velocity could be selected above the low velocity shaded section of the Duct Friction Loss Chart (I−P) A−1, and the Duct Friction Loss Chart (SI) A−2 if higher sound levels and energy conservation are not limiting factors. The charts on Appendix A.1 and A.2 are used to determine the friction loss by using the de− sign air quantity (cfm or L/s) and the selected velocity (fpm or m/s). A friction loss value commonly used for low pressure duct sizing is 0.1 in. wg per 100 feet of ductwork (0.8 Pa/m), although other values found in the low velocity shaded area, both lower and higher, are used by some designers as their standard or for spe−


7.9

cial applications. This same friction loss value gener− ally is maintained throughout the design, and the re− spective round duct diameters are obtained from the charts in Figures A−1 and A−2. The round duct diame− ters are used to select the equivalent rectangular duct sizes from chart A−1, unless round ductwork is to be used. Round ductwork is generally preferred on the higher pressure duct systems. The friction losses of each duct section may be cor− rected for otherN materials and construction methods by use of Table A−1 and Figure A−3. The correction factor from the Duct Friction Loss Correction Factors Figure A−3 is applied to the duct friction loss as only for the straight sections of the duct. The airflow rate used in the next section (and subsequent sections) of the main supply duct after each branch takeoff, is the original airflow rate (cfm or L/s) of the preceding sec− tion reduced by the amount of airflow into the branch. Using charts in Figures A−1 and A−2, the new airflow rate value (using the recommended friction rate of 0.1 in. wg per 100 ft (or 0.8 Pa/m) will determine the duct velocity and diameter for that section. The equivalent rectangular size of that duct section again is obtained from the Circulation Equivalents of Rectangular Ducts for Equal Friction and Capacity (I−P) (2) Dimen− sions in Inches Table A−2 and the (SI) version Table A−3 (if needed). All additional sections of the main supply air duct and all branch duct sections can be sized using charts in Figures A−1 and A−2 and approxi− mately the same friction loss rates. The pressure drop at each terminal device or air outlet (or inlet) of a small duct system, or of branch ducts of a larger system, should not differ more than 0.05 in. wg (12 Pa). If the pressure difference between the termi− nals exceeds that amount, dampering would be re− quired that could create objectionable air noise levels, and balancing may become more complicated. Example 7.1 (IP) A 70 foot section of N36 × 24 in. galvanized sheetN met− al duct is handling 10,000 cfm of air. What is the actual pressure loss and velocity of this duct section, and is it in the ?low velocity" category?

Solution Using Table A−2, a 36 × 24 in. duct has a circular equivalent of 32.0 inches. From Figure A−1, the 32.0 in. equivalent diameter duct has a velocity of 1800 fpm at a 10,000 cfm airflow rate and a pressure loss rate of 7.10

0.12 in. wg per 100 ft. It is in the low velocity shaded area. From Table A−1 the duct roughness category is medium smooth and from Figure A−3, a correction fac− tor is not needed. 0.12in.wg 70ft. Sectionpressureloss 100ft. 8 8 0.084in.wg NOTE: The pressure loss of any duct% fittings or ac− cessories contained in this 70 foot section of duct would be added to the above duct friction pressure loss.

Example 7.1 (SI) A 21 meter section of 900 × 600 mm galvanized sheet metal duct is handling 5000 L/s of air. What is the actu− al pressure loss and velocity of this duct section, and is it in the ?low velocity" category?

Solution Using Table A−3, a 900 × 600 mm duct has a circular equivalent of 799 mm. From Figure A−2, the 799 mm equivalent diameter duct has a velocity of 10 m/s at a 5000 L/s airflow rate, and a pressure loss rate of 1.1 Pa/ m. It is in the low velocity shaded area. From Table A−1, the duct roughness category is medium smooth, and from Figure A−3, a correction factor is not needed. Section pressure loss = 1.1 Pa/m × 21 m = 23.1 Pa 7.3.4

Modified Design Method

The modified equal friction method is used for sizing duct systems that are not symmetrical or that have both long and short runs. Instead of depending upon volume dampers to artificially increase the pressure drop of short branch runs, the branch ducts are sized (as nearly as possible) to dissipate (bleed off) the available pres− sure by using higher duct friction loss values. Only the main duct, which is usually the longest run, is sized by the original duct friction loss rate. Care should be exer− cised to prevent excessive high velocities in the short branches (with the higher friction rates). If calculated velocities are found to be too high, then duct sizes must be recalculated to yield lower velocities, and opposed blade volume dampers or static pressure plates must be installed in the branch duct at or near the main duct to dissipate the excess pressure. Regardless, it is good de− sign practice to always include balancing dampers in HVAC duct systems to balance the airflow to each branch.


Return Air Duct Systems

× 16"

A DAMPER

2000 cfm

Terminal Unit Ductwork

7.4

DUCT SIZING EXAMPLES

7.4.1

Sizing In I-P Units

Example 7.2 (I−P Units) What is the total pressure loss of the portion of the fi− brous glass HVAC duct system shown in Figure 7−8.

Solution Using the figures, tables, and charts from the Appen− dix:

B

× 16"

Low pressure ducts leaving terminal units are sized as any other conventional low pressure ductwork. Expe− rience indicates that the least expensive and the safest way to assure a quiet installation is to have some length of lined ductwork on the leaving side of terminal units. Lined ductwork, especially when it contains one or two elbows, can be a very effective sound attenuator. Lined ductwork helps if noise regeneration should oc− cur in the air distributing system because of poorly constructed ducts, fittings, and taps.

45’

35’

22"

7.3.6

32"

× 16"

Return air ducts should be sized using the equal fric− tion method at lower velocities. One scheme to simpli− fy return air systems is to use the space above hung ceilings or corridors for return air plenums and collect all return airflows at central points on each floor. In some localities there are code restrictions to using this method.

30’

32"

7.3.5

C

2000 cfm

25’

2000 cfm

D

14"

× 16"

×

28" 14" Grille PD=0.12 in. wg

FIGURE 7-8 SYSTEM LAYOUT (I-P UNITS)

Figure A−3, Correction factor = 1.43 Table A−14D, Opposed Blade Damper (set wide open), C = 0.52 Table A−10F , Sq. Elbow, 4.5 in. ?R", single thickness vanes, C = 0.23 Velocity = 6000 cfm/32 × 16/144 = 1688 fpm Table A−4, Vp = 0.18 in. wg

Duct AB: Duct ABC32 × 16 inches, 6000 cfm Table A−2, Circ. Equiv. = 24.4 in. diameter

DuctC45 ft + 30 ft = 75 ft/100 ft × 0.17 × 1.43 = 0.182 in. wg DamperCVp × C = 0.18 × 0.52 = 0.094 in. wg

Figure A−1, Friction loss = 0.17 in. wg/100 ft, velocity = 1850 fpm

ElbowCVp × C = 0.18 × 0.23′ = 0.041 in. wg

Table A−1, category = medium rough

Duct Section 1200 fpm

-0.72

-0.52

-0.23

0.34

0.76

1.14

1.83

For Main Loss Coefficient (C) see below.

G. CONVERGING TEE, ROUND BRANCH TO RECTANGULAR MAIN Use Vp of the downstream section When:

Ab/As

As/Ac

Ab/Ac

0.5

1.0

0.5

Branch, Coefficient C (See Note 4) Qb/Qc

Vc

0.1

0.2

0.3

0.4

0.5

0.6

0.7

< 1200 fpm

-0.63

-0.55

-0.13

0.23

0.78

1.30

1.93

> 1200 fpm

-0.49

-0.21

-0.23

0.60

1.27

2.06

2.75


H. CONVERGING TEE, RECTANGULAR MAIN AND BRANCH Use Vp of the downstream section When:

Ab/As

As/Ac

As/Ac

0.5

1.0

0.5

Vc

Branch, Coefficient C (See Note 4) Qb/Qc 0.1

0.2

0.3

0.4

0.5

0.6

0.7

< 1200 fpm

-0.75

-0.53

-0.03

0.33

1.03

1.10

2.15

> 1200 fpm

-0.69

-0.21

-0.23

0.67

1.17

1.66

2.67


Main Duct Loss Coefficient for Above Fittings Main Coefficeient C (See Note 4) 0.2 0.4 0.6 Vb / Vc C

0.03

0.04

0.07

0.8

1.0

1.2

1.4

1.6

1.8

0.12

0.13

0.14

0.27

0.30

0.25

Note 4: A = Area ( sq. in.). Q= airflow (cfm). V= Velocity (fpm)

Table A-12 Rectangular Branch Connection Loss Coefficients (continued) HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

A.21

I. CONVERGING WYE, RECTANGULAR Use Vp of the downstream section

Branch, Coefficient C (See Note 4)

R 1.0 W

Ab/As

Ab/Ac

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

Qb/Qc 0.1

0.2

0.3

0.4

0.5

0.6

0.7

-0 .50 -1 .2 -0 .50 -1 .0 -2 .2 -0 .60 -1 .2 -2 .1

0 -0 .40 -0 .20 -0 .60 -1 .5 -0 .30 -0 .80 -1 .4

0.50 0.40 0 -0 .20 -0 .95 -0 .10 -0 .40 -0 .90

1.2 1.6 0.25 0.10 -0 .50 -0 .04 -0 .20 -0 .5

2.2 3.0 0.45 0.30 0 0.13 0 - .20

3.7 4.8 0.70 0.60 0.40 0.21 0.16 0

5.8 6.8 1.0 1.0 0.80 0.29 0.24 0.20

Main, Coefficient C (See Note 4) AS/AC 0.75 1.0 0.75 0.5 1.0 0.75 0.5

Qb/Qc

Ab/AC

0.1

025 0.5 0.5 0.5 1.0 1.0 1.0

0.2

30 0.17 0.27 1.2 0.18 0.75 0.80

0.3

30 0.16 0.35 1.1 0.24 0.36 0.87

0.20 0.10 0.32 0.90 0.27 0.38 0.80

0.4

0.5

0.6

0.7

-0.10 0 0.25 0.65 0.26 0.35 0.68

-0.45 -0.08 0.12 0.35 0.23 0.27 0.55

-0.92 -0.18 -0.03 0 0.18 0.18 0.40

-1.5 -.027 -0.23 -0.40 0.10 0.05 0.25

J. TEE, RECTANGULAR MAIN TO CONICAL BRANCH Use Vp of the upstream section

Vb/Vc

0.40

C

0.80

Branch, Coefficient C (See note 4) 0.50 0.75 1.0 1.3 0.83

0.90

1.0

1.1

1.5 1.4

Main Duct Loss Coefficient for Above Fittings Main Coefficeient C (See Note 4) Vb / Vc 0.2 0.4 0.6 C

0.03

0.04

0.07

0.8

1.0

1.2

1.4

1.6

1.8

0.12

0.13

0.14

0.27

0.30

0.25

Note 4: A = Area ( sq. in.). Q= airflow (cfm). V= Velocity (fpm)

Table A-12 Rectangular Branch Connection Loss Coefficients (continued) A.22


A. TEE OR WYE, 30_ DEGREES TO 90_, ROUND Diverging FittingsCUse the Vp of the upstream section. Converging FittingsCUse the Vp of the downstream section. Wye = 30 Ab/Ac q

0.8 0.7 0.6 0.5 0.4 0.3 0.2 0.1

Branch, Coefficient C (See note 5) Qb/Qc 0.1

0.2

0.3

0.4

0.5

0.6

0.7

0.75 0.72 0.69 0.65 0.59 0.55 0.40 0.28

0.55 0.51 0.46 0.41 0.33 0.28 0.26 1.5

0.40 0.36 0.31 0.26 0.21 0.24 0.58 —

0.28 0.25 0.21 0.19 0.20 0.38 1.3 —

0.21 0.18 0.17 0.18 0.27 0.76 2.5 —

0.16 0.15 0.16 0.22 0.40 1.3 — —

0.15 0.16 0.20 0.32 0.62 2.0 — —

Wye = 45 As/Ac 0.8 0.7 0.6 0.5 0.4 0.3 0.2 0.1


0.2

0.3

0.4

0.5

0.6

0.7

0.78 0.77 0.74 0.71 0.66 0.66 0.56 0.60

0.62 0.59 0.56 0.52 0.47 0.48 0.56 2.1

0.49 0.47 0.44 0.41 0.40 0.52 1.0 —

0.40 0.38 0.37 0.38 0.43 0.73 1.8 —

0.34 0.34 0.35 0.40 0.54 1.2 — —

0.31 0.32 0.36 0.45 0.69 1.8 — —

0.32 0.35 0.43 0.59 0.95 2.7 — —

Wye = 60 Ab/Ac 0.8 0.7 0.6 0.5 0.4 0.3 0.2 0.1


0.2

0.3

0.4

0.5

0.6

0.7

0.83 0.82 0.81 0.79 0.76 0.80 0.77 1.0

0.71 0.69 0.68 0.66 0.65 0.75 0.96 2.9

0.62 0.61 0.60 0.61 0.65 0.89 1.6 —

0.56 0.56 0.58 0.62 0.74 1.2 2.5 —

0.52 0.54 0.58 0.68 0.89 1.8 — —

0.50 0.54 0.61 0.76 1.1 2.6 — —

0.53 0.60 0.72 0.94 1.4 3.5 — —

Wye = 90 Ab/Ac 0.8 0.7 0.6 0.5 0.4 0.3 0.2 0.1


0.2

0.3

0.4

0.5

0.6

0.7

0.95 0.95 0.96 0.97 0.99 1.1 1.3 2.1

0.92 0.94 0.97 1.0 1.1 1.4 1.9 —

0.92 0.95 1.0 1.1 1.3 1.8 2.9 —

0.93 0.98 1.1 1.2 1.5 2.3 — —

0.94 1.0 1.1 1.4 1.7 — — —

0.95 1.1 1.2 1.5 2.0 — — —

1.1 1.2 1.4 1.8 2.4 — — —

Note 5: A = Area (sq. in.). Q = Airflow (cfm). V = Velocity (fpm)

Table A-13 Round Branch Connection Loss Coefficients HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

A.23

B. 90_ CONICAL TEE, ROUND Diverging FittingsCUse the Vp of the upstream section. Converging FittingsCUse the Vp of the downstream section.

Vb/Vc

0

0.2

0.4

Branch, Coefficient C (See Note 5) 0.6 0.8 1.0 1.2 1.4

1.6

C

1.0

0.85

0.74

0.62

0.32

0.52

0.42

0.36

0.32


C. 45_ CONICAL WYE, ROUND Diverging FittingsCUse the Vp of the upstream section. Converging FittingsCUse the Vp of the downstream section.

Vb/Vc

0

0.2

0.4

0.6

C

1.0

0.84

0.61

0.41

Branch, Coefficient C (See Note 5) 0.8 1.0 1.2 1.4 0.27

0.17

0.12

0.12

1.6

1.8

2.0

0.14

0.18

0.27


Diverging Fitting Main Duct Loss Coefficients

Vs / Vc

0

0.1

0.2

C

0.35

0.28

0.22

Main, Coefficient C (See note 5) 0.3 0.4 0.5 0.17

0.13

0.09


Table A-13 Round Branch Connection Loss Coefficients (continued) A.24


0.6

0.8

0.06

0.02

D. CONVERGING WYE, ROUND Diverging FittingsCUse the Vp of the upstream section. Converging FittingsCUse the Vp of the downstream section.

Vb/Vc

45

Branch, Coefficient C (See Note 5) Ab/Ac 0.1

0.2

0.3

0.4

0.6

0.8

1.0

0.4

-0.56

-0.44

-0.35

-0.28

-0.15

-0.04

0.05

0.5

-0.48

-0.37

-0.28

-0.21

-0.09

0.02

0.11

0.6

-0.38

-0.27

-0.19

-0.12

0

0.10

0.18

0.7

-0.26

-0.16

-0.08

-0.01

0.10

0.20

0.28

0.8

-0.21

-0.02

0.05

0.12

0.23

0.32

0.40

0.9

0.04

0.13

0.21

0.27

0.37

0.46

053

1.0

0.22

0.31

0.38

0.44

0.53

0.62

0.69

1.5

1.4

1.5

1.5

1.6

1.7

1.7

1.8

2.0

3.1

3.2

3.2

3.2

3.3

3.3

3.3

2.5

5.3

5.3

5.3

5.4

5.4

5.4

5.4

3.0

8.0

8.0

8.0

8.0

8.0

8.0

8.0

Vb/Vc

Main Coefficient C Ab/Ac 0.1

0.2

0.3

0.4

0.6

0.8

1.0

0.1

-8.6

-4.1

-2.5

-1.7

-0.97

-0.58

-0.34

0.2

-6.7

-3.1

-1.9

-1.3

-0.67

-0.36

-0.18

0.3

-5.0

-2.2

-1.3

-0.88

-0.42

-0.19

-0.05

0.4

-3.5

-1.5

-0.88

-0.55

-0.21

-0.05

0.05

0.5

-2.3

-0.95

-0.51

-0.28

-0.06

0.06

0.13

0.6

-1.3

-0.50

-0.22

-0.09

0.05

0.12

0.17

0.7

-0.63

-0.18

-0.03

0.04

0.12

0.16

0.18

0.8

-0.18

0.01

0.07

0.10

0.13

0.15

0.17

0.9

0.03

0.07

0.08

0.09

0.10

0.11

0.13

1.0

-0.01

0

0

0.10

0.02

0.04

0.05


Table A-13 Round Branch Connection Loss Coefficients (continued) HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

A.25

E. CONVERGING TEE, 90_, ROUND Diverging FittingsCUse the Vp of the upstream section. Converging FittingsCUse the Vp of the downstream section. Branch, Coefficient C (See Note 5) Ab/Ac

Ob/Qc

90°

0.1

0.2

0.3

0.4

0.6

0.8

0.1

0.40

-0 .37

-0 .51

-0 .46

-0 .50

-0 .51

0.2

3.8

0.72

0.17

-0 .02

-0 .14

-0 .18

0.3

9.2

2.3

1.0

0.44

0.21

0.11

0.4

16

4.3

2.1

0.94

0.54

0.40

0.5

26

6.8

3.2

1.1

0.66

0.49

0.6

37

9.7

4.7

1.6

0.92

0.69

0.7

43

13

6.3

2.1

1.2

0.88

0.8

65

17

7.9

2.7

1.5

1.1

0.9

82

21

9.7

3.4

1.8

1.2

1.0

101

26

4.0

2.1

1.4

12

F. CONVERGING WYE, CONICAL ROUND Diverging FittingsCUse the Vp of the upstream section. Converging FittingsCUse the Vp of the downstream section.

Ab

Main, Coefficient C (See Note 5) Qb/Qs

As/Ac

Ab/Ac

0.2

0.4

0.6

0.8

1.0

1.2

1.4

1.6

1.8

0.3

0.2 0.3

-2.4 -2.8

-0.01 -1.2

2.0 0.12

3.8 1.1

5.3 1.9

6.6 2.6

7.8 3.2

8.9 3.7

9.8 4.2

0.4

0.2 0.3 0.4

-1.2 -1.6 -1.8

0.93 -0.27 -0.72

2.8 0.81 0.07

4.5 1.7 0.66

5.9 2.4 1.1

7.2 3.0 1.5

8.4 3.6 1.8

9.5 4.1 2.1

10 4.5 2.3

0.5

0.2 0.3 0.4 0.5

-.046 -.094 -1.1 -1.2

1.5 0.25 -0.24 -0.38

3.3 1.2 0.42 0.18

4.9 2.0 0.92 0.58

6.4 2.7 1.3 0.88

7.7 3.3 1.6 1.1

8.8 3.8 1.9 1.3

9.9 4.2 2.1 1.5

11 4.7 2.3 1.6

0.6

0.2 0.3 0.4 0.5 0.6

-0.55 -1.1 -1.2 -1.3 -1.3

1.3 0 -0.48 -0.62 -0.69

3.1 0.88 0.10 -0.14 -0.26

4.4 1.6 0.54 0.21 0.04

6.1 2.3 0.89 0.47 0.26

7.4 2.8 1.2 0.68 0.42

6.6 3.3 1.4 0.85 0.57

9.6 3.7 1.6 0.99 0.66

11 4.1 1.8 1.1 0.75

0.8

0.2 0.3 0.4 0.5 0.6 0.7 0.8

0.06 -0.52 -0.67 -0.73 -0.75 -0.77 -0.78

1.8 0.35 -0.05 -0.19 -0.27 -0.31 -0.34

3.5 1.1 0.43 0.18 0.05 -0.02 -0.07

5.1 1.7 0.80 0.46 0.28 0.18 0.12

6.5 2.3 1.1 0.68 0.45 0.32 0.24

7.8 2.8 1.4 0.85 0.58 0.43 0.33

8.9 3.2 1.6 0.99 0.68 0.50 0.39

10 3.6 1.8 1.1 0.76 0.56 0.44

11 3.9 1.9 1.2 0.83 0.61 0.47

0.2 0.3 0.4 0.5 0.6 0.8 1.0

— — — — — — —

2.1 0.54 0.21 0.05 -0.02 -0.10 -0.14

3.7 1.2 0.62 0.37 0.23 0.11 0.05

5.2 1.8 0.96 0.60 0.42 0.24 0.16

6.6 2.3 1.2 0.79 0.55 0.33 0.23

7.8 2.7 1.5 0.93 0.66 0.39 0.27

9.0 3.1 1.7 1.1 0.73 0.43 0.29

11 3.7 2.0 1.2 0.80 0.46 0.30

11 3.7 2.0 1.2 0.85 0.47 0.30

Qb

Qs As 45°

Qc Ac

1.0

Converging Fitting Main Duct Loss Coefficients Main, Coefficient C (See note 5) Qb/Qc 0.1 0.2 0.3 0.4 0.5 0.6 0.7 C

0.16

0.27

0.38

0.46

0.53

0.57

0.59

0.8

0.9

0.60

0.59


Table A-13 Round Branch Connection Loss Coefficients (continued) A.26


A. DAMPER, BUTTERFLY, THIN PLATE, ROUND Use Vp of the upstream section

q

0

10

Coefficient C 20 30

40

50

60

C

0.20

0.52

1.5

11

29

108

4.5

0 is full open

B. DAMPER, BUTTERFLY, THIN PLATE, RECTANGULAR Use Vp of the upstream section

q

0

10

Coefficient C 20 30

40

50

60

C

0.04

0.33

1.2

9.0

26

70

3.3

0 is full open

C. DAMPER, RECTANGULAR, PARALLEL BLADES Use Vp of the upstream section

Coefficient C L/R

q Damper blades with crimped leaf edges and 1/4” metal damper frame

0.3 0.4 0.5 0.6 0.8 1.0 1.5

80°

70°

60°

50°

40°

30°

20°

10°

0° Fully open

116 152 188 245 284 361 576

32 38 45 45 55 65 102

14 16 18 21 22 24 28

9.0 9.0 9.0 9.0 9.0 10 10

5.0 6.0 6.0 5.4 5.4 5.4 5.4

2.3 2.4 2.4 2.4 2.5 2.6 2.7

1.4 1.5 1.5 1.5 1.5 1.6 1.6

0.79 0.85 0.92 0.92 0.92 1.0 1.0

0.52 0.52 0.52 0.52 0.52 0.52 0.52

NW L R 2(H W)

where: N W L R H

is number of damper blades is duct dimension parallel to blade axis is sum of damper blade lengths is perimeter of duct is duct dimension on perpendicular to blade axis

Table A-14 Miscellaneous Fitting Coefficients HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

A.27

D. DAMPER, RECTANGULAR, OPPOSED BLADES Use Vp of the upstream section

Coefficient C 80°

70°

60°

50°

40°

30°

20°

10°

0° Fully open

807 915 1045 1121 1299 1521 1654

284 332 377 411 495 547 677

73 100 122 148 188 245 361

21 28 33 38 54 65 107

9.0 11 13 14 18 21 28

4.1 5.0 5.4 6.0 6.6 7.3 9.0

2.1 2.2 2.3 2.3 2.4 2.7 3.2

0.85 0.92 1.0 1.0 1.1 1.2 1.4

0.52 0.52 0.52 0.52 0.52 0.52 0.52

L/R

H

q

0.3 0.4 0.5 0.6 0.8 1.0 1.5

where:

NW L R 2(H W)

Damper blades with crimped leaf edges and 1/4” metal damper frame

N W L R H

is number of damper blades is duct dimension parallel to blade axis is sum of damper blade lengths is perimeter of duct is duct dimension on perpendicular to blade axis

E. PERFORATED PLATE IN DUCT, THICK, ROUND AND RECTANGULAR Use Vp of the upstream section

Coefficient C n t/D

0.20

0.25

0.30

0.40

0.50

0.60

0.70

0.80

0.90

0.015

52

30

18

8.2

4.0

2.0

0.97

0.42

0.13

0.2

48

28

17

7.7

3.8

1.9

0.91

0.40

0.13

0.4

46

27

17

7.4

3.6

1.8

0.88

0.39

0.13

42 where:

24 15 t = plate thickness

6.6

3.2

1.6

0.80

0.36

0.13

0.6 t/D 0.015

n

Ap A

d

= duameter of perforated holes

n = free area ratio of plate Ap = total flow area of perforated plate A = area of duct

Table A-14 Miscellaneous Fitting Coefficients (continued) A.28


F. RECTANGULAR DUCT WITH 4-90_ MITERED ELLS TO AVOID OBSTRUCTION Use Vp of the upstream section

Coefficient C 1.0 L/H Ratio 0.5

1.5

2

Single Blade Turning Vanes

—

0.86

0.83

0.77

Double Blade Turning Vanes

—

1.85

2.84

2.91

“S” type Splitter Vanes

0.61

0.65

—

—

No Vanes - Up to 1200 fpm

0.88

5.26

6.92

7.56

No Vanes - Over 1200 fpm

1.26

6.22

8.82

9.24

Where:

W/H = 1.0 to 3.0 B

= 12” to 24”

G. RECTANGULAR DUCT, DEPRESSED TO AVOID AN OBSTRUCTION Use Vp of the upstream section

Coefficient C L/H W/H

12”

0.125

0.15

0.25

0.30

1.0

0.26

0.30

0.33

0.35

0.4

0.10

0.14

0.22

0.30

15

H. ROUND DUCT, DEPRESSED TO AVOID AN OBSTRUCTION Use Vp of the upstream section

30

12”

When: L/D = 0.33 C = 0.24

Table A-14 Miscellaneous Fitting Coefficients (continued) HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

A.29

I. EXIT, ABRUPT, ROUND AND RECTANGULAR, WITH OR WITHOUT A WALL Use Vp of the upstream section

C = 1.0 With Screen: Cs = 1 + (C from Table K) WALL (OPTIONAL)

J. DUCT MOUNTED IN WALL, ROUND AND RECTANGULAR Use Vp of the upstream section Coefficient C L/D t/D 0

0.02 0.05 Rectangular : D

2HW (H W)

0

0.002

0.01

0.05

0.2

0.5

1.0

0.50 0.50 0.50

0.57 0.51 0.50

0.68 0.52 0.50

0.80 0.55 0.50

0.92 0.66 0.50

1.0 0.72 0.50

1.0 0.72 0.50

With Screen or Perforated Plate: a. Sharp Edge (t/De 0.05): Cs = 1 + C1 b. Thick Edge (t/De 0.05): Cs = C + C1 where: Cs is new coefficient C is from above table C1 is from Table K (screen) or Table E (perforated plate)

K. SCREEN IN DUCT, ROUND AND RECTANGULAR Use Vp of the upstream section

n

0.30

0.40

0.50

0.55

C

6.2

3.0

1.7

1.3

Coefficient C 0.60 0.65 0.97

0.75

0.70

0.75

0.80

0.90

0.58

0.44

0.32

0.14

Table A-14 Miscellaneous Fitting Coefficients (continued) A.30


HVAC EQUATIONS (I-P) Air Equations (I-P) a.

V p

V Vp d P

= = = =

T

=

Velocity (fpm) Velocity Pressure (in. wg) Density (lb/cu ft) Absolute Static Pressure (in. Hg) (Barometric pressure + static pressure) Absolute Temperature (460 + F)

Q (sens.) = 1.08 × cfm × ∆t

Q Cp d ∆t

= = = =

Heat Flow (Btu/hr) Specific Heat (Btu/lb × F) Density (lb/cu ft) Temperature Difference (F)

Q (lat.) = 4750 × cfm × ∆W (lb.)

∆W =

Humidity Ratio (lb or gr H2O/lb dry air)

V 1096

d or for standard air (d = 0.075 lb/cu ft): V 4005 V p to solve for d: d 1.325 b.

Pb T

Q (sens.) = 60 × Cp × d × cfm × ∆t

or for standard air (Cp = 0.24 Btu/lb × F):

c.

Q (lat.) = 0.67 × cfm × ∆W (gr.) d.

Q (total) = 4.5 × cfm × ∆h

∆h =

Enthalpy Difference (Btu/lb dry air)

e.

Q = A × U × ∆t

A U

= =

Area of Surface (sq ft) Heat Transfer Coefficient (Btu/sq ft × hr × F)

f.

R 1 U

R

=

Sum of Thermal Resistance (sq ft × hr × F / Btu)

g.

P1

P V R T M

= = = = =

Absolute Pressure (lb/sq ft) Total Volume (cu ft) Gas Constant Absolute Temperature (460 + F = R) Mass (lb)

h.

TP = Vp + SP

V1 V P 2 2 RM T1 T2

V

4005

2

TP = Vp = SP =

Total Pressure (in. wg) Velocity Pressure (in. wg) Static Pressure (in. wg)

i.

Vp

j.

V Vm

V = Vm = d =

Velocity (fpm) Measured Velocity (fpm) Density (lb/cu ft)

k.

cfm = A × V

A

=

Area of Duct cross section (sq ft)

l.

TP = C × Vp

C

=

Duct Fitting Loss Coefficient

d(otherthanstandard) 0.075(d std.air)


A.31

Fan Equations (I-P) cfm 2 rpm a. rpm2 1 cfm 1

2

b.

P2 rpm rpm2 P1 1

c.

bhp 2 rpm rpm2 1 bhp 1

d.

rpm d2 rpm2 1 d1

e.

rpmfan Pitchdiam.motorpulley rpmmotor Pitchdiam.fanpulley

3

2

Pump Equations (I-P) gpm rpm a. gpm 2 rpm2 1 1 gpm 2 D2 b. gpm 1 D1 H2 rpm2 2 c. rpm H1 1

d. e. f.

cfm = rpm =

Cubic feet per minute Revolutions per minute

P

Static or Total Pressure (in. wg)

=

bhp =

Brake horsepower

d

Density (lb/cu ft)

=

rpm =

Revolutions per minute

D

=

Impeller diameter

H

=

Head (ft wg)

2

H2 D2 D1 h1 bhp 2 rpm rpm2 1 bhp 1 3 bhp 2 D2 D1 bhp 1

bhp =

Brake horsepower

3

Hydronic Equivalents (I-P) a.

One gallon water = 8.33 pounds

b.

Specific heat (Cp ) water = 1.00 Btu/lb ⋅ F (@68F)

c.

Specific heat (Cp ) water vapor = 0.45 Btu/lb ⋅ F (@68F)

d.

One ft of water = 0.433 psi

e.

One psi = 2.3 ft wg = 2.04 in. Hg

f.

One cu ft of water = 62.4 lb = 7.49 gal.

g.

One in. of mercury (Hg) = 13.6 in. wg = 1.13 ft. wg

h.

Atmospheric Pressure = 29.92 in. Hg = 14.696 psi

A.32


Hydronic Equations (I-P) a.

Q 500 gpm Dt

b.

DP 2 gpm gpm2 1 DP 1

c.

gpm

DP C

2

2

v

d. e. f.

gpm H Sp.Gr. 3960 gpm H Sp.Gr. whp bhp Ep 3960 E p whp 100 Ep (inpercent) bhp whp

g.

NPSHA P a Ps

h.

V h f L 2 D 2g

V2 P vp 2g

Electric Equations (I-P) a. b.

Q = gpm = ∆t =

Heat Flow (Btu/hr) Gallons per minute Temperature Difference (F)

∆P = Cv =

Pressure difference (F) Valve constant (dimensionless)

whp = gpm = bhp = H = Sp.Gr.= Ep =

Water horsepower Gallons per minute Brake horsepower Head (ft wg) Specific gravity (use 1.0 for water) Efficiency of pump

NPSHA Pa = Ps = g =

= Net positive suction head available Atm. pressure (use 34 ft wg) Pressure at pump centerline (ft wg) Gravity acceleration (32.2 ft/sec2)

h f L D V

Head Loss (ft) Friction factor (dimensionless) Length of pipe (ft) Internal diameter (ft) Velocity (ft/sec)

= = = = =

I E P.F. Eff. (Single Phase) 746 I E P.F. Eff. 1.73 Bhp (Three Phase) 746 Bhp

c.

E IR

d.

P EI

e.

F.L.Amps Voltage * ActualF.L.Amps ActualVoltage

NOTE: * refers to Nameplate ratings. Water Temperature F

60F

150F

200F

250F

300F

340F

Ft. head differential per inch Hg. differential

1.046

1.07

1.09

1.11

1.15

1.165

Table A-15 Converting Pressure In Inches of Mercury to Feet of Water at Various Water Temperatures


A.33

Altitude (ft)

Sea Level

Barometer (in. Hg)

29.92

28.86

(in. wg)

407.5

392.8

−40°

1.26

1.22

0°

1.15

40°

1000

2000

3000

4000

5000

6000

7000

8000

9000

10,000

27.82

26.82

25.84

24.90

23.98

23.09

22.22

21.39

20.58

378.6

365.0

351.7

338.9

326.4

314.3

302.1

291.1

280.1

1.17

1.13

1.09

1.05

1.01

0.97

0.93

0.90

0.87

1.11

1.07

1.03

0.99

0.95

0.91

0.89

0.85

0.82

0.79

1.06

1.02

0.99

0.95

0.92

0.88

0.85

0.82

0.79

0.76

0.73

Air Temp. °F

70°

1.00

0.96

0.93

0.89

0.86

0.83

0.80

0.77

0.74

0.71

0.69

100°

0.95

0.92

0.88

0.85

0.81

0.78

0.75

0.73

0.70

0.68

0.65

150°

0.87

0.84

0.81

0.78

0.75

0.72

0.69

0.67

0.65

0.62

0.60

200°

0.80

0.77

0.74

0.71

0.69

0.66

0.64

0.62

0.60

0.57

0.55

250°

0.75

0.72

0.70

0.67

0.64

0.62

0.60

0.58

0.56

0.58

0.51

300°

0.70

0.67

0.65

0.62

0.60

0.58

0.56

0.54

0.52

0.50

0.48

350°

0.65

0.62

0.60

0.58

0.56

0.54

0.52

0.51

0.49

0.47

0.45

400°

0.62

0.60

0.57

0.55

0.53

0.51

0.49

0.48

0.46

0.44

0.42

450°

0.58

0.56

0.54

0.52

0.50

0.48

0.46

0.45

0.43

0.42

0.40

500°

0.55

0.53

0.51

0.49

0.47

0.45

0.44

0.43

0.41

0.39

0.38

550°

0.53

0.51

0.49

0.47

0.45

0.44

0.42

0.41

0.39

0.38

0.36

600°

0.50

0.48

0.46

0.45

0.43

0.41

0.40

0.39

0.37

0.35

0.34

700°

0.46

0.44

0.43

0.41

0.39

0.38

0.37

0.35

0.34

0.33

0.32

800°

0.42

0.40

0.39

0.37

0.36

0.35

0.33

0.32

0.31

0.30

0.29

900°

0.39

0.37

0.36

0.35

0.33

0.32

0.31

0.30

0.29

0.28

0.27

1000°

0.36

0.35

0.33

0.32

0.31

0.30

0.29

0.28

0.27

0.26

0.25

Standard Air Density, Sea Level, 70°F = 0.0 75 lb/cu ft at 29. 92 in. Hg

Table A-16 Air Density Correction Factors (I-P)

A.34


HVAC EQUATIONS (SI) Air Equations (SI) a.

V 1.414

Vd

p

or for standard air (d = 1.204 kg/m3): V 1.66V p To solve for d: P d 3.48 b T b.

Q C p d Ls Dt

or for standard air (CP = 1.005 kJ/kg ⋅ C) Q(sens.) 1.23 Ls Dt (in watts)

V Vp d Pb

= = = =

T

=

Velocity (m/s) Velocity Pressure (pascals or Pa) Density (kg/m3) Absolute Static Pressure (kPa) (Barometric pressure + static pressure) Absolute Temp. (273 + C = K)

Q Cp d L/s

= = = =

Heat Flow (watts or kilowatts) Specific Heat (kJ/kg ⋅ C) Density (kg/m3) Airflow (liters per second)

Q(sens.) 1.23 m3s Dt (in kilowatts)

t = m3/s =

Temperature Difference (C) Airflow (cubic meters per second)

c.

Q(lat.) 3.0 Ls DW

W=

Humidity Ratio (g H2O/kg dry air)

d.

Q(totalheat) 1.20 Ls Dh

h =

Enthalpy Diff. (kJ/kg dry air)

e.

Q A U Dt

A U

= =

Area of Surface (m2) Heat Transfer Coefficient (W/m2 ⋅ C)

f.

R 1 U

R

=

Sum of Thermal Resistances (m2 ⋅ C/W)

g.

P 1V 1 PV 2 2 RM T1 T2

P V T R M

= = = = =

Absolute Pressure (kPa) Total Volume (m3) Absolute Temp. (273 + C = K) Gas Constant Mass (kg)

h.

TP V P SP

TP = Vp = SP =

Total Pressure (Pa) Velocity Pressure (Pa) Static Pressure (Pa)

V p d V 2 0.602V2 2 d(otherthanstandard) V V m 1.204(d std.air)

d = V = Vm =

Density (kg/m3) Velocity (m/s) Measured Velocity (m/s)

A C

Area of duct cross section (m2) Duct Fitting Loss Coefficient

i. j.

k.

Ls 1000 A V

l.

TP C VP

= =


A.35

Fan Equations (SI) Ls 2 m3s rads 2 revs 2 a. 3 2 Ls 1 m s 1 rads 1 revs 1

2

b.

rads 2 P2 P1 rads 1

c.

rads 2 kW 2 kW 1 rads 1

d.

rads 2 d2 d1 rads 1

e.

rads(fan) Pitchdiam.motorpulley Pitch.diam.fanpulley rads(motor)

3

2

Pump Equations (SI) Ls 2 m3s rads 2 revs 2 a. 3 2 Ls 1 m s 1 rads 1 revs 1 m 3s 2 d2 b. d1 m 3s 1

c.

rads 2 H2 H1 rads 1

d.

H2 D2 H1 D1

e.

rads 2 BP 2 BP 1 rads 1

2

2

L/s = Liters per second m3/s = Cubic meters per second rad/s = Radians per second rev/s = Revolutions per second P = Static or Total Pressure (Pa) kW = Kilowatts d = Density (kg/m3) NOTE: m3/h = Cubic meters per hour (is used in lieu of m3/s in some countries.)

L/s = Liters per second m3/s = Cubic meters per second rad/s = Radians per second rev/s = Revolutions per second D = Impeller diameter H = Head (kPa) BP. = Brake horsepower NOTE: m3/h = Cubic meters per hour (is used in lieu of m3/s in some countries.)

3

Hydronic Equations (SI) a.

Q 4190 m3s Dt

2

b.

m3s 2 DP 2 DP 1 m3s 1

c.

m 3s Ls DP Cv Cv

2

Q

Ls 2 Ls 1

2

2

WP(kW) 9.81 m3s H(m) Sp.8 Gr. Ls H(Pa) Sp.8 Gr. or WP(W) 1002

=

Heat flow (kilowatts)

t = m3/s =

Temperature difference (C) Cubic meters per second (used for large volumes)

L/s =

Liters per second

P = Cv =

Pressure diff. (Pa or kPa) Valve constant (dimensionless)

d.

e.

BP WP Ep

f.

E p WP 100 (inpercent) BP

A.36

WP = Water power (kW or W) m3/s = Cubic meters per second H = Head (Pa or m) L/s = Liters per second Sp.Gr. = Specific gravity (use 1.0 for water) BP = Ep =

Brake power (kW) Efficiency of Pump


g.

NPSHA P a Ps V P vp 2g

NPSHA = Net positive suction head available Pa = Atm. press (Pa) (Std. Atm. press. = 101,325 Pa) Ps = Pressure at pump centerline (Pa) V 2 = Velocity head at point P (m) s 2g Pvp = Absolute vapor pressure (Pa)

h.

h f L V D 2g

h g f L D V

= = = = = =

Head loss (m) Gravity acceleration (9.807 m/s2) Friction factor (dimensionless) Length of pipe (m) Internal diameter (m) Velocity (m/s)

I E P.F. Eff. (Single Phase) 1000 I E P.F. Eff. 1.73 kW 1000 (Three Phase)

kW I E P.F. R P.

= = = = = =

Kilowatts Amps (A) Volts (V) Power factor ohms (Ω) watts (W)

a. b.

2

2

kW

c.

E IR

d.

P EI F.L.Amps * Voltage * ActualF.L.Amps ActualVoltage

e.

NOTE: * Refers to Nameplate Ratings


A.37

Altitude (m) Barometer (kPa) Air Temp °C 0° 20° 50° 75° 100° 125° 150° 175° 200° 225° 250° 275° 300° 325° 350° 375° 400° 425° 450° 475° 500° 525°

Sea Level 101.3

250 98.3

500 96.3

750 93.2

1000 90.2

1250 88.2

1500 85.1

1750 83.1

2000 80.0

2500 76.0

3000 71.9

1.08 1.00 0.91 0.85 0.79 0.74 0.70 0.66 0.62 0.59 0.56 0.54 0.51 0.49 0.47 0.46 0.44 0.42 0.41 0.39 0.38 0.37

1.05 0.97 0.89 0.82 0.77 0.72 0.68 0.64 0.61 0.58 0.55 0.52 0.50 0.48 0.46 0.44 0.43 0.41 0.40 0.38 0.37 0.36

1.02 0.95 0.86 0.80 0.75 0.70 0.66 0.62 0.59 0.56 0.53 0.51 0.49 0.47 0.45 0.43 0.41 0.40 0.38 0.37 0.36 0.35

0.99 0.92 0.84 0.78 0.72 0.68 0.64 0.62 0.57 0.54 0.52 0.49 0.47 0.45 0.43 0.42 0.40 0.39 0.37 0.36 0.35 0.34

0.96 0.89 0.81 0.75 0.70 0.66 0.62 0.59 0.56 0.53 0.50 0.48 0.46 0.44 0.42 0.41 0.39 0.38 0.36 0.35 0.34 0.33

0.93 0.87 0.79 0.73 0.68 0.64 0.60 0.57 0.54 0.51 0.49 0.47 0.45 0.43 0.41 0.39 0.38 0.37 0.35 0.34 0.33 0.32

0.91 0.84 0.77 0.71 0.66 0.62 0.59 0.55 0.52 0.50 0.47 0.45 0.43 0.41 0.40 0.38 0.37 0.35 0.34 0.33 0.32 0.31

0.88 0.82 0.75 0.69 0.65 0.60 0.57 0.54 0.51 0.48 0.46 0.44 0.42 0.40 0.39 0.37 0.36 0.34 0.33 0.32 0.31 0.30

0.86 0.79 0.72 0.67 0.63 0.59 0.55 0.52 0.49 0.47 0.45 0.43 0.41 0.39 0.38 0.36 0.35 0.33 0.32 0.31 0.30 0.29

0.81 0.75 0.68 0.63 0.59 0.55 0.52 0.44 0.47 0.44 0.42 0.40 0.38 0.37 0.35 0.34 0.33 0.32 0.31 0.29 0.28 0.27

0.76 0.71 0.64 0.60 0.56 0.52 0.49 0.46 0.44 0.42 0.40 0.38 0.36 0.35 0.33 0.32 0.31 0.30 0.29 0.28 0.27 0.26

Table A-17 Air Density Correction Factors (SI)

A.38


SI UNITS AND EQUIVALENTS

UNIT ampere candela Celsius coulomb farad henry hertz joule kelvin

SYMBOL A cd °C C F H Hz J K

QUANTITY Electric current Luminous intensity Temperature Electric charge Electric capacitance Electric inductance Frequency Energy, work, heat Thermodynamic y temperature

kilogram liter lumens lux meter mole newton ohm pascal

kg L lm lx m mol N Ω Pa

Mass Liquid volume Luminous flux Illuminance Length Amount of substance Force Electrical resistance Pressure, stress

radian second siemens steradian volt watt

rad s S sr v w

Plane angle Time Electric conductance Solid angle Electric potential Power, heat flow

EQUIVALENT OR RELATIONSHIP Same as I−P 1 cd/m2 = 0.292 ft. lamberts °F = 1.8 °C + 32° Same as I−P Same as I−P Same as I−P Same as cycles per second 1 J = 0.7376 ft−lb = 0.000948 Btu °K = °C + 273.15° °F 459.67 1.8 1 kg = 2.2046 lb 1 L = 1.056 qt = 0.264 gal 1 lm/m2 = 0.0929 ft candles 1 lx = 0.0929 ft candles 1 m = 3.281 ft C 1 N = kg • m/s2 = 0.2248 lb (force) Same as I−P 1 Pa = N/m2 = 0.000145 psi = 0.004022 in.wg 1 rad = 57.29° Same as I−P C C Same as I−P 1 W = J/s = 3.4122 Btu/hr 1 W = 0.000284 tons of refrigeration

Table A-18 SI Units (Basic and Derived)


A.39

QUANTITY acceleration angular velocity area atmospheric pressure density density, air density, water duct friction loss enthalpy gravity heat flow length (normal) linear velocity mass flow rate moment of inertia power pressure specific heatCair (Cp ) specific specific specific thermal

heatCair (Cv ) heatCwater volume conductivity

volume flow rate

SYMBOL m/s2 rad/s m2 C kg/m 3 C C Pa/m kJ/kg W m m/s kg/s kg⋅m2 W kPa Pa C C C m3/kg W mm/m2⋅°C m3/s L/s C C m3/h

UNIT meters per second squared radians per second square meter 101.325 kPa kilograms per cubic meter 1.2 kg/m3 1000 kg/m3 pascals per meter kilojoule per kilogram 9.8067 m/s2 watt meter meters per second kilograms per second kilograms × square meter watt kilopascal (1000 pascals) pascal 1000 J/kg °C 717 J/kg °C 4190 J/kg °C cubic meters per kilogram watt millimeter per square meter °C cubic meters per second liters per second 1 m3/s = 1000 L/s 1 mL = liters/1000 cubic meters per hour

I−P RELATIONSHIP 1 = 3.281 ft/sec2 1 rad/sec = 9.549 rpm = 0.159 rps 1 m2 = 10.76 ft2 29.92 in Hg = 14.696 psi 1 kg/m3 = 0.0624 lb/ft3 0.075 lb/ft3 62.4 lb/ft3 1 Pa/m = 0.1224 in.wg/100 ft 1 kJ/kg = 0.4299 Btu/lb dry air 32.2 ft/sec2 1 W = 3.412 Btu/hr 1 m = 3.281 ft = 39.37 in. 1 m/s = 196.9 fpm 1 kg/s = 7936.6 lb/hr 1 kg⋅m2 = 23.73 lb ft2 1 W = 0.00134 hp 1 kPa = 0.296 in Hg = 0.145 psi 1 Pa = 0.004015 in.wg 1000 J/kg °C = 1 kJ/kg °C = 0.2388 Btu/lb °F 0.17 Btu/lb °F 1.0 Btu/lb °F 1 m3/kg = 16.019 ft3/lb 1 W mm/m2 ⋅°C = 0.0069 Btuh in/ft2 °F 1 m3/s = 2118.88 cfm (air) 1 L/s = 2.12 cfm (air) 1 m3/s = 15,850 gpm (water) 1 mL/s = 1.05 gph (water) 1 m3/h = 0.588 cfm (air) 1 m3/h = 4.4 gpm (water) m/s2

Table A-19 SI Equivalents

A.40


L w Lp 10 log10

4prQ A4 10.5dB

L p Lw 10 log10

4prQ A4 10.5dB

2

2

R Sa or R A 1a 1a

Lp = Sound pressure level, dB re 0.0002 microbar Lw = Sound power level, dB re 10−12 watt r = Distance from the sound source, feet A%orNSα= Total sabins in the room Q = Directivity factor R = Room constant α = Absorption coefficient of the surface treat− ment S = Surface area in square feet a = Average sabin absorption coefficient for the room D V f

= = =

Characteristic dimension, in. Velocity, fpm Octave band center frequency, Hz

Octave Band Sound Power Level = F G HindBre10 12watts

F G H

= = =

Function (Use with charts) Function (Use with charts) Function (Use with charts)

L WB LWD TL 10 log 10 S A

LWB = LWD = TL = S = A =

Sound power level breakout Sound power level in duct Transmission loss of duct wall Radiating surface area of duct wall (sqNft) Cross−sectional area of duct component (sqNft)

L w Lp 10 log10 A 10dB

Lw = Lp = A =

Sound power level that enters duct Sound pressure level in source room Opening in duct (sq ft.)

lc f

λ c f

Wavelength in feet Speed of sound, 1125 fps Cycles per second, hertz

L w 10 log W W ref

Lw = W = Wref =

Sound power level, dB Acoustic power output of noise source 10−12 watt

Lp = P = Pref =

Sound power level in dB rms sound pressure Ref. rms sound pressure

Strouhal Number N str

fD 5 V

= = =

dBre10 12watt dBre1013watt 10 L p 20 log P Pref

D

C1 C2 C ... n T1 T2 Tn

Cn = Tn =

Actual duration of exposure, hours Noise exposure limit

B f

RPM No.ofBlades 60

Bf =

Blade frequency

Lw Kw Q P

Est sound power level, dB re 10−12 watt Specific sound power level Volume flow rate, cfm Pressure, in. wg

L w Kw 10 log Q 20 log P

= = = =

Table A-20 Sound Design Equations


A.41

NR TL 10 log S 10 log A

TL = S A

= =

The transmission loss of the partition separating the two spaces The surface area of the partition (sq ft) Total of the sound absorptive materials in the receiving room (sq ft sabins)

TL 20 log M 20 log F 33dB

M = F =

Mass of construction (lb/sq ft) Frequency (hertz)

D 0.5 A

D A

Distance (in feet) from the noise source Total absorption in room (sq ft sabins)

A2 A1 Noise Reduction in dB

NR, dB 10 log NR =

L P1 LP2 20 log D 1 20 log D 2

= =

A1 =

Total absorption (Sabins, sq ft) in the room before adding the sound absorption

LP1 = LP2 = D1 =

Sound pressure level at position #1 Sound pressure level at position #2 Distance (in ft) from noise source to position #1 Distance (in ft) from noise source to position #2

D2 = L p Lw 20 log D 0.5dB

D

a (AntilogL a20)105

a = La =

Acceleration in meters/sec2 (g = 9.8 m/s2) Acceleration level in dB re 10−5 (m/s2)

v (AntilogL v20)108

v

Velocity in meters/sec (1 g@ 100 Hz = 0.015 m/s) Velocity level in dB re 10−8 (m/s)

=

=

Lv = d (AntilogL d20)1011

d

=

Ld =

Distance (in ft) from the point source to the point where sound pressure is measured

Displacement in meters (1g@100 Hz = 0.0249 mm) Displacement level in dB re 10−11 (meters)

Table A-20 Sound Design Equations (continued)

A.42


FITTING EQUIVALENTS (WATER) Velocity, ft/s 1 2 3 4 5 6 7 8 9 10

1_w

3_r

1

11_r

I1_w

2

21_w

1.2 1.4 1.5 1.5 1.6 1.7 1.7 1.7 1.8 1.8

1.7 1.9 2.0 2.1 2.2 2.3 2.3 2.4 2.4 2.5

2.2 2.5 2.7 2.8 2.9 3.0 3.0 3.1 3.2 3.2

3.0 3.3 3.6 3.7 3.9 4.0 4.1 4.2 4.3 4.3

3.5 3.9 4.2 4.4 4.5 4.7 4.8 4.9 5.0 5.1

4.5 5.1 5.4 5.6 5.9 6.0 6.2 6.3 6.4 6.5

5.4 6.0 6.4 6.7 7.0 7.2 7.4 7.5 7.7 7.8

Pipe Size 3 31_w 6.7 7.5 8.0 8.3 8.7 8.9 9.1 9.3 9.5 9.7

7.7 8.6 9.2 9.6 10.0 10.3 10.5 10.8 11.0 11.2

4

5

6

8

10

12

8.6 9.5 10.2 10.6 11.1 11.4 11.7 11.9 12.2 12.4

10.5 11.7 12.5 13.1 13.6 14.0 14.3 14.6 14.9 15.2

12.2 13.7 14.6 15.2 15.8 16.3 16.7 17.1 17.4 17.7

15.4 17.3 18.4 19.2 19.8 20.5 21.0 21.5 21.9 22.2

18.7 20.8 22.3 23.2 24.2 24.9 25.5 26.1 26.6 27.0

22.2 24.8 26.5 27.6 28.8 29.6 30.3 31.0 31.6 32.0

Table A-21 Equivalent Length in Feet of Pipe for 90_ Elbows Pipe Size, mm

Velocity, m/s 0.33

15

20

25

32

40

50

65

75

100

125

150

200

250

300

0.4

0.5

0.7

0.9

1.1

1.4

1.6

2.0

2.6

3.2

3.7

4.7

5.7

6.8

0.67

0.4

0.6

0.8

1.0

1.2

1.5

1.8

2.3

2.9

3.6

4.2

5.3

6.3

7.6

1.00

0.5

0.6

0.8

1.1

1.3

1.6

1.9

2.5

3.1

3.8

4.5

5.6

6.8

8.0

1.33

0.5

0.6

0.8

1.1

1.3

1.7

2.0

2.5

3.2

4.0

4.6

5.8

7.1

8.4

1.67

0.5

0.7

0.9

1.2

1.4

1.8

2.1

2.6

3.4

4.1

4.8

6.0

7.4

8.8

2.00

0.5

0.7

0.9

1.2

1.4

1.8

2.2

2.7

3.5

4.3

5.0

6.2

7.6

9.0

2.33

0.5

0.7

0.9

1.2

1.5

1.9

2.2

2.8

3.6

4.4

5.1

6.4

7.8

9.2

2.67

0.5

0.7

0.9

1.3

1.5

1.9

2.3

2.8

3.6

4.5

5.2

6.5

8.0

9.4

3.00

0.5

0.7

0.9

1.3

1.5

1.9

2.3

2.9

3.7

4.5

5.3

6.7

8.1

9.6

3.33

0.5

0.8

0.9

1.3

1.5

1.9

2.4

3.0

3.8

4.6

5.4

6.8

8.2

9.8

Table A-22 Equivalent Length in Meters of Pipe for 90_ Elbows Iron Pipe

Fitting

Copper Tubing

Elbow. 90

1.0

1.0

Elbow, 45

0.7

0.7

Elbow. 90 long turn

0.5

0.5

Elbow. welded, 90

0.5

0.5

Reduced coupling

0.5

0.4

Open return bend

1.0

1.0

Angle Radiator valve

2.0

3.0

Radiator or convector

3.0

4.0

Boiler or heater

3.0

4.0

Open gate valve

0.5

0.7

Open globe valve

12.0

17.0

Table A-23 Iron and Copper Elbow Equivalents FIGURE A-7 ELBOW EQUIVALENTS OF TEES AT VARIOUS FLOW CONDITIONS


A.43

PROPERTIES OF STEAM

Pressure Gage psi Vacuum 25 in. Hg 9.56 in. Hg 0 2 5 15 50 100 150 200

Absolute psia

Saturation Temperature F

2.4 10 14.7 16.7 19.7 29.7 64.7 114.7 164.7 214.7

134 193 212 218 227 250 298 338 366 388

Specific Volume ft3/lb

Enthalpy Btu/lb

Liquid Vf

Steam Vg

Liquid hf

0.0163 0.0166

146.4 38.4 26.8 23.8 20.4 13.9 6.7 3.9 2.8 2.1

101 161 180 187 195 218 267 309 339 362

0.0167 0.0168 0.0168 0.0170 0.0174 0.0179 0.0182 0.0185

Evap. hfg

Steam hg

1018 982 970 966 961 946 912 881 857 837

1119 1143 1150 1153 1156 1164 1179 1190 1196 1179

Table A-24 Properties of Saturated Steam (I-P)

Specific Volume L/kg Pressure, kPa 19.9 47.4 101.00 199.00 362 618 1003 1555

Enthalpy, kJ/kg

Saturation Temperature °C

Liquid Vf

Steam Vg

Liquid hf

Evap. hfg

Steam hg

60 80 100 120 140 160 180 200

1.02 1.03 1.04 1.06 1.08 1.10 1.13 1.16

7669 3405 1672 891 508 307 194 127

251 335 419 504 589 676 763 852

2358 2308 2256 2202 2144 2082 2015 1941

2609 2643 2775 2706 2733 2758 2778 2793

Table A-25 Properties of Saturated Steam (SI)

A.44


STEAM PIPING (I-P)

Pressure DropCpsi per 100 Ft in Length

Nom. Pipe Size, in. 3_r 1 11_r 11_w

1_qy psi (1 oz)

1_i psi (2 oz)

1_r psi (4 oz)

1_w psi (8 oz)

3_r psi (12 oz)

1 psi

2 psi

Sat. press psig

Sat. press. psig

Sat. press psig

Sat. press psig

Sat. press. psig

Sat. press. psig

Sat. press. psig

3.5

12

3.5

12

3.5

12

3.5

12

3.5

12

3.5

12

3.5

12

9 17 36 56

11 21 45 70

14 26 53 84

16 31 66 100

20 37 78 120

24 46 96 147

29 54 111 174

35 66 138 210

36 68 140 218

43 82 170 260

42 81 162 246

50 95 200 304

60 114 232 360

73 137 280 430

2 21_w 3 31_w

108 174 318 462

134 215 380 550

162 258 465 670

194 310 550 800

234 378 660 990

285 460 810 1218

336 540 960 1410

410 660 1160 1700

420 680 1190 1740

510 820 1430 2100

480 780 1380 2000

590 950 1670 2420

710 1150 1950 2950

850 1370 2400 3450

4 5 6 8

640 1200 1920 3900

800 1430 2300 4800

950 1680 2820 5570

1160 2100 3350 7000

1410 2440 3960 8100

1690 3000 4850

1980 3570 5700

2400 4250 5700

2450 4380 7000

3000 5250 8600

2880 5100 8400

3460 6100

4200 7500

4900 8600

10,000

11,900

14,200

10,000

11,400

14,300

14,500

17,700

16,500

20,500

24,000

29,500

10 12

7200

8800

10,200

12,600

15,000

18,200

21,000

26,000

26,200

32,000

30,000

37,000

42,700

52,000

11.400

13,700

16,500

19,500

23,400

28,400

33,000

40,000

41,000

49,500

48,000

57,500

67,800

81,000

Table A-26 Steam Piping (I-P) Flow Rate of Steam in Schedule 40 Pipe at Initial Saturation Pressure of 3.5 and 12 psig (Flow Rate expressed in Pounds per Hour)

Pitch of Pipe Pipe Size, in.

1_r in.

1_w in.

1 1_w in.

1 in.

Capa− city

Max. Vel.

Capa− city

Max. Vel.

Capa− city

3.2 6.8 11.8 19.8 42.9

8 9 11 12 15

4.1 9.0 15.9 25.9 54.0

11 12 13 16 18

5.7 11.7 19.9 33.0 68.8

Max. Vel.

Capa− city

2 in.

Max. Vel.

Capa− city

3 in.

Max. Vel.

4 in.

Capa− city

Max. Vel.

Capa− city

8.3 17.3 31.3 46.8 99.6

17 22 25 26 32

9.9 19.2 33.4 50.8 102.4

5 in.

Max. Vel.

Capa− city

Max. Vel.

22 10.5 24 20.5 26 38.1 28 59.2 32 115.0

22 25 31 33 33

Capacity Expressed in Pounds Per Hour

3_r 1 11_r 11_w 2

13 15 17 19 24

6.4 12.8 24.6 37.4 83.3

14 17 20 22 27

7.1 14.8 27.0 42.0 92.9

16 19 22 24 30

Table A-27 Comparative Capacity of Steam Lines at Various Pitches for Steam and Condensate Flowing in Opposite Directions (Pitch of Pipe in Inches per 10 Feet – Velocity in Feet per Second) HVAC SYSTEMS Testing, Adjusting & Balancing • Third Edition

A.45

Initial Steam Pressure, psig

Total Pressure Drop in Steam Supply Piping, psi

Pressure Drop Per 100 Ft, psi

Subatmos, or vacuum return 0 1 2 5 10 15 30 50 100 150

2–4 oz. 1_w oz. 2 oz. 2 oz. 4 oz. 8 oz.

1–2 psi 1 oz. 1–4 oz. 8 oz. 11_w psi 3 psi

1 psi 2 psi 2–5 psi 2–5 psi 2–10 psi

4 psi 5–10 psi 10–15 psi 15–25 psi 25–30 psi

Table A-28 Pressure Drops In Common Use for Sizing Steam Pipe (For Corresponding Initial Steam Pressure)

Length in Feet to be Added to Run Size of Pipe Inches 1_w 3_r 1 1 1_r 1 1_w 2 2 1_w 3 3 1_w 4 5 6 8 10 12

Standard Elbow

Side Out− let Teeb

Gate Valve 2

Globe Valve 2

Angle Valve 2

1.3 1.8 2.2 3.0

3 4 5 6

0.3 0.4 0.5 0.6

14 18 23 29

7 10 12 15

3.5 4.3 5.0 6.5

7 8 11 13

0.8 1.0 1.1 1.4

34 46 54 66

18 22 27 34

8 9 11 13

15 18 22 27

1.6 1.9 2.2 2.8

80 92 112 136

40 45 56 67

17 21 27

35 45 53

3.7 4.6 5.5

180 230 270

92 112 132

Table A-29 Length in Feet of Pipe to be Added to Actual Length of Run — Owing to Fittings — to Obtain Equivalent Length

A.46


Capacity in Pounds per Hour Two Pipe Systems Nominal Pipe Size, Size Inches A

One−Pipe Systems

Condensate Flowing Against Steam Vertical

Horizontal

Ba

Ca

Supply Risers Up−Feed Db

Radiator Valves and Vertical Connections E

Radiator and Riser Runouts Fc

3_r 1 1 1_r 1 1_w 2

8 14 31 48 97

7 14 27 42 93

6 11 20 38 72

C 7 16 23 42

7 7 16 16 23

2 1_w 3 3 1_w 4 5

159 282 387 511 1,050

132 200 288 425 788

116 200 286 380 C

C C C C C

42 65 119 186 278

1,800 3,750 7,000 11,500 22,000

1,400 3,000 5,700 9,500 19,000

C C C C C

C C C C C

545 C C C

6 8 10 12 16

NOTE: Steam at an average pressure of 1 psig is used as a basis of calculating capacities. a Do

not use Column B for pressure drops of less than 1_qy ft. of equivalent run.

not use Column D for pressure drops of less than 1_wr psi per 100 ft. of equivalent run except on sizes 3 in. and over. b Do

c Pitch of horizontal runouts to risers and radiators should be not less than 1_w in. per ft.

Where this pitch cannot be obtained, runouts over 8 ft. in length should be one pipe size larger than called for in this table.

Table A-30 Steam Pipe Capacities for Low Pressure Systems (For Use on One-Pipe Systems or Two-Pipe Systems in which Condensate Flows Against the Steam Flow)


A.47

1_er psi or 1_w oz Drop per 100 ft Pipe Size Inch− es G

1_wr psi or 2_e oz Drop per 100 ft

1_qy psi or 1 oz Drop per 100 ft

Wet

Dry

Vac.

Wet

Dry

Vac.

H

I

J

K

L

M

1_i psi or 2 oz Drop per 100 ft

1_r psi or 4 oz Drop per 100 ft

Wet

Dry

Vac.

Wet

Dry

Vac.

Wet

Dry

Vac.

N

O

P

Q

R

S

T

U

V

Return Main 3_r

C

C

C

C

C

42

C

C

100

C

C

142

C

C

200

1

125

62

C

145

71

143

175

80

175

250

103

249

350

115

350

11_r

213

130

C

248

149

244

300

168

300

425

217

426

600

241

600

11_w

338

206

C

393

236

388

475

265

475

675

340

674

950

378

950

700

470

C

810

535

815

1,000

575

1,000

1,400

740

1,420

2,000

825

2,000

21_w

2

1,180

760

C

1,580

868

1,360

1,680

950

1,680

2,350

1,230

2,380

3,350

1,360

3,350

3

1,880

1,460

C

2,130

1,560

2,180

2,680

1,750

2,680

3,750

2,250

3,800

5,350

2,500

5,350

31_w

2,750

1,970

C

3,300

2,200

3,250

4,000

2,500

4,000

5,500

3,230

5,680

8,000

3,580

8,000

4

3,880

2,930

C

4,580

3,350

4,500

5,500

3,750

5,500

7,750

4,830

7,810 11,000

5,380 11,000

5

C

C

C

C

C

7,880

C

C

9,680

C

C 13,700

C

C 19,400

6

C

C

C

C

C 12,600

C

C 15,500

C

C 22,000

C

C 31,000

Table A-31 Return Main and Riser Capacities for Low-Pressure Systems—Pounds per Hour (Reference to this table will be made by column letter G through V)

A.48


STEAM PIPING (SI) Pressure Drop—Pa/m

Nom. Pipe Size mm 20 25 32 40

14 Pa/m

28 Pa/m

58 Pa/m

113 Pa/m

170 Pa/m

225 Pa/m

450 Pa/m

Sat. press. kPa

Sat. press. kPa

Sat. press. kPa

Sat. press. kPa

Sat. press. kPa

Sat. press. kPa

Sat. press. kPa

25 4 8 16 25

85 5 10 20 32

25 6 12 24 38

85 7 14 30 45

25 9 17 35 54

85 11 21 44 67

25 13 24 50 79

85 16 30 63 95

25 16 31 64 99

85 20 37 77 118

25 19 37 73 112

85 23 43 91 138

25 27 52 105 163

85 33 62 127 195

50 65 80 90

49 79 144 210

61 98 172 249

73 117 211 304

88 141 249 363

106 171 299 449

129 209 367 552

152 245 435 640

186 299 526 771

191 308 540 789

231 372 649 953

218 354 626 907

268 431 758 1100

322 522 885 1340

386 621 1090 1560

100 125 150 200

290 544 871 1770

363 649 1040 2180

431 762 1280 2530

526 953 1520 3180

640 1110 1800 3670

767 1360 2200 4540

898 1620 2590 5170

1090 1930 2590 6490

1110 1990 3180 6580

1360 2380 3900 8030

1310 2310 3910 7480

1570 2770 4540 9300

1910 3400 5400

2220 3900 6440

10,900

13,400

250 300

3270 5170

3990 6210

4630 7480

5720 8850

6800

8260

9530

11,800

11,900

14,500

13,600

16,800

19,400

23,600

10,600

12,900

15,000

18,100

18,600

22,500

21,800

26,100

30,800

36,000

Table A-32 Flow Rate in kg/h of Steam in Schedule 40 Pipe at Initial Saturation Pressure of 15 and 85 kPa Above Atmospheric Pitch of Pipe

20

40

80

120

170

250

350

420

Pipe Size mm

Capa− city kg/h

Max Vel. m/s

Capa− city kg/h

Max Vel. m/s

Capa− city kg/h

Max Vel. m/s

Capa− city kg/h

Max Vel. m/s

Capa− city kg/h

Max Vel. m/s

Capa− city kg/h

Max Vel. m/s

Capa− city kg/h

Max Vel. m/s

Capa− city kg/h

Max Vel. m/s

20

1.5

2.4

1.9

3.4

2.6

4.0

2.9

4.3

3.2

4.9

3.8

5.2

4.5

6.7

4.8

6.7

25

3.1

2.7

4.1

3.7

5.3

4.6

5.8

5.2

6.7

5.8

7.8

6.7

8.7

7.3

9.3

7.6

32

5.4

3.4

6.8

4.3

9.0

5.2

11.2

6.1

12.2

6.7

14.2

7.6

15.2

7.9

17.3

9.4

40

9.0

3.7

11.7

4.9

15.0

5.8

17.0

6.7

19.1

7.3

15.1

7.9

23.0

8.5

26.9

10.1

50

19.5

4.6

24.5

5.5

31.2

7.3

37.8

8.2

42.1

9.1

45.2

9.8

46.4

9.8

52.2

10.1

Capacity in kg/h, Velocity in m/s

Table A-33 Comparative Capacity of Steam Lines at Various Pitches for Steam and Condensate Flowing in Opposite Directions


A.49

Length in Meters to be Added to Run Size of Pipe mm 15 20 25 32

Standard Elbow 0.4 0.5 0.7 0.9

Side Outlet Teeb 1 1 1 2

Gate Valve 0.1 0.1 0.1 0.2

Globe Valve 4 5 7 9

Angle Valve a 2 3 4 5

40 50 65 80

1.1 1.3 1.5 1.9

2 2 3 4

0.2 0.3 0.3 0.4

10 14 16 20

6 8 8 10

100 125 150

2.7 3.3 4.0

5 7 8

0.5 0.7 0.9

28 34 41

14 17 20

200 250 300 350

5.2 6.4 8.2 9.1

11 14 16 19

1.1 1.4 1.7 1.9

55 70 82 94

28 34 40 46

NOTE:

a Valve

in full open position.

b Valve

given only to a tee used to divert the flow in the main to the last riser.

Table A-34 Equivalent Length of Fittings to be Added to Pipe Run

A.50


Capacity, kg/h Two−Pipe Systems Nominal Pipe Size, mm A

One−Pipe System

Condensate Flowing Against Steam

Supply Risers Up−feed Db

Radiator Valves and Vertical Connections E

Radiator and Riser Runouts Fc

Vertical

Horizontal

Ba

Ca

20 25 32 40 50

4 6 14 22 44

3 6 12 19 42

3 5 9 17 33

C 3 7 10 19

3 3 7 7 10

65 80 90 100 125

72 128 176 232 476

60 91 131 193 357

53 91 130 172 C

C C C C C

19 29 54 84 126

150 200 250 300 400

816 1700 3180 5220 9980

635 1360 2590 4310 8620

C C C C C

C C C C C

247 C C C C

NOTE: Steam at an average pressure of 7 kPa above atmospheric is used as a basis of calculating capacities. a Do

not use Column B for pressure drops of less than 13 Pa/m of equivalent run.

b Do not use Column D for pressure drops of less than 9 Pa/m of equivalent run except on sizes 88 mm and over. c Pitch of horizontal runouts to risers and radiators should not be less than 40 mm/m. Where this pitch cannot be obtained, runouts over 2.5 m in length should be one pipe size larger than called for in this table.

Table A-35 Steam Pipe Capacities for Low-Pressure Systems (For Use on One-Pipe Systems or Two-Pipe Systems in which Condensate Flows Against the Steam Flow)


A.51

7 Pa/m

9 Pa/m

14 Pa/m

28 Pa/m

57 Pa/m

Pipe Size, mm

Wet

Dry

Vac.

Wet

Dry

Vac.

Wet

Dry

Vac.

Wet

Dry

Vac.

Wet

Dry

Vac.

G

H

I

J

K

L

M

N

O

P

Q

R

S

T

U

V

20 25 32 40 50

C 57 97 153 318

C 28 59 93 213

C C C C C

C 66 112 178 367

C 32 68 107 243

19 65 111 176 370

Return Main C C 79 36 136 76 215 120 454 261

45 79 136 215 454

C 113 193 306 635

C 47 98 154 336

64 113 193 306 644

C 159 272 431 907

C 52 109 171 374

91 159 272 431 907

65 80 90 100 125 150

535 853 1,125 1,760 C C

345 662 894 1,330 C C

C C C C C C

717 967 1,500 2,080 C C

394 708 998 1,520 C C

616 989 1,400 2,040 3,570 5,720

762 1,220 1,810 2,490 4,390 7,030

1,070 1,700 2,490 3,520 C C

558 1,020 1,470 2,190 C C

1,080 1,720 2,580 3,540 6,210 9,980

762 1,220 1,810 2,490 C C

431 794 1,130 1,700 C C

1,520 617 1,520 2,430 1,130 2,430 3,630 1,620 3,630 4,990 2,440 4,990 C C 8,800 C C 14,100

Riser 20 25 32 40 50

C C C C C

22 51 112 170 340

C C C C C

C C C C C

22 51 112 170 340

65 111 176 370 616

C C C C C

22 51 112 170 340

79 136 215 454 762

C C C C C

22 51 112 170 340

113 193 306 644 1,080

C C C C C

65 70 80 100 125

C C C C C

C C C C C

C C C C C

C C C C C

C C C C C

989 1,470 2,030 3,570 5,720

C C C C C

C C C C C

1,220 1,810 2,490 4,390 7,030

C C C C C

C C C C C

1,720 2,580 3,540 6,210 9,980

C C C C C

Table A-36 Return Main and Riser Capacities for Low-Pressure Systems — kg/h

A.52


22 51 112 170 340

159 272 431 907 1,520

C 2,430 C 3,630 C 4,990 C 8,800 C 14,100

Sample Specification for Testing, Adjusting and Balancing The minimum requirements for testing, adjusting and balancing (TAB) of heating, ventilating and air conditioning (HVAC) distribution systems shall be as follows: 1.

The TAB Contractor shall review and become thoroughly familiar with the basic duct and piping installation layout prior to ceiling and wall installation. Prior to any closing−in of ductwork and piping, verify that all fittings, dampers, control devices, test devices and valves are properly located and installed.

2.

Examine each air and hydronic distribution system to see that is is free from obstructions. Confirm that all dampers, registers and valves are operating and in a set or full open position; that moving equipment is lubri− cated and functioning properly; and that the required filters are clean and installed. Request that the installing contractor perform any adjustments necessary for proper functioning of the system.

3.

The TAB Contractor shall use test instruments that have been calibrated within a time period recommended by the manufacturer or in the SMACNA HVAC SYSTEMS Testing, Adjusting and Balancing manual, and that they have been checked for accuracy prior to the start of the testing, adjusting and balancing activity.

4.

Verify that all equipment performs as specified. Adjust variable type drives, column dampers, control damp− ers, balancing valves and control valves as required by the TAB work.

5.

Adjust each register, diffuser and terminal unit to handle and properly distribute the design airflow within 10N% of the specified quantities.

6.

Adjust all balancing equipment so that each heating/cooling coil is furnished with the design fluid within 10N% of the specified quantities.

7.

Document the results of all testing on SMACNA TAB Report Forms and submit specified copies for approval and record.

8.

All TAB work shall be performed in accordance with the methods and the procedures described in the SMACNA HVAC SYSTEMS Testing, Adjusting and Balancing manual.

9.

The TAB Contractor shall become familiar with and comply with the provisions of all national, state, and local codes, ordinances, and safety acts that affect the TAB work.

10. HVAC systems utilizing microprocessor temperature controls, variable speed fans and pumps, and variable air terminal boxes should have the control system contractor on site during all primary system TAB work to provide any software programming or setpoint modifications required.


A.53

REFERENCES Publications from the following sources can be used as reference material to supplement the information con− tained in this manual. The members of the SMACNA Duct Design Committee express their thanks and ap− preciation to these firms for allowing selected text, fig− ures and tables to be used in this manual. 1.

Alnor Instrument Company

2. AMCA Fan Application Manuals C Air Move− ment and Control Association, Inc.

A.54

3. ASHRAE Handbooks C American Society of Heating, Refrigeration and Air Conditioning Engi− neers 4.

Bell and Gossett, Fluid Handling Division, ITT

5. Carrier System Design Manuals C Carrier Cor− poration 6. Fans and Their Application in Air Condition− ingCTrane Company 7. Products of Environmental Elements Corporation C Titus Corporation


GLOSSARY

GLOSSARY −A− A−Scale − A filtering system that has characteristics which roughly match the response characteristics of the human ear at low sound levels (below 55 dB Sound Pressure Level, but frequently used to gauge levels to 85 dB). A−scale measurements are often referred to as dB(A). Absolute Filter − Obsolete term (See HEPA filter) Absolute Pressure − Air at standard conditions 70F (20C ) air at sea level with a barometric pressure of 29.92 in.Hg. (760 mm Hg) exerts a pressure of 14.696 psi (101.325 kPa). This is the pressure in a system when the pressure gauge reads zero. So the absolute pressure of a system is the gauge pressure in pounds per square inch (kPa) added to the atmospheric pres− sure of 14.696 psi (use 14.7 psi in environmental sys− tem work) and the symbol is ?psia." Add 101.325 kPa to the gauge pressure for metric units. Absorbent − A material which, due to an affinity for certain substances, extracts one or more such sub− stances from a liquid or gaseous medium with which it contacts and which changes physically or chemical− ly, or both, during the process. Calcium chloride is an example of a solid absorbent, while solutions of lithi− um chloride, lithium bromide, and ethylene glycols are liquid absorbents.

Absorption Unit − Is a factory tested assembly of component parts producing refrigeration for comfort cooling by the application of heat. This definition shall apply to those absorption units which also produce comfort heating. Acceleration − The time rate of change of velocity, i.e., the derivative of velocity; with respect to time. Acceleration Due to Gravity − The rate of increase in velocity of a body falling freely in a vacuum. Its val− ue varies with latitude and elevation. The International Standard is 32.174 ft. per second per second (9.807 m/ s/s). Acceptance Test − A test made upon completion of fabrication, receipt, installation or modification of a component unit or system to verify that it meets the re− quirements specified. Accuracy − The extent to which the value of a quantity indicated by an instrument under test agrees with an accepted value of the quantity. Actuator − A controlled motor, relay or solenoid in which the electric energy is converted into a rotary, lin− ear, or switching action. An actuator can effect a change in the controlled variable by operating the final control elements a number of times. Valves and damp− ers are examples of mechanisms which can be con− trolled by actuators.

Absorber Surface − The surface of the collector plate which absorbs solar energy and transfers it to the col− lector plate.

Adiabatic Process − A thermodynamic process during which no heat is added to, or taken from, a substance or system.

Absorptance − The ratio of the amount of radiation ab− sorbed by a surface to the amount of radiation incident upon it.

Adjustable Differential − A means of changing the difference between the control cut−in and cutout points.

Absorption − A process whereby a material extracts one or more substances present in an atmosphere or mixture of gases or liquids accompanied by the materi− al’s physical and/or chemical changes.

Adsorbent − A material which has the ability to cause molecules of gases, liquids, or solids to adhere to its in− ternal surfaces without changing the adsorbent physi− cally or chemically. Certain solid materials, such as silica gel and activated alumina, have this properly.

Absorption Coefficient − For a surface, the ratio of the sound energy absorbed by a surface of a medium (or material) exposed to a sound field (or to sound radi− ation) divided by the sound energy incident on the sur− face. The conditions under which measurements of ab− sorption coefficients are made must be stated explicitly. The absorption coefficient is a function of both angle of incidence and frequency. Tables of ab− sorption coefficients usually list the absorption coeffi− cients at various frequencies, the values being those obtained by averaging over all angles of incidence.

Adsorption − The action, associated with the surface adherence, of a material in extracting one or more sub− stances present in an atmosphere or mixture of gases and liquids, unaccompanied by physical or chemical change. Commercial adsorbent materials have enor− mous internal surfaces. Aerodynamic Noise − Also called generated noise, self−generated noise; is a noise of aerodynamic origin in a moving fluid arising from flow instabilities. In


G.1

duct systems, aerodynamic noise is caused by airflow through elbows, dampers, branch wyes, pressure re− duction devices, silencers and other duct components. Air, Ambient − Generally speaking, the air surround− ing an object. Airborne Noise − Noise which reaches the observer by transmission through air. Air, Dry − Air without contained water vapor; air only. Air, Outdoor − Air taken from outdoors and, therefore, not previously circulated through the system. Air, Outside − External air, atmosphere exterior to re− frigerated or conditioned space; ambient (surround− ing) air. Air, Recirculated − Return air passed through the con− ditioner before being again supplied to the conditioned space. Air, Reheating of − In an air conditioning system, the final step in treatment, in the event the temperature is too low. Air, Return − Air returned from conditioned or refrig− erated space. Air, Saturated − Moist air in which the partial pressure of the water vapor is equal to the vapor pressure of wa− ter at the existing temperature. This occurs when dry air and saturated water vapor coexist at the same dry− bulb temperature. Air, Standard − Dry air at a pressure of 29.92 in.Hg (760 mm Hg) at 70F (20C) temperature and with a specific volume of 13.33 ft.3/lb (0.8305 m3/kg). Air Change Rate − The number of times the total air volume of a defined space is replaced in a given unit of time. Ordinarily computed by dividing the total vol− ume of the subject space (in cubic feet) into the total volume of air exhausted from the space per unit of time. Air Changes − A method of expressing the amount of air leakage into or out of a building or room in terms of the number of building volumes or room volumes exchanged. Air Conditioner, Unitary − An evaporator, compres− sor, and condenser combination; designed in one or G.2

more assemblies, the separate parts designed to be as− sembled together. Air Conditioning, Comfort − The process of treating air so as to control simultaneously its temperature, hu− midity, cleanliness and distribution to meet the com− fort requirements of the occupants of the conditioned space. Air Conditioning Unit − An assembly of equipment for the treatment of air so as to control, simultaneously, its temperature, humidity, cleanliness and distribution to meet the requirements of a conditioned space. Air Cooler − A factory−encased assembly of elements whereby the temperature of air passing through the de− vice is reduced. Air Diffuser − A circular, square, or rectangular air distribution outlet, generally located in the ceiling and comprised of deflecting members discharging supply air in various directions and planes, and arranged to promote mixing of primary air with secondary room air. Air Gap − An air gap in a potable water distribution system is the unobstructed vertical distance through the free atmosphere between the lowest opening from any pipe or faucet supplying water to a tank, plumbing fixture or other device and the floor level rim of the re− ceptacle. Air Shower − A relatively small, isolated ?chamber" normally located at the main entrance of a cleanroom to remove particulate from personnel and garments by high velocity air. Air, Supply − That air delivered to the conditioned space and used for ventilation, heating, cooling, hu− midification or dehumidification. Air, Transfer − The movement of indoor air from one space to another. Air, Ventilation − That portion of supply air which is outdoor air plus any recirculated air that has been treated for the purpose of maintaining acceptable in− door air quality. Air Washer − A water spray system or device for cleaning, humidifying, or dehumidifying the air. Airborne Sound − Sound which reaches the point of interest by radiation through the air. Airlock − An area between the entrance to the clean− room and the entry from an outside area. The airlock


receives the same clean, filtered air as the cleanroom, and is designed to prevent contaminated air in the out− side area from flowing into the cleanroom. (Also re− ferred to as ?Ante−Room.")

bulb temperature of the entering air. In a conduction heat exchanger device, the temperature difference be− tween the leaving treated fluid and the entering work− ing fluid.

Algae − A minute fresh water plant growth which forms a scum on the surfaces of recirculated water ap− paratus, interfering with fluid flow and heat transfer.

Area − Generally used to designate a portion of a build− ing at a given level of protection or contamination con− trol, as set off from adjoining portions of different con− tamination levels. Used somewhat interchangeably with ?space" or ?zone."

Alternating Current (AC) − A source of power for an electrical circuit which periodically reverses the po− larity of its charge. Ambient − The existing surrounding environmental conditions (Temperature, Relative Humidity, Pres− sure, etc...) of a particular area of consideration. Ampacity − A wire’s ability to carry current safely, without undue heating. The term formerly used to de− scribe this characteristic was current−capacity of the wire. Amperage − The flow of current in an electrical circuit measured in ?amperes," abbreviated ?amps" (A). Amplitude of Ground Surface Temperature Varia− tion − Peak Annual fluctuation of ground surface tem− perature about a mean value. Anemometer − An instrument for measuring the ve− locity of a fluid. Anemometer, Shielded Hot−Wire − An instrument for measuring air velocities based on the convective cool− ing effect of airflow on a heated wire. Instruments of this type are specifically designed for low air speeds, ranging from about 25 to 300 feet per minute (0.12 to 1.5 m/s). Anticipating Control − One which, by artificial means, is activated sooner than it would be without such means, to produce a smaller differential of the controlled property. Heat and cool anticipators are commonly used in thermostats. Anticipators − A small heater element in two−position temperature controllers which deliberately cause false indications of temperature in the controller in an at− tempt to minimize the override of the differential and smooth out the temperature variation in the controlled space. Approach − In an evaporative cooling device, the dif− ference between the average temperature of the circu− lating water leaving the device and the average wet−

As−Built Facility − A cleanroom which is complete and operating, with all services connected and func− tioning, but has no production equipment or operating personnel within the facility. As−Found Data − Data comparing the response of an instrument to known standards as determined without adjustment after the instrument is made operational. Aspect Ratio − In air distribution outlets, the ratio of the length of the core opening of a grille, face, or regis− ter to the width. In rectangular ducts, the ratio of the width to the depth. Aspiration − Production of movement in a fluid by suction created by fluid velocity. At−Rest Facility − A cleanroom which is complete and has the production equipment installed, but has no per− sonnel within the facility. Attenuation − The transmission loss or reduction in magnitude of a signal between two points in a trans− mission system. Autumnal Equinox (See Also Vernal Equinox) − The position of the sun midway between its lowest and highest altitude during the autumn; it occurs Septem− ber 21. Auxiliary Contacts − A set of contacts that perform a secondary function, usual in relation to the operation of a set of primary contacts. Averaging Element − A thermostat sensing element which will respond to the average duct temperature. Azimuth Angle (Solar) − The angular direction of the sun with respect to true south. −B− Backflow − The unintentional reversal of flow in a po− table water distribution system which may result in the transport of foreign materials or substances into the other branches of the distribution system.


G.3

Background Noise − Sound other than the signal wanted. In room acoustics, it is the irreducible noise level measured in the absence of any building occu− pants when all of the known sound sources have been turned off.

Bypass − A pipe or duct, usually controlled by valve or damper, for conveying a fluid around an element of a system.

Barometer − Instrument for measuring atmospheric pressure.

Calibration − Comparison of a measurement standard or instrument of unknown accuracy with another stan− dard or instrument of known accuracy to detect, corre− late, report, or eliminate by adjustment, any variation in the accuracy of the unknown standard or instrument.

Basic Principles − Essential theory and understanding of operation. Bimetallic Element − One formed of two metals hav− ing different coefficients of thermal expansion such as are used in temperature indicating and controlling de− vices. Boiling Point − The temperature at which the vapor pressure of a liquid equals the absolute external pres− sure of the liquid−vapor interface. Branch Circuit − Wiring between the last overcurrent device and the branch circuit outlets. Breakout Noise − The transmission or radiation of noise through some part of the duct system to an occu− pied space in the building. British Thermal Unit (Btu) − The Btu is defined as the heat required to raise the temperature of a pound of wa− ter from 59F to 60F. Btuh − Number of Btu’s transferred during a period of one hour. Bulb − The name given to the temperature sensing de− vice located in the fluid for which control or indication is provided. The bulb may be liquid−filled, gas−filled, or gas−and−liquid filled. Changes in temperature pro− duce pressure changes within the bulb which are trans− mitted to the controller. Building Envelope − The elements of a building which enclose conditioned spaces through which energy may be transferred to or from the exterior. Bus Bar − A heavy, rigid metallic conductor which car− ries a large current and makes a common connection between several circuits. Bus bars are usually uninsu− lated and located where the electrical service enters a building; that is, in the main distribution cabinet. Bus Duct − An assembly of heavy bars of copper or aluminum that acts as a conductor of large capacity. G.4

−C−

Calibration, Field − Calibration test performed in the field in accordance with the manufacturer’s recom− mended and/or accepted industry practices. Calibration, On−Line − Calibration performed using the reference system built into the instrument, in ac− cordance with manufacturer’s recommendations and/ or accepted industry standards. Capacitance − The property of an electric current that permits the storage of electrical energy in an electro− static field and the release of that energy at a later time. Capacitor (condenser) − An electrical device that will store an electric charge used to produce a power factor change. Capacity, Latent − The available refrigerating capac− ity of an air conditioner for removing latent heat from the space to be conditioned. Capillary − The name given to the thin tube attached to the bulb which transmits the bulb pressure changes to the controller or indicator. The cross sectional area of the capillary is extremely small compared to the cross section of the bulb so that the capillary, which is usually outside of the controlled fluid, will introduce the smallest possible error in the signal being trans− mitted from the bulb. Capillary Tube − The capillary tube is a metering de− vice made from a thin tube approximately 2 to 20 feet (0.6 to 6 m) long and from 0.025 to 0.090 inches (0.6 mm to 2.3 mm) in diameter which feeds liquid directly to the evaporator. Usually limited to systems of 1 ton or less, it performs all of the functions of the thermal expansion valve when properly sized. Cathodic Protection − The process of providing cor− rosion protection against electrolytic reactions that could be deleterious to the performance of the pro− tected material or component. Ceiling Outlet − A round, square, rectangular or linear air diffuser located in the ceiling which provides a hor−


izontal distribution pattern of primary and secondary air over the occupied zone and induces low velocity secondary air motion through the occupied zone. Celsius (Formerly Centigrade) − A thermometric scale in which the freezing point of water is called 0C and its boiling point 100C at normal atmospheric pressure (101.325 kPa). Certificate of Compliance (Conformance) − A writ− ten statement, signed by a qualified party, attesting that the items or services are in accordance with speci− fied requirements, and accompanied by additional in− formation to substantiate the statement. Certification − The process of validation required to obtain a certificate of compliance. Certification Agency − A company providing on−site, field certification services for profit or gain. Change of State − Change from one phase, such as sol− id, liquid or gas, to another. Changeover − The process of switching an air condi− tioning system from heating to cooling, or vice versa. Channel − Term used to describe output of a load man− agement system. Usually corresponds to a specific relay. Chemical Compatibility − The ability of materials and components in contact with each other to resist mutual chemical degradation, such as that caused by electrolytic action.

Clearing a Fault − Eliminating a fault condition by some means. Generally taken to mean operation of the over−circuit device that opens the circuit and clears the fault. Clo Value − A numerical representation of a clothing ensemble’s thermal resistance. 1Clo 0.88°F ft 2 hrBtu (0.155°C m 2W) Coanda Effect − The diversion of the normal fluid flow path from a jet by its attachment to an adjacent surface (wall or ceiling) caused by a low pressure re− gion between the fluid flow path and the surface. Coefficient of Discharge − For an air diffuser, the ratio of net area or effective area of vena contracta of an orificed airstream to the tree area of the opening. Coefficient of Expansion − The change in length per unit length or the change in volume per unit volume, per degree. change in temperature. Coefficient of Performance (COP), Heat Pump − The ratio of the compressor heating effect (heat pump) to the rate of energy input to the shaft of the compres− sor, in consistent units, in a complete heat pump, under designated operating conditions. Coil − A cooling or heating element made of pipe or tubing. Cold Deck − The cooling section of a mixed air zoning system. Collector Azimuth − The horizontal angle between true south and a line which is perpendicular to the plane of the collector that is projected on a horizontal plane.

Circuit − An electrical arrangement requiring a source of voltage, a closed loop of wiring, an electric load and some means for opening and closing it.

Collector Plate − The component of a solar collector which transfers the heat from solar energy to a circulat− ing fluid.

Circuit Breaker − A switch−type mechanism that opens automatically when it senses an overload (ex− cess current).

Collector (Solar) − An assembly of components in− tended to capture usable solar energy.

Cleanroom − A specially constructed room in which the air supply, air distribution, filtration of air supply, materials of construction, and operating procedures are regulated to control airborne particle concentra− tions to meet appropriate cleanliness levels as defined by Federal Standard 209E. Clean Zone − A defined space in which the concentra− tion of airborne particles is controlled to specified lim− its.

Combustion − The act or process of burning. Comfort Chart − A chart showing effective tempera− ture with dry−bulb temperatures and humidities (and sometimes air motion) by which the effects of various air conditions on human comfort may be compared. Comfort Cooling − Refrigeration for comfort as op− posed to refrigeration for storage or manufacture. Comfort Zone − (average) the range of effective tem− peratures over which the majority (50 percent or more)


G.5

of adults feels comfortable − (extreme) the range of ef− fective temperatures over which one or more adults feel comfortable. Commissioning Plan − A documented, systematic process to ready new building systems for active ser− vice. Common Neutral − A neutral conductor that is com− mon to, or serves, more than one circuit. Compressibility − The ease which a fluid may be re− duced in volume by the application of pressure, de− pends upon the state of the fluid as well as the type of fluid itself. In TAB work, consider that water may not be compressed. Air is a compressible gas, but that fac− tor is usually not considered during normal testing and balancing procedures. Compressor − The pump which provides the pressure differential to cause fluid to flow and in the pumping process increases pressure of the refrigerant to the high side condition. The compressor is the separation be− tween low side and high side. Concentration − The quantity of one constituent dis− persed in a defined amount of another. Concentrator − A reflective surface or refracting lens for directing insolation onto the absorber surface. Condensate − The liquid formed by condensation of a vapor. In steam heating, water condensed from steam; in air conditioning, water extracted from air, as by con− densation on the cooling coil of a refrigeration ma− chine. Condensation − Process of changing a vapor into liq− uid by extracting heat. Condensation of steam or water vapor is effected in either steam condensers or dehu− midifying coils, and the resulting water is called con− densate. Condenser − The heat exchanger in which the heat ab− sorbed by the evaporator and some of the heat of com− pression introduced by the compressor are removed from the system. The gaseous refrigerant changes to a liquid, again taking advantage of the relatively large heat transfer by the change of state in the condensing process. Condenser − ElectricalCsee ?capacitor". Condensing Unit, Refrigerant − An assembly of re− frigerating components designed to compress and liq− G.6

uify a specific refrigerant, consisting of one or more refrigerant compressors, refrigerant condensers, liq− uid receivers (when required) and regularly furnished accessories. Conditioned Space − Space within a building which is provided with heated and/or cooled air or surfaces and, where required, with humidification or dehumidifica− tion means so as to maintain a space condition falling within the ?comfort zone." Conditions, Standard − A set of physical, chemical, or other parameters of a substance or system which de− fines an accepted reference state or forms a basis for comparison. Conductance, Electrical − The reciprocal (opposite) of resistance and is the current carrying ability of any wire or electrical component. Resistance is the ability to oppose the flow of current. Conductance, Surface Film − Time rate of heat flow per unit area under steady conditions between a sur− face and a fluid for unit temperature difference be− tween the surface and fluid. Conductance, Thermal − Time rate of heat flow through a body (frequently per unit area) from one of its bounding surfaces to the other for a unit tempera− ture difference between the two surfaces, under steady conditions. Conductivity, Thermal − The time rate of heat flow through unit area and unit thickness of a homogeneous material under steady conditions when a unit tempera− ture gradient is maintained in the direction perpendic− ular to area. Materials are considered homogeneous when the value of the thermal conductivity is not af− fected by variation in thickness or in size of sample within the range normally used in construction. Conductor, Thermal − A material which readily transmits heat by means of conduction. Conduit − A round cross−section electrical raceway, of metal or plastic. Connected Load − The sum of all loads on a circuit. Connection in Parallel − System whereby flow is di− vided among two or more channels from a common starting point or header. Connection in Series − System whereby flow through two or more channels is in a single path entering each succeeding channel only after leaving the first or pre− vious channel.


Contamination − The presence of any unwanted sub− stance, material or energy which adversely affects a product or procedure in a cleanroom. Contactor − Electromagnetic switching device. Contaminant − An unwanted airborne constituent that may be reduced acceptability of the air. Control − A device for regulation of a system or com− ponent in normal operation, manual or automatic. If automatic, the implication is that it is responsive to changes of pressure, temperature or other property whose magnitude is to be regulated. Control Diagram (ladder diagram) − A diagram that shows the control scheme only. Power wiring is not shown. The control items are shown between two ver− tical lines; hence, the nameCladder diagram. Control Point − The value of the controlled variable which the controller operates to maintain. Controlled Area − An air conditioned work space or room in which the particle concentration is lower than normal air conditioned spaces. A controlled area is not to be classified as a cleanroom, but some special filtra− tion is required.

Cooling, Evaporative − Involves the adiabatic ex− change of heat between air and water spray or wetted surface. The water assumes the wet−bulb temperature of the air, which remains constant during its traverse of the exchanger. Cooling, Regenerative − Process of utilizing heat which must be rejected or absorbed in one part of the cycle to function usefully in another part of the cycle by heat transfer. Cooling Coil − An arrangement of pipe or tubing which transfers heat from air to a refrigerant or brine. Cooling Effect, Sensible − Difference between the to− tal cooling effect and the dehumidifying effect, usual− ly in watts (Btuh). Cooling Effect, Total − Difference between the total enthalpy of the dry air and water vapor mixture enter− ing the cooler per hour and the total enthalpy of the dry air and water vapor mixture leaving the cooler per hour, expressed in watts (Btuh). Cooling Range − In a water cooling device, the differ− ence between the average temperatures of the water entering and leaving the device. Core Area − The total plane area of that portion of a grille, included within lines tangent to the outer edges of the openings through which air can pass.

Controlled Device − One which receives the con− verted signal from the transmission system and trans− lates it into the appropriate action in the environmental system. For example − a valve opens or closes to regu− late fluid flow in the system.

Corresponding Values − Simultaneous values of vari− ous properties of a fluid, such as pressure, volume, temperature, etc., for a given condition of fluid.

Controller − An instrument which receives the signal from the sensing device and translates that signal into the appropriate corrective measure. The correction is then sent to the system controlled devices through the transmission system.

Counterflow − In heat exchange between two fluids, opposite direction of flow, coldest portion of one meet− ing coldest portion of the other.

Convection − Transfer of heat by movement of fluid. Convection, Forced − Convection resulting from forced circulation of a fluid, as by a fan, jet or pump. Convection, Natural − Circulation of gas or liquid (usually air or water) due to differences in density re− sulting from temperature changes. Conventional Flow (Nonlaminar Flow) Cleanroom − A cleanroom with non−uniform or mixed patterns and velocities.

Corrosive − Having chemically destructive effect on metals (occasionally on other materials).

Critical Surface − The surface of the work part to be protected from particulate contamination. Critical Velocity − The velocity above which fluid flow is turbulent. Cross Connection − Any physical connection or ar− rangement between two otherwise separate piping sys− tems, one of which contains potable water and the oth− er either water of unknown or questionable safety or steam, gas, chemicals, or other substances whereby there may be a flow from one system to the other, the direction of flow depending on the pressure differen− tial between the two systems.


G.7

Crossflow − Horizontal airflow. Crystal Formation, Zone of Maximum − Tempera− ture range in freezing in which most freezing takes place, i.e., about 25F to 30F (−4C to −1C) for wa− ter.

The referenced power for sound power level is 10−12 watts. In noise control work, the decibel notation is used to indicate the magnitude of sound pressure and sound power.

Curb Box − Access to an underground valve at the street curb. It controls water service to a house or building.

Combining Decibels − In sound survey work, it is fre− quently necessary to combine sound pressure level readings. An example would be to evaluate the effect of adding a noise source in a room where the noise lev− el is already considered borderline. Since the decibel scale is logarithmic, decibel values cannot be added directly. The correct procedure is to convert the deci− bels to intensity ratios, add the intensity ratios, and re− convert this sum into decibels.

Current (I) − The electric flow in an electric circuit, which is expressed in amperes (amps). Cycle − A complete course of operation of working fluid back to a starting point, measured in thermody− namic terms (functions). Also in general for any re− peated process on any system. Cycle, Reversible − Theoretical thermodynamic cycle, composed of a series of reversible processes, which can be completely reversed. −D− DWV − Drainage, waste and vent. Dalton’s Law of Partial Pressure − Each constituent of a mixture of gases behaves thermodynamically as if it alone occupied the space. The sum of the individu− al pressures of the constituents equals the total pres− sure of the mixture. Damper − A device to vary the volume of air passing through an airoutlet, air inlet or duct. Deadband − In HVAC, a temperature range in which neither heating nor cooling is turned on; in load man− agement, a kilowatt range in which loads are neither shed nor restored.

Degree Day − A unit, based upon temperature differ− ence and time, used in estimating fuel consumption and specifying nominal heating load of a building in winter. For any one day, when the mean temperature is less than 65F (18C), there exist as many degree days as there are Fahrenheit degrees difference in tem− perature between the mean temperature for the day and 65F (18C). Dehumidification − The condensation of water vapor from air by cooling below the dewpoint or removal of water vapor from air by chemical or physical methods. Dehumidifier − (1) an air cooler or washer used for lowering the moisture content of the air passing through it; (2) an absorption or adsorption device for removing moisture from air. Dehydration − (1) removal of water vapor from air by the use of absorbing or adsorbing materials, (2) remov− al of water from stored goods. Delta Service − An arrangement of the utility trans− formers. Commonly shown ?."

Decay Rate − The rate at which the sound pressure lev− el in an enclosed space decreases after the sound source has stopped. it is measured in decibels per sec− ond.

Demand − The probable maximum rate of water flow as determined by the number of water supply fixture units.

Decibel (dB) − The unit ?bel" is used in telecommu− nication engineering as a dimensionless unit for the logarithmic ratio of two power quantities. The decibel is one−tenth of a bel.

Demand Charge − The part of an electric bill based on kW demand and the demand interval, expressed in dol− lars per kilowatt. Demand charges offset construction and maintenance of a utility’s need for a large generat− ing capacity.

Therefore, L 10 log 10

G.8

sound power reference power

Demand Control − A device which controls the kW demand level by shedding loads when the kW demand exceeds a predetermined set point.


Demand Interval − The period of time during which kW demand is monitored by a utility service, usually 15 or 30 minutes long. Demand Load − The actual amount of load on a circuit at any time. The sum of all the loads which are ON. Equal to the connected load minus the loads that are OFF. Demand Reading − Highest or maximum demand for electricity an individual customer registers in a given interval, example, 15 minute interval. The metered de− mand reading sets the demand charge for the month. Density − The ratio of the mass of a specimen of a sub− stance to the volume of the specimen. The mass of a unit volume of a substance. When weight can be used without confusion, as synonymous with mass, density is the weight per unit volume. Desiccant − Any absorbent or adsorbent, liquid or sol− id, that will remove water or water vapor from a mate− rial. In a refrigeration circuit, the desiccant should be insoluble in the refrigerant. Design Working Pressure − The maximum allowable working pressure for which a specific part of a system is designed. Dewpoint, Apparatus − That temperature which would result if the psychrometric process occurring in a dehumidifier, humidifier or surface−cooler were car− ried to the saturation condition of the leaving air while maintaining the same ratio of sensible to total heat load in the process. Dewpoint Depression − The difference between dry bulb and dewpoint temperatures. Dewpoint Temperature − (tdp) The temperature at which moist air becomes saturated (100% relative hu− midity) with water vapor when cooled at constant pres− sure.

Diffuse Sound Field − A diffuse sound field is a space in which at every point the flow of sound energy in all directions is equally probable. (It is often assumed that in a diffuse field, the sound pressure level, averaged through time, is everywhere the same.) Diffuser − A circular, square, or rectangular air dis− tribution outlet, generally located in the ceiling and comprised of deflecting members discharging supply air in various directions and planes, and arranged to promote mixing of primary air with secondary room air. Direct Acting − Instruments that increase control pres− sure as the controlled variable (such as temperature or pressure) increases; while reverse acting instruments increase control pressure as the controlled variable de− creases. Direct Current (DC) − A source of power for an elec− trical circuit which does not reverse the polarity of its charge. Direct Field − The sound in a region in which all or most of the sound arrives directly from the source without reflection. Directivity Factor − The ratio of the sound pressure squared at some fixed distance and direction divided by the mean−squared sound pressure at the same dis− tance averaged over all directions from the source. Discharge Stop Valve − The manual service valve at the leaving connection of the compressor. Discrete Logic − Electronic circuitry composed of standard transistors, resistors, capacitors, etc., as compared to microprocessor circuits where the logic is condensed on a single chip (integrated circuit). Domestic Hot Water − Potable hot water as distin− guished from hot water used for house heating.

Dielectric Fitting − An insulating or nonconducting fitting used to isolate electrochemically dissimilar ma− terials.

D.O.P. (Dioctyl Phthalate) − An aerosol generated by blowing air through liquid dioctyl phthalate. Thermal− ly generated D.O.P. is an aerosol generated by con− densing vapor that has been evaporated from liquid (D.O.P.) by heat. The aerosol mean particle diameter is between 0.2 and 0.4 micron with a maximum geo− metric standard deviation of 1.3.

Differential − The difference between the points where a controller turns ?on" and ?off." If a thermostat turns a furnace on at 68F (20C) and the differential is 2F (1C), the burner will be turned off at 70F (21C).

D.O.P. Aerosol Generator, Air Operated − A device for producing a D.O.P. aerosol, operated by com− pressed air at room temperature, equipped with Laskin nozzles to produce a heterogeneous D.O.P. test aero− sol.


G.9

D.O.P. Aerosol Generator, Pressurized Gas−Ther− mal − A device for producing D.O.P. aerosol, operated by pressurized gas and equipped with heating means. Downflow − Vertical airflow (from ceiling to floor). Draft − a) A current of air, when referring to the pres− sure difference which causes a current of air or gases to flow through a flue, chimney, heater, or space; or b) to a localized effect caused by one or more factors of high air velocity, low ambient temperature, or direc− tion of air flow,whereby more heat is withdrawn from a person’s skin than is normally dissipated. Drier − A manufactured device containing a desiccant placed in the refrigerant circuit. Its primary purpose is to collect and hold within the desiccant, all water in the system in excess of the amount which can be tolerated in the circulating refrigerant. Drift − Term used to describe the difference between the set point and the actual operating or control point. Drip − A pipe, or a steam trap and a pipe considered as a unit, which conducts condensation from the steam side of a piping system to the water or return side of the system. Drop − The vertical distance that the lower edge of a horizontally projected airstream drops between the outlet and the end of its throw. Dry Bulb, Room − The dry bulb temperature of the conditioned room or space. Dry Bulb Temperature − The temperature of a gas or mixture of gases registered by an accurate thermome− ter after correction for radiation. The dry bulb repre− sents the measure of sensible heat, or the intensity of heat. Dry Bulb Temperature, Adjusted (tadb) − The average of the air temperature (ta) and the mean radiant tem− perature (tr) at a given location. The adjusted dry bulb temperature (tadb) is approximately equivalent to op− erative temperature (to) at air motions less than 80 fpm (0.4 m/s) when tr is less than 120F (49C).

Dynamic Discharge Head − Static discharge head plus friction head plus velocity head. Dynamic Insertion Loss − The dynamic insertion loss of a silencer, duct lining, or other attenuating device is in the performance measured in accordance with ASTM E 477 when handling the rated airflow. it is the reduction in sound pressure level, expressed in deci− bels, due solely to the placement of the sound attenuat− ing device in the duct system. Dynamic Suction Head − Positive static suction head minus friction head and minus velocity head. Dynamic Suction Lift − The sum of suction lift and ve− locity head at the pump suction when the source is be− low pump centerline. −E− Economizer − A system of dampers, temperature and humidity sensors, and motors which maximizes the use of outdoor air for cooling. Effect, Humidifying − Latent heat of water vaporiza− tion at the average evaporating temperature times the number of pounds (kilograms) of water evaporated per hour in Btuh (watts). Effect, Sun − Solar energy transmitted into space through windows and building materials. Effect, Total Cooling − The difference between the to− tal enthalpy of the dry air and water vapor mixture en− tering a unit per hour and the total enthalpy of the dry air and water vapor (and water) mixture leaving the unit per hour, expressed in Btu per hour (watts). Effective Area − The net area of an outlet or inlet de− vice through which air can pass, equal to the free area times the coefficient of discharge. Effectiveness (Efficiency) − The ratio of the actual amount of heat transferred by a heat recovery device to the maximum heat transfer possible between the air− streams (sensible heat/sensible heat, sensible heat/to− tal heat, or total heat/total heat).

Duct − A passageway made of sheet metal or other suit− able material, not necessarily leak tight, used for con− veying air or other gas at low pressures.

Elasticity of Demand − The change of quantity of electricity (or other commodity) purchased as a result of a change in its price. Demand for electricity is ?elas− tic" when it increases or decreases in response to de− creases or increases, respectively, in the price for the electricity.

Dust − An air suspension (aerosol) or particles of any solid material, usually with particle size less than 100 microns.

Electrical Circuit − A power supply, a load, and a path for current flow are the minimum requirements for an electrical circuit.

G.10


Electromechanical − Converting electrical input into mechanical action. A relay in an electromechanical switch. Electro−Pneumatic (EP) Switches − Switches that open or close an air line valve from an electrical im− pulse. Electrostatic Discharge (E.S.D.) − A transfer of elec− trostatic charge between objects at different electro− static potentials caused by direct contact or induced by electrostatic field. Emissivity − The property of a surface that determines its ability to give off radiant energy. Emittance − The ratio of the radiant energy emitted by a body to the energy emitted by a black body at the same temperature. End Reflection − When a duct system opens abruptly into a large room, some of the acoustic energy at the exit of the duct is reflected upstream with the result that the amount of the acoustic energy radiated into the room is reduced. This decrease in radiated energy in− creases as the frequency decreases. Energy − Expressed in kilowatt−hours (kWh) or watt− hours (Wh), and is equal to the product of power and time. energy = power × time kilowatt−hours = kilowatts × hours watt−hours = watts × hours

Enthalpy, Specific − A term sometimes applied to en− thalpy per unit weight. Entrainment − The capture of part of the surrounding air by the airstream discharged from an outlet (some− times called secondary air motion). Entropy − The ratio of the heat added to a substance to the absolute temperature at which it is added. Entropy, Specific − A term sometimes applied to en− tropy per unit weight. Equal Friction Method − A method of duct sizing wherein the selected duct friction loss value is used constantly throughout the design of a low pressure duct system. Equivalent Duct Diameter − The equivalent duct di− ameter for a rectangular duct with sides of dimensions a and b is 4/abp. Evaporation − Change of state from liquid to vapor. Evaporative Cooling − The adiabatic exchange of heat between air and a water spray or wetted surface. The water approaches the wet bulb temperature of the air, which remains constant during its traverse of the exchanger. Evaporator − The heat exchanger in which the me− dium being cooled, usually air or water, gives up heat to the refrigerant through the exchanger transfer sur− face. The liquid refrigerant boils into a gas in the pro− cess of the heat absorption. Exfiltration − Air leakage outward through cracks and interstices and through ceilings, floors and walls of a space or building.

Energy (Consumption) Charge − That part of an elec− tric bill based on kWh consumption (expressed in cents per kWh). Energy charge covers cost of utility fuel, general operating costs, and part of the amortiza− tion of the utility’s equipment.

Extended Surface − Heat transfer surface, one or both sides of which are increased in area by the addition of fins, discs, or other means.

Energy Efficiency Ratio (EER), Cooling − The ratio of net cooling capacity in Btuh to total electric input in watts under designated operating conditions.

Face Area − The total plane area of the portion of a grille, coil, or other items bounded by a line tangent to the outer edges of the openings through which air can pass.

Engine − Prime mover; device for transforming fuel or heat energy into mechanical energy.

Face Velocity − The velocity obtained by dividing the air quantity by the component face area.

Enthalpy − The total quantity of heat energy contained in a substance, also called total heat; the sum of the sensible heat and latent heat in an exchange process.

Fahrenheit − A thermometric scale in which 32 (F) denotes freezing and 212 (F) the boiling point of wa− ter under normal pressure at sea level (14.696 psi).

−F−


G.11

Fail Safe − In load management, returning all loads to conventional control during a power failure. Accom− plished by a relay whose contacts are normally closed. Fan, Centrifugal − A fan rotor or wheel within a scroll type housing and including driving mechanism sup− ports for either belt drive or direct connection. Fan performance Curve − Fan performance curve re− fers to the constant speed performance curve. This is a graphical presentation of static or total pressure and power input over a range of air volume flow rate at a stated inlet density and fan speed. It may include static and mechanical efficiency curves. The range of air volume flow rate which is covered generally extends from shutoff (zero air volume flow rate) to free deliv− ery (zero fan static pressure). The pressure curves are generally referred to as the pressure−volume curves. Fan propeller − A propeller or disc type wheel within a mounting or plate and including driving mechanism supports for either belt drive or direct connection. Fan, Tubeaxial − A propeller or disc type wheel within a cylinder and including driving mechanism supports for either belt drive or direct connection. Fan, Vaneaxial − A disc type wheel within a cylinder, a set of airguide vanes located either before or after the wheel and including driving mechanism supports for either belt drive or direct connection. Fault − A short circuitCeither line to line, or line to ground. Feed Line − A pipe that supplies water to items such as a boiler or a domestic hot water tank. Filter − A device to remove solid material from a fluid. Filter−Drier − A combination device used as a strainer and moisture remover.

If this ?first" room has the same noise criterion (NC) or a lower NC value than rooms further away from the fan, it may be assumed that, if the acoustical attenua− tion of the duct system from the fan to this ?first" room satisfies the requirements for this ?first" room, it also satisfies the acoustical requirements for rooms further away from the fan. Fire Damper − A device, installed in an air distribu− tion system, designed to close automatically upon detection of heat, to interrupt migratory airflow, and to restrict the passage of flame. A combination fire and smoke damper shall meet the requirements of both. Fire Resistance Rating − The time, in minutes or hours, that materials or assemblies have withstood a fire exposure as established in accordance with the test procedures of NFPA 251, Standard Methods of Fire Tests of Building Construction and Materials. Fire Wall − A wall having adequate fire resistance and structural stability under fire conditions to accomplish the purpose of subdividing buildings to restrict the spread of fire. First Air − The air which issues directly from the HEPA filter before it passes over any work location. First Work Location − The work location nearest the downstream side of the HEPA filters in a laminar air− flow device or cleanroom. Fixed Collector − A permanently oriented collector that has no provision for seasonal adjustment or track− ing of the sun. Flame Spread Rating − The flame spread rating of a material refers to a number or classification of materi− al obtained according to NFPA 255, Method of Test of Surface Burning Characteristics of Building Materi− als.

Fin − An extended surface to increase the heat transfer area, as metal sheets attached to tubes.

Flanking (Sound) Transmission − The transmission of sound between two rooms by any indirect path of sound transmission.

Final Filter − The last stage of filtration before the air− stream enters the clean space. The performance grade of this filter determines the air quality entering the clean space.

Flat−Plate Collector − A solar collector without exter− nal concentrators or focusing devices, usually consist− ing of an absorber plate, cover plates, back and side in− sulation and a container.

First Acoustically Critical Room − Most duct system service a number of rooms. The room that has the shortest duct run from the fan is usually exposed to more fan noise than rooms further away from the fan.

Floating Action Controllers − Essentially two posi− tion type controllers which vary the position of the controlled devices but which are arranged to stop be− fore reaching a maximum or minimum position.

G.12


Flow, Laminar − Fluid flow in which each fluid par− ticle moves in a smooth path substantially parallel to the paths followed by all other particles. Flow, Turbulent − Fluid flow in which the fluid moves transversely as well as in the direction of the tube or pipe axis. Flue − A special enclosure incorporated into a building for the removal of products of combustion to the out−of−doors. Type

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