Waste heat boiler deskbook

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Author: V Ganapathy

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Waste Heat Boiler Deskbook by v. Ganapathy

Library of Congress Cataloging-in-Publication Data Ganapathy, V. Waste heat boiler deskbook / by V. Gartapathy. p. cm. Includes Index. ISBN 0-88173-122-6 1. Waste heat boilers--Handbooks, manuals, etc. I. Title. TJ319.G36

1991

621.1'83--dc20

90-85871

elP

Waste Heat Boiler Deskbook / By V. Ganapathy. ©1991 by The Fairmont Press, Inc. All rights reserved. No part of this publication may be reproduced or transmitted in any form or by any' means, ~lectronic or mechan~cal, including photocopy, r.ecprd~g,. 01: finy information .storage and retneval system, wIthout permIssIon m wntmg from the publisher. Published by The Fairmont Press, Inc. 700 Indian Trail Lilburn, GA 30247 Printed in the United States of America

10 9 8 7 6 5 4 3 2 1

ISBN 0-88173-122-6

FP

ISBN 0-13-950890-2

PH

While every effort is made to provide dependable information, the publisher, authors, and editors cannot be held responsible for any errors or omissions. Distributed by Prentice-Hall, Inc. A Simon & Schuster Company Englewood Cliffs, NJ 07632 Prentice-Hall International (UK) Limited, London Prentice-Hall of Australia Ply. Limited, Sydney Prentice-Hall Canada Inc., Toronto Prentice-Hall Hispanoamericana, S.A., Mexico Prentice-Hall of India Private Limited, New Delhi Prentice-Hall of Japan, Inc., Tokyo Simon & Schuster Asia Pte. Ltd., Singapore Editora Prentice-Hall do Brasil, Ltda., Rio de Janeiro

Contents

Preface ......................................................................................... xi Acknow ledgements ..................................................................... xiii List bf Frequently Used Abbreviations .......................................... xv Introduction .............................................................................. xvii CHAPTER 1: Heat Recovery Systems ............................................. 1 HRSGs for gas turbines ........................................................... 5 Natural versus forced circulation boilers ................................. 6 Auxiliary firing ..................................................................... 9 Computing fuel requirements ................................................. 20 HRSG system efficiency ....................... ;............................... 30 Fresh air firing .................................................................... 31 HRSG design features ........................................................... 33 Finned surfaces and design ................. ;.................................. 33 Steaming economizers ........................................................... 34 Emissions of NOx and CO ..................................................... 35 Methods of reducing pollutants ............................................. 37 Bypass dampers ................................................................... 41 Recent trends ........................................................................ 44 STIG and Cheng cycle systems ............................................... 47 Enhanced on recovery applications ...................................... 51 Reciprocating engine heat recovery ....................................... 52 Hydrogen plant waste heat boilers ....................................... 52 Boilers for sulfuric acid plant ............................................... 57 Incineration and heat recovery ............................................. 62 Solid waste incineration ....................................................... 62 RDF firing ........................................................................... 72 Fluid bed combustors for MSW .............................................. 74 Hazardous waste incineration .............................................. 76 High temperature corrosion .................................................. 77

v

Heat recovery boilers ........................................................... 80 Incineration of wood wastes, tires ......................................... 84 Incineration of liquids, fumes, VOCs ..................................... 87 Air heating applications ...................................................... 91 References ............................................................................ 93 CHAPTER 2: Fire Tube Boilers ...................................................... 97 Guidelines for fire tube boilers .............................................. 99 Design procedure ................................................................. 100 Determination of tube side coefficient .................................. 101 Example of design ""'"'''' ""'" ....................... ""'"'' ............. 103 Effect of tube size on design .................. """ .......................... 108 Simplified approach to design ............................................ 109 Predicting boiler performance .............................................. 117 Simplified approach to predicting performance ................... 119 Checks for fouling ............................................................... 120 Effect of scale on boiler performance .............. " ...... " ............. l21 Hydrogen plant boilers ..................... """"""'''''''' ............... 123 Gas bypass flow calculations ............................................... 124 Determining heat losses from boiler ..................................... 125 References ........................................................................... 126 Nomenclature ..................................................................... 126 CHAPTER 3: Water Tube Boilers ................................................ .131 Guidelines for water tube boilers .......................................... 133 Heat transfer calculations ................................................... 135 Convective heat transfer coefficient .................................... 135 Determination of tube side coefficient .................................. 139 Non-luminous heat transfer coefficient ................................ 140 c~ Gas pressure drop calculations ............................................. 148 In-line versus staggered arrangement ................................... 153 Design of evaporators .......................................................... 155 Performance calculations ..................................................... 162 Selecting designs with low pinch and approach points ............................................................ 165 Comparison of bare versus finned evaporator .......... " ............ 167 Radiant heat transfer ............................. "" ........................ 168 HRSG configuration and circulation""""""" ........ "" ........ " .170 vi

Design of superheaters ........................................................ 176 Design procedure ................................................................. 179 Performance calculations ..................................................... 181 NTU method of performance calculations ............................. 181 Metal temperature calculations .../ ...................................... .182 External radiation .............................................................. 186 Flow in parallel streams ..................................................... 189 Minimizing tube wall temperatures ..................................... 190 Steam temperature control. .................................................. 191 Design of economizers ......................................................... .196 Performance of complete HRSG ............................................ 197 References ...........................................................................201 Nomenclature .....................................................................201 CHAPTER 4: Simulation of HRSG Design and Performance ............................................ 205 Importance of HRSG simulation ........................ '" ................ 205 Design and performance calculations .................................... 206 Design temperature profile ..................................................207 Guidelines for selecting pinch and approach points ............................................................209 Example of design ............................................................... 213 Performance calculation procedure ....................................... 216 Software for HRSG simulation - COGEN ............................. 228 Supplementary firing and HRSG efficiency .......................... 231 Improving efficiency of HRSG ............................................. 231 Deaeration steam calculations ............................................. 238 Steam turbine calculations ................................................... 241 Optimizing temperature profiles using COGEN .................... 245 Using field data to simulate HRSG performance ..................248 Multi-pressure HRSG design and performance simulation ................................................. 250 References ...........................................................................254 Nomenclature ..................................................................... 255 CHAPTER 5: Specifying Waste Heat Boilers ............................... 257 Application or system design ............................................~ .. 257 Space and layout guidelines ................................................ 259 vii

~r---------------------------------------------------------

Gas data ............................................................................. 260 Boiler duty .........................................................................263 Auxiliary fuel data .............................................................. 264 Emission data .................................................................... .265 Feed water analysis, blow down .......................................... 265 Surface area, fin configuration ............................................. 270 Cost data for fuel, electricity and steam............................... 271 Drum sizing ......................................................................... 271 References ........................................................................... 274

APPENDIX A: Finned Tubes ........................................................ 275 Heat transfer calculations ................................................... 276 Fin efficiency and effectiveness ........................................... 278 Gas pressure drop ................................................................ 278 Tube wall and fin tip temperature ........................................ 280 Design example .................................................................. .281 Comparison of bare versus finned evaporator ........................ 286 Comparison of in-line versus staggered arrangement ................................................. .287 Fin configuration and performance ....................................... 290 Importance of tube side coefficient ....................................... 291 Effect of fouling factors ........................................................ 292 Surface area and duty .......................................................... 300 Nomenclature .....................................................................304 References ...........................................................................306

APPENDIX B: Low Temperature Corrosion .................................. 307 Causes and cures .................................................................. 307 Methqds of avoiding cold end corrosion ................................312 Condensation on surfaces ......................................................314 Corrosion is stacks, ducts ......................................................315 Heat loss calculations through multi-layer insulation ...................................................320 Hot casing design ................................................................322 Nomenclature .....................................................................326 References ...........................................................................326

viii

APPENDIX C: Heat Transfer Equipment Vibration ........................................................... 327 APPENDIX D: Gas Turbine Data ................................................. 337 APPENDIX E: Gas and Steam Properties ...................................... 349 Specific heat, viscosity, thermal conductivity of gases ........................................349 Enthalpy of gases ............................................................... .351 Estimating flue gas properties .............................................351 Effect of pressure on heat transfer ........................................353 Converting % volume to % weight ........................................ 355 Properties of steam and compressed water ............................355 APPENDIX F: Tube Thickness Calculations ..................................377 Tubes and pipes subject to internal pressure ........................... 377 Designing vessels and tubes subject to external pressure ..................................... 381 APPENDIX G: Conversion Factors ................................................391 INDEX ....................................................................................... 395

ix

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F

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Preface During the past 20 years I have had the opportunity of engineering a wide variety of industrial boilers and Heat Recovery Steam generators. During the past 7 years at ABeO Industries in particular I have had the pleasure of custom designing over two hundred fire tube and water tube waste heat boilers, each with different gas/steam parameters, which are in operation in the USA and abroad; these units were built for diverse heat recovery applications such as gaseous, liquid, solid waste and hazardous waste incineration systems, gas turbine exhaust, effluents from chemical plants such as sulfuric acid and hydrogen plants, petrochemical plants, cat crackers in refineries and for effluents from clean as well as dirty processes; the gas flow ranged from 2000 to 1.5 million pounds per hour, which implies a wide variety of boiler configurations and design features as you will see in the text. Energy management programs are vital to the economic life of any industry and heat recovery boilers playa dominant role in those projects which otherwise waste energy from hot flue gases. I decided to write this book after reviewing hundreds of specifications for heat recovery boilers prepared by consultants and would be users of the equipment; unfortunately several of them are poorly written without emphasis on the process aspects and optimization of installed plus operating costs, with a result that the end user or the owner gets an equipment which perhaps meets the budget requirement but which could incur significant operating costs in the form of higher gas pressure drop or fuel consumption or lower steam production year after year. The book addresses various aspects of heat recovery boilers, such as engineering, specifying, system design, optimization and performance evaluation. Hence engineers and managers involved in several disciplines of energy management including plant operation will find the book useful and informative. xi

The book is dedicated to professionals involved in any way with energy conservation and heat recovery. As pointed out by one, the earth is not for man, but man is for earth. Hence let us use the limited natural energy resources wisely with the future of mankind and the next generation in mind.

V. Ganapathy

xii

F

I I I

Acknowledgements

I

I would like to thank ABCO Industries for their encouragement and support in the preparation of this book and for the use of several ABCO illustrations and photographs.

I

I would also like to thank the following publications for permitting me to use my articles, which originally appeared in them:

I I

i

Power Power Engineering Chemical Engineering Oil and Gas Journal Hydrocarbon Processing Heating, piping and Air-conditioning Sci-Tech Publications Pennwell Books Marcel Dekker Inc. I would also like to thank ESCOA Corp for permitting me to use their correlations for extended surface heat transfer calculations. Several readers from various continents have been writing to me regarding my publications during the past several years, which has been indeed been motivating and I would like to thank them for their interest.

V. Ganapathy

xiii

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List of Frequently Used Abbreviations ABMA - American Boiler Manufactures Association ASME - American Society of Mechanical Engineers CO - Carbon Monoxide EOR - Enhanced Oil recovery FrB - Fire tube boiler GTE - Gas turbine exhaust HRSG - Heat Recovery Steam Generator MSW - Municipal Solid Waste NIMBY - Not in my back yard NOx - Nitrogen oxides NWL - Normal water level PPB - Parts per billion PPM - Parts per million RDF - Refuse Derived Fuel SCR - Selective Catalyst Reduction System STIG - Steam Injected Gas Turbine TDS - Total dissolved solids WHB - Waste Heat Boiler WTB - Water tube boiler VOC - Volatile Organic Compounds

xv

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Introduction The book is aimed at engineers, consultants and managers involved in specifying, operating, engineering, marketing and procuring waste heat boilers (WHBs) or heat recovery steam generators (HRSGs). It offers valuable information on not only the heat recovery systems in chemical plants, gas turbine cogeneration and combined cycle plants, solid waste, liquid and gaseous incineration systems and flue gas heat recovery in general, but also provides the characteristics of each system such as gas analysis, fouling and slagging tendencies, high and low temperature corrosion potential and the impact of these on design and performance aspects of HRSGs. During the past 15 years and particularly during the last 7 years at ABCO Industries I have had the opportunity of engineering a wide variety of fire tube and water tube waste heat boilers for different types of applications as mentioned above. Having designed over two hundred boilers with gas flows varying from 2000 to 1.5 million pounds per hour and steam flows varying from 2000 to 250,000 pounds per hour, I feel that custom designing HRSGs is an art as well as a science, as there are numerous configurations possible depending on economics, cleanliness of gas, gas and steam parameters and layout considerations. I have had also the opportunity of authoring four books and over 175 articles on heat recovery boilers and steam plant systems in journals such as Power, Power Engineering, Chemical Engineering, Heating Piping Air-Conditioning, Oil and Gas Journal, Hydrocarbon Processing, Plant Engineering; the feed back from the readers has been very encouraging, which prompted me to bring out this work. Another reason was that in the course of reviewing specifications on HRSGs from various consultants and engineering organizations, I felt that less emphasis was being placed on process and optimization aspects, which is very important in the long run to the owner of the plant. Due to lack of knowledge on HRSGs and their performance

xvii

aspects, several of the specifications are poorly written and do not furnish adequate information to engineer an economically and technically sound design. Many engineers also lack knowledge or do not know how to evaluate alternate design options. For example if you read Appendix A, several examples are given to show that with finned tubes one can have a lower surface area and still transfer more energy by proper choice of fin configuration. Several engineers and purchase managers still purchase HRSGs for critical applications based on surface area and are of the view that more the surface area the better and a design with a lower surface area would not perform. Also, I have come across several specifications which do not place emphasis on HRSG operating costs; during the life time of the HRSG, the cost of moving the gas through the system due to high gas pressure drop or the cost of fuel which is required to generate a desired quantity of steam may be very significant. While the consultant looks at the initial cost alone, to the owner of the plant who has to pay for the fuel and electricity for years to come, the life cycle cost of the HRSG is important. Hence addition of secondary heat recovery surfaces such as condensate h;eater or economizer may have to be looked into, though the initial cost may be slightly more. A few examples on the subject of evaluating operating and life cycle costs are discussed in the book. This book offers useful information on design and off-design performance aspects of Heat recovery systems and components such as superheaters, evaporators and economizers, which are elaborated by over 65 fully worked out examples. You will find quantitative answers to commonly asked questions on heat recovery boilers and systems; some of them are: • • • • • •

How can one improve the efficiency of a HRSG system? What is the effect of auxiliary firing on system efficiency? How to compute the fuel requirements and oxygen consumption for gas turbine exhaust boilers? How to select pinch and approach points? How do they vary with gas inlet conditions? What is the effect of scale on boiler performance and tube wall temperatures and heat flux? How to compute the dew points of hydrochloric acid, sulfuric acid, hydrobromic acid, nitric acid?

xviii

• • • • • • • • • • • •

Which is better arrangement for bare and finned tubes, in-line or staggered? How do boilers with finned tubes compare with bare tube design for the same duty? How to compute the gas temperature at the SCR at off-design conditions. How to avoid high and low temperature corrosion problems? How to use field data to predict off-design performance or fouling of HRSGs? With finned tubes can you transfer more duty with less surface area? What is the effect of fin configuration? How to size and predict off-design performance of fire tube and bare/finned water tube boilers, superheaters, economizers? How to compute tube wall and fin tip temperatures? How to compute thickness of tubes subject to internal or external pressure? How much gas should be bypassed for gas temperature control? What is the effect of gas pressure on heat transfer? How to evaluate HRSGs for possible noise and vibration problems?

The first chapter deals with heat recovery systems. HRSGs are used in various applications such as gas turbine exhaust, incineration systems, chemical plants and refineries to mention a few. In order to design a HRSG for any application, the characteristics of the gas stream are important. For example, auxiliary firing in gas turbine HRSGs is discussed in depth with examples on computing fuel requirements, oxygen consumption, impact on system efficiency and emissions. Features of boilers such as natural or forced circulation, single or multiple gas pass design, insulated casing or fully water cooled membrane wall construction are discussed, along with methods of minimizing steaming concerns in economizers. Various aspects of WHBs in Municipal Solid Waste (MSW) applications, Refuse Derived Fuel (RDF) fired units and other incineration systems are discussed with emphasis on type of boilers, whether fire tube or water tube, fouling and slagging concerns and high and low temperature corrosion potential. Methods of minimizing these concerns through boiler design and selection of steam parameters are addressed. The second and third chapters deal with Design and off-design performance calculation procedures for fire tube and water tube

xix

boilers with bare and extended surfaces. Plant engineers can use the simplified procedures described in the text for instance to check for fouling, estimate bypass flow for gas temperature control or estimate the gas temperature at the Selective Catalytic Reduction system SCR) at different load conditions. Effect of tube size and arrangement whether in-line or staggered on design and performance is elaborated quantitatively. Examples are also given on how to compute the tube wall temperatures, including the effect of scale. Arrangement of headers on flow mal-distribution in superheaters and the effect of tube configuration on direct radiation to tubes are also discussed. Circulation aspects are also discussed along with various configur- . ations available for superheaters, evaporators and economizers. Simulation of single or multi-pressure unfired or fired HRSGs can be performed using the methodology described in chapter 4. Guidelines on selecting pinch and approach points are discussed. One can predict the performance of complex unfired and fired multipressure HRSGs under different load conditions without actually designing the unit. Such studies would be helpful to consultants in simulating the entire combined cycle or cogeneration plant behavior and economics. Methods of improving the efficiency of HRSG systems through addition of condensate heater, deaerator coil or heat exchanger are addressed. Examples illustrate how one can also optimize the temperature profiles b;y rearranging the heating surfaces. Methods of computing deaeration steam requirements and power output from steam turbines are discussed with examples. The software COGEN which is used in HRSG evaluations is recommended to those involved in engineering combined cycle and cogeneration projects, as on can simulate complex HRSG systems without actually designing the plant, saving a lot of engineering time Chapter 5 shows how one should specify waste heat boilers from the process view point. Adhering to the guidelines will save a lot of time for both the boiler designer and the purchasing manager responsible for evaluating alternate bids. Advantages of extended surface over bare tube is discussed in detail with examples in Appendix A. Effect of arrangement of tubes i.e. in-line versus staggered and the selection of fin configuration are elaborated. Examples also show how one can transfer more energy with less surface area with finned tubes. The effect of tube side xx

coefficient and tube and gas side fouling factors on tube wall and fin tip temperatures are illustrated with examples. Appendix B cites the causes of low temperature corrosion and suggests methods of minimizing the problems. Dew points of hydrochloric, sulfuric and hydrobromic acid may be computed using the correlations given. Heat losses thrqugh casing may be evaluated using the program described. One can evaluate an HRSG design for possible noise and vibration due to vortex shedding using the methods discussed in Appendix C. Gas turbine based HRSGs are widely used in cogeneration and combined cycle plants. Appendix D gives the exhaust gas data for several widely used machines. Gas and steam properties are provided in Appendix E with correlations for saturated and superheated steam. Example illustrates how one can compute gas mixture properties. The effect of gas pressure on heat transfer inside and outside tubes is also addressed. Appendix F shows the method of computation of thickness of tubes subject to internal and external pressures according to recent ASME code procedures. In sum, over sixty five examples from real life situations are worked out covering design and off-design performance aspects of various types of waste heat boilers and systemsi in addition, elaborate matter of fact discussions on systems and equipment should make this book indispensable to engineers involved in various disciplines of heat recovery. This book in the authors view would be an invaluable addition to the library of engineers and consultants involved in operation, maintenance, engineering, specifying or purchasing waste heat boilers. Since no single book can cover all of the aspects of the subject, the author suggests that serious professionals involved with heat recovery systems and waste heat boilers should acquire the other books and the software COGEN written by himi for more information, please contact the author at : V. Ganapathy, P.O. Box 673, Abilene, Texas 79604, USA.

xxi

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Chapter 1

Waste Heat Boilers

Waste heat boilers (WHBs) or Heat Recovery Steam Generators (HRSGs) as they are often called are used to recover energy from waste gas streams such as those encountered in sulfuric acid or hydrogen plants, refineries, solid, liquid and gaseous incineration systems, power plants and in cogeneration systems using gas turbines and reciprocating engines. With rising fuel costs and limited supply of premium fuels, it is prudent to maximize the energy recovered from waste gas streams whenever possible. Basically HRSGs can be classified into two broad categories: 1. Those which are required to cool gas streams to a desired temperature range from process considerations; examples could be found in hydrogen or sulfuric acid plants; in these plants, the energy recovery aspect is of secondary importance; the exit gas temperature from the boiler has to be controlled within a narrow range of temperatures for further downstream process purposes and methods such as gas bypassing would be used to achieve this objective. 2. In the other category of waste heat boilers, the objective is to maximize energy recovery compatible with considerations of high or low temperature corrosion and economics. Examples could be found in gas turbine based combined cycle or cogeneration systems, incineration plants and flue gas heat recovery in general. There is no standard design methodology or procedure for engineering of waste heat boilers, since one comes across a wide range of gas temperatures, gas analysis, pressures and steam parameters Table 1-1 shows some of the gas streams encountered in the industry. If the gas pressure is high, a fire tube boiler is preferred, Figure 1-1.

1

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Table 1-1 Composition of Typical Waste Gases

Gas 1 2 3 4 5 6 7 8 9 10 11 12 13

Temp.,oC

Pressure, atm

300-1,000 250-500 250-850 200-1,100 300-1,100 500-1,000 200-500 300-1,200 100-600 175-1000 250-1350 150-1000 300-1450

1 1 3-10 1 30-50 25-50 200-450 40-80 1 1 1 1 1.5

N2

NO

78-82 80-82 65-67 8-10 70-72 12-13 13-15 18-20 0.2-0.5 70-80 70-75 75-80 65-72 50-55

H2O

18-20 16-18 40-41 34-36

~

A

8-10 10-12 5-7 2-3

502

503

8-11 0.5-1.0

6-8

C~

CO

CI-4

9-10 6-8 13-15

7-9

0.3

0.2-1

H2S

H2

0-0.8

30-32 38-40 56-60 45-49

1-5 0.3-0.5 6-10 8-12 6-10 16-25 20-25

13-16 5-8 3-5 1-3 3-5

4-6 46-48 3-4 10-13 6-8 4-6 5-7 2-3

0.2-0.5

NH3

Hcl

18-20

traces 5-7

:s~

C;

2-3

3-4

1. Raw suller gases 2. ~ from converter 3. Nitrous gases 4. Primary reformer flue gases 5. Secondary reformer gases 6. Converted gases 7. Synthesis gas 8. Shell gasifier effluent 9. Gas turbine 10. Modular MSW incinerator 11. Chlorinated plastics incinerator 12. Fume incinerator 13. Sulfur condenser

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Figure 1-1. Elevated drum fire tube boiler [courtesy ABea Industries]

If the gas flow is large and several levels of steam pressure are used,

a water tube boiler as shown in Figure 1-2 may be appropriate. More discussions on fire tube versus water tube type could be found in Chapters 2 and 3. Another important aspect to the type of boiler is the cleanliness of the gas stream; if clean, extended surfaces could be used and the boiler may be made compact as in gas turbine applications; on the other hand if the gas stream is dirty as in municipal solid waste systems, the tube surfaces should be bare, with provisions for cleaning and ash removal. A large water cooled membrane wall radiant section may be required to cool the gases below the fusion points of eutectics before entering the convection section. Ample consideration should be given to high and low temperature corrosion aspects.

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198

Waste Heat Boiler Deskbook

Figure 3-18d. A Large Economizer Coil [ABCO Industries]

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Water Tube Boilers

Figures 3-18 e & f. Finned Tube Economizer and Boiler With Vertical Gas Flow Economizer [ABCO Industries]

199

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Figure 3-19. Printout of Results From HRSG Program WHB PERFORMANCE •• V.GANAPATHY ---PROJECT

•• SAMPLE

GAS FLOW-PPH = 400000 GAS TEMP IN -F =1500 GAS MW =28.24 GAS PRES-PSIA= 14.5 DRUM PRES-PSIA = 655 SAT TEMP-F = 496 BLOW ON = .040 FW TEMP-F =250 EXT DUTY-MMB/H = 9.00 HEAT LOSS-%= 2.00 FOUL FTR IN=.0010 PROS STM-PPH= 0 CO2 6.00

H2O 12.00

N2 70.00

HCL 0.00

H2S 0.00

DO 2.000 NW 40

DI 1.738 NO 5

EVAPORATOR CPG DUTY TG1 TG2 16.95 1500 1356 0.3001 H B L N 9.00 0.00 0.000 0.000

U 19.34 WS 0.000

ARRGT=IN

TG1 1356 1264

NO

L 12.50 12.50

6 8

10 1. 738 1.738

CONF OD CF 2.000 CF 2.000 FOUL FTR = .001

DI 1.738 1.738 1.738

NW

NO

40 40 40

2 2 12

N 0.00 0.00

H2 0.00

U 16.71 16.24

CO 0.00

SURFP 942 ST 3.750

16.945 GAS DP =

SUPERHEATER DUTY CPG 0.2963 10.64 13.02 0.2929

TG2 1264 1150

ARRGT=IN

S02 0.00

FOUL FTR =0.001 DUTY=

NW 24 24

DO 2.000 2.000 2.000

02 12.00

SURFP 942 1256

FOUL FTR =0.001 DUTY=

DELT DELPG MAXVL 930 0.75 110.8 SL TFIN TWAL 4.000 546 546

DELT· 676 638

DELPG 0.46 0.58

ST 4.500 4.500

SL 4.000 4.000

DUTY STM IN STM OUT STH HTC 566 701 280 10.64 13.02 496 641 291

U 14.62 10.76 7.76 WS 0.000 0.000 0.000

64.508 GAS DP =

MAXVL

86.2 81.9 PROP STRMS 11.41 24 12.58 24

TWAL TFIN 774 774 706 706

ARGT=IN STM FLO = 123227

EVAPORATOR TG1 TG2 CPG DUTY 1150 1104 0.2903 5.20 1104 928 0.2865 19.81 928 563 0.2771 39.50 L N H B 13.45 0.00 0.000 0.000 11.50 3.00 0.750 0.102 11.00 4.00 0.750 0.050

S03 0.00

0.749

H B WS 0.000 0.000 0.0000 0.000 0.000 0.0000

STH PRES -PSIA= 625

CH4 0.00

SURFP 563 3573 25987 ST 3.750 3.750 3.750

SPRAY= 5665

DELT DELPG 631 0.12 515 0.71 196 2.51 TFIN SL 4.000 521 4.000 841 4.000 759

MAXVL

60.9 93.9 77.5 TWAL

521 619 576

3.339

ECONOMIZER DO 2.000 NW 40

01 1.738 NO 20

TG1 563 L 12.00

WAT TEMP IN= 250 ECO FOUL FTR= .001

TG2 CPG DUTY 312 0.2666 26.30 B N H 4.00 0.750 0.050

OUT =

454

ARRGT=IN

U SURFP DELT DELPG MAXVL 7.05 44791 83 1.99 44.7 WS ST SL WATDP STRMS 0.157 4.000 4.000 16.77 10

WAT HTC= 1451 % STH = 0

BOILER DUTY= 131.413 TOT GAS PROP=

WAT FLOW= 122264

STM SURF= 0

SPRY TEHP = 250

7.11 STEAM GEN= 123228 TOT BLR DUTY= 140.413

##############################################################################

Water Tube Boilers

201

automatically arrives at the firing temperature required to generate the desired final steam quantity if the fuel analysis is specified. The mechanical configuration is inputted along with gas inlet conditions entering the burner system (if used). The steam and gas side pressure drops, tube wall temperatures, spray water, firing temperature, gas analysis and properties are all automatically computed in seconds; provision exists to superheat steam from other boilers or to withdraw a certain quantity of saturated steam from the waste heat boiler for process purposes and superheat the balance quantity of steam. Fouling factors or tube configurations such as pitch, size could be varied for each section to simulate operating conditions.

REFERENCES 1. V. Ganapathy, "Applied Heat Transfer," Pennwell Books,

Tulsa, 1982 2. V. Ganapathy, "Charts can help give quick engineering estimates of gas pressure drop in tube banks," Oil and Gas Journal, March 1, 1982 3. V. Ganapathy, "Charts simplify estimation of non-luminous heat transfer coefficients," Hydrocarbon processing 4. V. Ganapathy, "Basic programs for steam plant engineers," Marcel Dekker, New York, 1984 5. V Ganapathy, "Simplified approach to designing heat transfer equipment," Chemical Engineering, April 13, 1987 6. V Ganapathy, "Determine spray water to desuperheat steam," Heating, Piping, Air-conditioning, December 1987 7. V. Ganapathy, "Steam Plant calculations manual," Marcel Dekker, New York, 1984

NOMENCLATURE C - factor used in Equation (3-9) and (3-44) Cp - gas specific heat, Btu/lb F d - tube diameter, in-subscript i and 0 refer to inside and outside f - friction factor for pressure drop; subscript g stands for gas.

-----~

202

"~---


if - fouling factor, sq ft h F /Btu -

subscripts i and 0 refer to inside and outside tubes. F - fraction of energy absorbed by direct radiation F1,F2" F3 - factors used in equations (3-25), (3-35) G - gas mass velocity, Lb/sq ft h H - enthalpy, Btu/lb L1 denotes difference in enthalpy. Subscript s stands for stearn h - heat transfer coefficient, Btu/ sq ft h F; subscripts c-convective, iinside tubes, n- non-luminous, o--outside hI heat loss factor; 1 % heat loss means hI =.99 k - gas thermal conductivity, Btu/ft h F K -factor used in equations (3-16) , (3-18) Km - tube metal thermal conductivity, Btu/ft h F L - length, ft or m; subscript e stands for effective length, ft. Nd, NH - number of tube rows deep N w _ number of rows wide NTU - Number of Transfer Units Nu - Nusselt Number Pc, Pw - partial pressure of carbon dioxide and water vapor Pr - Prandtl Number ~p - pressure drop; subscripts 0 or g stand for gas, in wc and subscript i for tube side, psi q - heat flux, Btu/sq ft h; subscripts i and 0 stand for inside and outside Q - duty, Btu/h; subscripts a and t stand for assumed and transferred: c, nand r for convective, non-luminous and radiant Re - Reynolds Number 5 - surface area, sq ftsl, stlongitudinal and transverse pitch, in L1 T -log-mean temperature difference, F T1, T2 - temperature entering and leaving, for gas Tg - average gas temperature, F or K ts - saturation stearn temperature, F U - overall heat transfer coefficient, Btu/ sq ft h F; subscripts i and 0 - stand for inside and outside tubes. v, V - specific volume, cu ft/Lb and velocity, ft/ s Wo, Wg - gas flow, pph Wi, Ws, Wf - Flow inside tubes, pph, inside, stearn and feed water E - effectiveness of exchanger, Equation (3-42)

J

--------

Water Tube Boilers

gas emissivity p-- gas density, lb/cu ft cr - Steffan Boltzman constant Y - factor defined in Equation (3-37) Eg-

203

~

Chapter 4

HRSG Design and Performance Simulation IMPORTANCE OF HRSG SIMULATION Water tube HRSGs (Heat Recovery Steam Generators) or WHBs (Waste Heat Boilers) as they are sometimes called are widely used in various applications as discussed earlier in Chapters 1 and 3. Methods of performing their design and performance calculations were elaborated in Chapter 3. The mechanical configuration of the WHB had to be known in these calculations and the heat transfer coefficient U was evaluated in each case in order to predict the duty and heat balance. However when consultants and system engineers are in the early stages of developing a project they need some basic information about HRSG performance and capabilities at different plant loads and operating points. For example in a gas turbine cogeneration plant, if the HRSG performance can be simulated at different· ambient conditions and loads, they can study the complete system performance knowing the steam generation, auxiliary fuel consumption, system efficiency etc., and use the information to select plant auxiliaries such as steam turbine, condenser and deaerator or evaluate the overall plant economics. They may obtain this information from HRSG suppliers; however, during the early stages of a project, a lot of time can be wasted by consultants if they depend on HRSG suppliers for this information for the simple reason that too many alternatives or gas turbines may be involved and HRSG suppliers may take a long time to design a HRSG for each alternative and then provide the 205

206


performance data for different operating conditions; also, they may not respond as quickly as the consultants would like to, particularly if it is a study and the gas inlet conditions are not firm. Also it would be in the interests of consultants to be able to simulate the HRSG thermal performance and not be tied to any particular HRSG design or supplier. Since they are only interested in the thermal performance, the mechanical details are not important at this stage. Once the study is completed and a particular gas turbine or heat source is selected, then the consultants can approach different HRSG suppliers for proposals, knowing what kind of performance and steam production they can expect for their system; this also makes them knowledgeable and the process of evaluation of HRSG bids can be quickened. Another reason for consultants to perform this simulation themselves is that they can optimize the steam system (whether to maximize HP or IP steam, where to take off the steam for deaeration, how much LP steam should be generated etc.,) as they are more familiar with cost of steam, fuel and the utilities and the . steam needs of the customer. Is there a method of evaluating HRSG performance without knowing the mechanical design details such as tube size, length, fin density etc? Can one obtain information on steam generation capabilities and temperature profiles without doing a mechanical design? The answer is YES and this chapter will elaborate the method, which can be effectively used by plant engineers, consultants and even HRSG designers to simulate and optimize a HRSG, be it a single pressure system or a complex multi-pressure unit and integrate it into the steam plant.

DESIGN AND PERFORMANCE CALCULATIONS There are two basic types of calculations performed while selecting a HRSG; one is the "design" calculation; in this mode, the basic configuration or the disposition of various surfaces is arrived at including the gas and steam temperature profiles and the steam production. Once a unit is designed, it means that the surface areas of

j

r 2rJ7

HRSG Design and Performance Simulation

the various components such as the superheater, evaporator and economizer and the configuration are frozen. Two important variables namely the PINCH point and APPROACH point determine the complete design, the design temperature profiles and the steam production, as explained below. The "performance calculation" tells the user how the HRSG that has been designed performs at different other gas flows, inlet temperatures, or steam parameters. This is a complex iterative procedure and will be discussed later. There is only one design point but there are several off-design performance points.

DESIGN CALCULATIONS AND DESIGN TEMPERATURE PROFILE Figure 4-1 shows the temperature profile for a simple HRSG for a single pressure case with a superheater, evaporator and economizer. The desired superheated stea:n:Lpressu~e aEd temperature ani kno'Y.-n as al~ thelee
Ity\


232

RESULTS •••• DESIGN CASE

UNFIRED

amb temp-f= 60 rel hum-%= 0 heat loss-%= 1 gas temp to HRSG= 900 gas flow-pph= 150000 %·vol C02= .:j H20= 7 N2 = 75 02 = 15 S02= 0

gas temp in-out-F EVAP ECO

900 408

408 327

wat/stm in-out-F 368 240

388 368

duty press MM b/h psig 19.57 3.10

200 210

flow pph

pstm pinch apprch % F F

22779 100.0 23462 0.0

20

(a)

20

RESULTS ••• PERFORMANCE CASE ••• FIRED amb temp-f= 60

reI hum-%= 60 heat loss-%= 1 gas temp to HRSG= 900 gas flow-pph= 150000 % vol C02= 3 H20= 7 N2 = 75 02 = 15 S02= 0 gas temp in-out-F BURN 900 1289 EVAP 1289 419 419 315 ECO

wat/stm in-out-F 0 336 240

gas flow after HRSG 0.00 fuel

0 388 336

duty press MM b/h psig 17.30 35.80 4.07

0 200 210

flow pph

pstm pinch apprch % F F

808 0 40029 100.0 41230 0.0

(I)

32

52

150808 % vol c02= 3.92 h20 = 8.79 n2=74.30 02=12.98 s02=

GAS: vol %

1 methane= 95 2 ethane= 5 lhv-btu/cu ft= 949 lhv-btu/lb = 21422 aug air-pph = 0

RESULTS ••• PERFORMANCE CASE ••• FIRED amb temp-f= 60 reI hum-%= 60 heat loss-%= 1 gas temp to HRSG= 900 gas flow-pph= 150000 % vol C02= 3 H20= 7 N2 = 75 02 = 15 S02= 0

gas temp in-out-F

wat/stm in-out-F

duty press MM b/h psig

flow pph

pstm pinch apprch % F F (c)

BURN 900 1714 EVAP 1714 430 ECO 430 310

0 315 240

gas flow after HRSG = 0.00 fuel

0 388 315

37.53 55.16 4.74

0 200 210

1752 0 60128 100.0 61932 0.0

42

73

151752 % vol c02= 4.97 h20 =10.84 n2=73.50 02=10.67 s02=

GAS: vol %

1 methane= 95 2 ethane= 5 lhv-btu/cu ft= 949 lhv-btu/lb = 21422 aug air-pph = 0

Figure 4-7. Printout of results of example 2 from COGEN software,

233


m ·ffi I

i i i

i! ~1I:l=-:::f:=~U._

!

-'-4

i i

i

i ........L i ~ ......

iI

i Figure 4-8. HRSG with condensate heater and glycol heater. [ABeD Industries]

Improving HRSG Efficiency Using Secondary Surfaces Figures 4-9, 4-10, 4-11, 4-12 show the concept of adding a condensate heater, a heat exchanger or a deaerator to improve the efficiency. The system consists of a HRSG generating superheated steam for producing electrical power via a steam turbine condenser system. The steam for deaeration is taken from an extraction point in the turbine as it is more economical to do so than take it off the HRSG. The reason for this will be given later. Condensate heater. This is basically an economizer used for preheating the mixture of condensate returns and make up before it enters the deaerator, Figure 4-10. A 20°F approach is used; that is if the deaeration temperature is 240°F, the design condensate heater exit temperature is 220°F. The following aspects of this option should be considered. 1. The log-mean temperature difference between the gas and water is higher than in the deaerator option described below; as a result the surface area requirements will be lower.

Figure 4-9. Basic arrangement: Option 1 has heat recovery only provided by HRSG [Power]

Figure 4-10. Condensate heater is added downstream of HRSG economizer in Option 2 to preheat condensate reducing deaeration steam. [Power]

Figure 4-11. In Option 3, heat exchanger is used to preheat makeup/condensate mixture before entering deaerator. [Power]

Figure 4-12. A low pressure evaporator can be used to generate low pressure deaeration steam, maximizing electrical output. [Power]

236


2. Due to the low water temperature at the inlet, the tube wall temperature will be lower and hence condensation of water vapor can occur resulting in corrosion .. It is suggested that the tube wall temperature be above the water vapor dew point of the exhaust gas. If there is steam injection in the gas turbine, the water vapor dew point is higher and hence suitable precautions have to be taken such as better choice of tube materials. If the exhaust gases contain sulfuric acid vapor or sulfur dioxide, then this option is not recommended. 3. The gas pressure drop will also be lower due to the lower surface area requirement.

Table 4-4. Results of Study of Deaeration Steam and Electric Output * Option Design reference Exhaust gas, 103 Ib/hr Gas-inlet temperature, F Stack-gas temperature, F Steam to turbine, 1031b/hr ** Steam to deaerator (from turbine), Ih/hr Feedwater temperature, F Mixture temperature (to deaerator),F Electric power, kW

(2) Cond htr 550 975 310 80

(3) Heat exch 550 975 323 80

(4) L-p evap 550 975 297 80

10,250 1730 240 240

3400 151

0 240

220 6830

200 6770

107 6890

(1 ) Base case 550 975 374 80

107 6528

* Assumptions: Natural-gas fired gas turbine; gas analysis, % vol: COz =3.5, HzO = 10, Nz = 74, Oz = 13.5; pinch point = 20°F; approach point = 20°F, blowdown = 2%; deaerator pressure = 10 psig; makeup temperature = 60°F; condenser pressure = 2.5 in. Hg abs 0.22 psia); heat loss in HRSG = 1%. **Conditions: 620 psig, 650°F.


237

Heat Exchanger. Figure 4-11 shows a heat exchanger used to improve the efficiency of a HRSG system. A water to water exchanger preheats the make up using the feed pump discharge before it enters the economizer. The make up water temperature increases thereby reducing the de aerating steam requirements. Since the feed water to the economizer is lower, the stack gas temperature reduces. 1. If the gas stream contains sulfuric acid, then this scheme is unsuitable due to the low tube wall temperature at the economizer inlet. Better tube materials may have to be used if water vapor dew point is high. 2. This scheme is attractive if the make up water temperature is low say 40 to 60°F. Then, the feed water can be cooled further in the exchanger. The size of the exchanger will also be smaller due to the larger log-mean temperature difference. Also, if the make up temperature is very high, on the order of 150 to 180°F, the exchanger may not be feasible due to the resulting low log-mean temperature and possible temperature cross conditions in the exchanger. 3. Due to the lower feed water temperature, the size of the economizer will increase. This cost has to be evaluated along with the cost of the exchanger, piping and associated valves. Deaerator coil. Figure 4-12 shows the deaerator coil used to recover additional energy from the gas stream. The bundle generates low pressure steam for de aeration and is essentially another evaporator. This may be an integral deaerator or simply an evaporator generating low pressure steam, which is taken to a deaerator. 1. The surface area required will be more due to the low logmean temperature difference compared to the condensate heater. 2. This scheme is suitable even if the gas stream contains sulfuric acid as the evaporator pressure can be raised to a value close to or above the acid dew point. The tubes will be at least 5 to 10°F above the saturation temperature. 3. This system is more expensive than the others due to the size and the use of drums and associated trim and controls. 4. The gas pressure drop will also be higher than the other options, due to the size.

238


USE OF COGEN TO ANAL YZE THE OPTIONS Example 3: A HRSG generates about 80,000 pph of superheated steam at 620 psig and 650°F from 550,000 pph of gas turbine gases at 975°F; the steam is expan9.ed in a steam turbine to generate power, with deaeration steam taken from an extraction point in the steam turbine. Study the options assuming that the gas stream contains no sulfur dioxide. Solution: Figure 4-13 shows the results from COGEN in the design mode for all of the options. A 20°F pinch and approach were used. The steam required for deaeration was estimated for each case through heat balance as described below. The power output from the steam turbine was obtained through computation of turbine steam rates. A Basic program for this is given later. Table 4-4 shows the summary of results. Since the deaerator option generates the maximum steam output to match de aeration requirements, the electrical power is the highest. One may also predict the HRSG performance at any other condition using COGEN and see how the system behaves.

DEAERATION STEAM CALCULATIONS Estimating steam quantity for deaeration is an important aspect of plant mass and energy balance calculations. Steam for de aeration should preferably be taken from an extraction point in the steam turbine if available. This results in a better system efficiency compared to the case where the steam is taken from the HRSG exit. Figure 4-14 shows the two schemes.

COMPUTING DEAERATION STEAM AND ELECTRIC OUTPUT To illustrate the procedure, consider two sources for deaeration steam for the base case (Option 1), Example 3. Assume, first, that the

239



UNFIRED

8mb temp-f= 60 reI hum-If;= 0 heat 10ss-%= 1 gas temp to HRSG= 975 gas.f10w-pph= 550000 'I; vol C02= 3.5 H20= 10 NZ = 73 02 = 13.5 S02= 0 gas temp

in-out-F

wat/stm in-out-F


flow pph

pstm pinoh approh 'I; F F (a)

SH EVA!> ECO

975 914 515

914 515 374

495 475 240

650 495 475


9.31 59.57 20.36

620 635 645

80085 .100.0 80085 100.0 81687 0.0

20

20

UNFIRED

amb temp-f= 60 reI hum-%= 0 heat 10ss-%= 1 gas temp to HRSG= 975 gas flow-pph= 550000 % vol C02= 3.5 H20= 10 N2 = 73 02 = 13.5 S02= 0 gas temp

in-out-F

wat/stm in-out-F


flow pph

pstm pinoh approh % F F (b)

SH EVAP ECO

975 914 515

914 515 324

495 475 151

650 495 475


9.31 59.57 27.65

620 635 645

80085 100.0 80085 100.0 81687 0.0

20

20

UNFIRED

amb temp-f= 60 reI hum-%= 0 heat 1099-%= 1 gas temp to HRSG= 975 gas flow-pph= 550000 % vol C02= 3.5 H20= 10 N2 = 73 02 = 13.5 S02= 0 gas temp in-out-F

wat/stm in-out-F


flow pph

pstm pinch apprch F F

"

SH EVAP ECO

975 914 515

914 515 374

495 475 240

650 495 475

9.31 59.57 20.36

620 635 645

80085 100.0 80085 100.0 81687 0.0

ECO

374

310

107

220

9.02

650

79870


(c)

20

20

0.0

UNFIRED

amb temp-f= 60

reI hum-%= 0 heat loss-%= 1 gas temp to HRSG= 975. gas flow-pph= 550000 % vol C02= 3.5 H20= 10 N2 73 02 = 13.5 S02= 0 gas temp in-out-F

=

wat/stm in-out-F


flow ppb

pstm pinoh approh F F

'"

(d)

SH EVAP ECO

975 914 515

914 515 374

495 475 240

650 495 475

9.31 59.57 20.36

620 635 645

80085 100.0 80085 100.0 81687 0.0

20

EVAP

374

297

107

240

10.93

10

10050 100.0

57 132

20

Figure 4-13. a, b, c, d: Results from COGEN program for various options.

240


Makeup. Y 60F (17001b:hr)

C"ndensate: 'OOF (OO,OOO·X /b/hr)

Mixture: '07F

Feedwater: 240F• • 01,700 /b/hr

Condensate: 10SF

Figure 4-14. Two alternatives for deaeration steam take off location; steam from HRSG (top) and from steam turbine (bottom), [Power]


241

steam is taken from the HRSG exit (Figure 4-14 left) rather than from the turbine. Makeup temperature is 60°F and condensate enters the mixing tank at 108°F (2.5-in. Hg condenser pressure). Neglecting flashed and vented steam, a mass balance around the mixing tank gives: 81,700 - X = Y + (80,000 - X), where X is the deaeration-steam flow and Y the makeup flow. Very simply, Y = 1700 lb/hr. (Flashed steam from blowdown and vent steam are neglected in this analysis.) An energy balance is then performed around the deaerator, giving: (81,700) (208) = (1700) (28) +(80,000 - X) (76) + 1319X, where 1319 Btu/lb is the enthalpy of steam at 620 psig and 650°F, 28 Btu/lb the make-up enthalpy, 76 Btu/lb the enthalpy of condensate returns, and 208 Btu/lb that of feedwater after deaeration. Solving this equation, X =87411b/hr steam: using 87851b/hr allows for losses. To obtain the electric output, we must compute the actual steam rate (ASR) for the expansion from 620 psig and 650"F to 1.22 psia (2.5 in. Hg) A program developed for this purpose gives ASR = 11.14 Ib/kWh. Allowing for 4% mechanical losses, P = 0.96 (80,000 - 8785) /11.14 = 6137 kW. Next, consider deaeration steam taken from the turbine extraction point (Figure 4-14 right) at about 30 psia. For an expansion efficiency of 70%, this corresponds to an enthalpy of 1140.6 Btu/lb (calculated from the program referenced). An energy balance in this case gives: (81,700) (208) = 1140.6 X + (80,000 - X) (76) + (1700) (28). Solving for X gives 10,206 lb/hr for deaeration-steam flow. To include losses, use 10,250 lb/hr. The program then gives 19 Ib/kWh for the ASR from 620 psig to 30 psia, and 11.14 Ib/kWh for the entire expansion. For the net electric output, P =0.96 [(10,250/19) + (80,000 10,250) /11.14] = 6528 kW. Thus, greater electric output is obtained by taking steam from the steam turbine, and so this is the preferred procedure for all other options. Results of similar calculations for those options are summarized in the Table 4-4.

STEAM TURBINE CALCULATIONS Engineers involved in cogeneration projects and power plant studies often need to calculate the steam properties during expansion in a steam turbine to evaluate the theoretical and actual steam rates

I Input P" rio

p" E

I Yes

Figure 4-15a (left): Expansion of steam in a turbine. Figure 4-15b (right): Logic diagram to obtain expansion results [Hydrocarbon Processing]

Sat steam properties


243

and hence, the electrical power output. With the help of this program written in BASIC, one can quickly evaluate all the pertinent data. Correlations used for steam property evaluation are also presented. Theory: Figure 4-15a & b shows the typical expansion process in a steam turbine.

TSR

= 3,413/(h1 -h2s)

(4-26) (4-27) (4-28)

Superheated steam enthalpy is computed from the equations given P, T as shown in Appendix E. For wet steam enthalpy, use: (4-29) h = xhv + (1 - x) hf where the saturated vapor and liquid properties are obtained from Appendix E given the pressure. The dryness fraction, x, is computed for each state from either the enthalpy relation in (4) or from the entropy relation:

s = XSv + (1-x)sf

(4-30)

The logic used in the program is shown in Figure 4-15b. Example 4: Two examples illustrate the use of the program. Example 4: Superheated steam at 650 psia and 750°F is expanded in a steam turbine to 150 psia with an expansion efficiency of 80%. Calculate the steam properties at inlet and exit as also the actual and theoretical steam rates.

Solution: Key in the program in Figure 4-16a. In the RUN mode the screen asks for the inlet pressure and temperature. If saturated, key 0 for

244


10 REM,PROGRAM COMPUTES TURBINE STEAM RATES 15 DIM A(8) ,B(8) ,C(8) ,0(8) ,E(8) ,F(8) ,G(8) ,HW(8,28), VW(8,28) 20 FOR 1=1 TO 8,READ A(I) ,B(I) ,C(I) ,0(1) ,E(I) ,F(I) ,G(I) ,NEXT I 25 DATA -.17724,3.83836,11.48345,31.1311,8.762969E-5,-2.78794E-8,86.594,-5.28012 6E-7, 2. 99461E-5, 1. 521874E-4, 6. 62512E-5, 8. 408856E-I0, 1. 86401E-14, .01596, -. 48799,3 04.717614,9.8299035 30 DATA -16.455274,9.474745E-4,-1.363366E-6,19.53953,2.662E-3,457.5802,-.176959, .826862,-4.601876E-7,6.3181E-l1,-2.3928,-.15115567,3.671404,11.622558,30.832667, 8. 74117E-5 35 DATA -2.62306E-8,54.55,-.14129,2.258225,3.4014802,14.438078,4.222624E-5,-1.56 9916E-8,1100.5,-1.67772E-4,4.2726S8E-3,.0104S04S,.95SO1509,9.101291E-S,-2.7592E-· 11, • l1S01 40 DATA -1.476933E-4,1.2617946E-3,3.44201E-3,-.OS49412S,6.S913SE-S,-2.4941E-11,1 .97364 45 PRINT" STEAM PROPERTIES AFTER EXPANSION •• BY V.GANAPATHY "'PRINT"" 50 INPUT" INLET PRESS-PSIA,TEMP-F(IF SATURATED INPUT 0 FOR TEMP)=";P1,Tl,PRINT" 55 INPUT" EXIT PRESS-PSIA,EXPN EFF-'=";P2,EF 60 IF T1=0 THEN GOTO 100 70 P=Pl,T=Tl,GOSUB 300 80 H1=Z,Sl=SV,GOTO 120 100 P=P1,GOSUB 400 110 Hl=HV,Sl=SV,T1=TSAT 120 P=P2,GOSUB 400 130 IF Sl>SV THEN GOTO 200 140 X=(Sl-SL)/(SV-SL),H2S=HV*X+(1-X)*HLIQ 150 H2=Hl-.0l*EF*(Hl-H2S) 160 IF H2>HV THEN GOTO 200 170 XF=(H2-HLIQ)/(HV-HLIQ),DELH=H1-H2,T2=TSAT,GOTO 290 200 T2=TSAT,RA=1000-TSAT 210 P=P2,T=T2,GOSUB 300 220 S2C=SV,IF ABS«Sl-S2C)/Sl)

150,000

< < < < < <
> > > > > 20140 7.17 3.15 21 52295 484

Tubes wide = 18,length =1 Oft, square pitch = 4.0 in.; finned tubes use

:4 fins/in. serrated fins, .75 in. high, .05 in. thick.

The advantages of using extended surfaces are obvious. The finned tube design is more compact as it has fewer rows deep; this also results in lower labor cost. The length of drums or casing would also be smaller as a result of fewer rows deep, resulting in savings in material cost. The gas pressure drop is also lower, resulting in lower operating costs. It can also be shown that the weight of the finned bundle is much lower. However the heat flux inside the tubes is much higher with finned tubes, as also the tube wall temperature. This is due to the larger ratio of external to internal surface area. The heat flux difference is more pronounced when the gas inlet temperature is higher, as in fired HRSGs. Hence care must be taken to use appropriate finning. A few bare rows of bare tubes, followed by tubes with low fin density and then with high fin density tubes is recommended.

COMPARISON OF INLINE VERSUS STAGGERED ARRANGEMENT Both inline and staggered arrangements have been used with extended surfaces. The advantages of staggered arrangement are higher overall heat transfer coefficients and lesser surface area; cost

288


could be marginally lower depending on the configurationj gas pressure drop could be higher or lower depending on the gas mass velocity used. If cleaning lanes are required for soot blowing, an inline arrangement is preferred. Solid as well as serrated fins are used in the industry. Generally solid fins are used in applications where the deposition of solids are likely. The following examples compare the effect of arrangement on boiler performance. Example 4: 150,000 pph of turbine exhaust gases at 10000P enter an evaporator of a waste heat boiler generating stearn at 235 psig. Determine the performance using solid and serrated fins and inline versus staggered arrangement. Tube size is 2 x 1.77 in.j tubes/row = 18, length = 10 ft. use 2 x .75 x .05 and 5 x .75 x .05 fins. Solution; Using the Escoa correlations and the methodology discussed in Chapter 3 for evaporator performance, the following results shown in Table A-5 were arrived at. Data in column 3 are for a staggered design with a duty close to the inline arrangement. Table A-5 Results of boiler performance (solid/serrated fins and inline/staggered arrangements)

1.

2. 3. 4. 5. 6. 7.

8. 9. 10.

Table 5a-5 x .75 x .05 Table 5b-5 x .75 x .05 x .157 Solid fins Serrated fins 1 2 3 col.no 1 2 3 in arrgt in st st st st 18 18 18 18 18 18 tubes/row 20 no. deep 20 18 20 17 20 length 10 10 10 10 10 10 6.51 7.71 7.71 7.18 8.67 8.69 Uo 4.87 4.38 LlPg 2.76 3.19 4.62 5.45 ,23.30 23.14 23.46 23.28 23.24 23.55 dutyQ 410 415 exit gas,F 418 416 414 408 surface (sqft) 21677 21677 19509 20524 20524 17446

-

Appendix A - Extended Surface Heat Transfer

1. 2. 3. 4. 5. 6. 7.

8. 9. 10.

289

Table 5c-2 x .75 x .05 Table 5d-2 x .75 x .05 x .157 Solid fins Serrated fins 1 col. no 2 2 1 2 3 in arrgt in st st st st 18 18 18 18 18 18 tubes/row no. deep 20 20 18 20 18 20 10 10 length 10 10 10 10 9.75 10.85 10.86 Uo 10.02 11.42 11.45 1.72 2.51 2.27 2.33 1.79 2.59 L1Pg 21.68 22.21 21.70 duty Q 21.59 22.22 21.72 exit gas 455 443 455 454 442 458 surface 9802 9802 8822 8407 9341 9341

duty-MMBtu/h; L1Pg - in we; surlaee-sq It; temp--F; Uo-Btu/sq It h F. The following observations may be made: 1. Staggered arrangement results in lower surface area for the same duty but higher gas pressure drop for both types of fins and fin density if the gas mass velocity is the same. For the same surface area you can transfer more energy with staggered configuration. 2. Serrated fins have a higher overall heat transfer coefficient for the same mass velocity; the surface area is lower than that of solid fins for the same duty; also, the gas pressure drop is slightly higher than that of solid fins for the same duty. Using 5 and 2 fins/in., the above design was revised to obtain a staggered arrangement with a lower pressure drop, closer to the inline configuration, for the same duty. Results are shown below. Table A-5e compares inline versus staggered designs for nearly the same gas pressure drop and same duty. It may be seen that due to the use of a lower gas mass velocity with staggered arrangement (more tubes per row), the gas pressure drop is reduced. It turns out that less surface area is required with staggered arrangement for the same duty and pressure drop. The staggered design could be marginally less in cost but there are other aspects to look into such as the effect of pitch on ligament efficiency, wider headers or longer tubes and above all, cleaning

.


290

requirements. Selection of type of arrangement is mostly based on the past experience of the company. Table A-5e Comparison between staggered and inline designs for nearly same duty and pressure drop 1.

2. 3. 4. 5. 6. 7. 8. 9. 10. 11.

col no arrgt fin type fin config tubes/row no deep length

Uo ,1Pg Q surface

2 1 in st serr serr 5x .75 x .05 x .157 20 18 16 20 10 10 8.36 7.18 3.19 3.62 23.24 23.31 20524 18244

4 3 in st solid solid 2x.75x.05xO 18 20 20 16 10 11 9.75 10.02 1.72 1.42 21.68 21.71 9802 9584

FIN CONFIGURATION AND PERFORMANCE Now that we have a feel for the effect of type of fins and arrangement on the performance, the next important question is how should the fin configuration be selected? Does the tube side coefficient influence the choice? Should one select a design simply because it has more surface area than another one? Can we transfer more energy with less surface area? We will answer the above questions in the following sections. 1. Higher fin denSity or height means lower U. From Figure A3, it can be seen that for a given mass velOcity, higher the fin denSity or height, lower the gas side coefficient or effectiveness, which results in lower Uo. 2. Higher fin density or height means higher ,1Pg. Even after adjusting for the increased surface area per row, it can be shown that higher the fin density or height, higher will be the gas pressure drop for mass velocities which are close.


291

IMPORTANCE OF TUBESIDE COEFFICIENT A simple calculation may be done to show the effect of tube side coefficient on U o. It was mentioned earlier that higher the tube side coefficient, higher can be the ratio of external to internal surface area. In other words, it makes no sense to use the same fin configuration, say 5 fins/in. fin density, for a superheater as well as for an evaporator. Rewriting Equation A-I based on tube side area and neglecting other resistances: (A-27)

Using the data from Figure A-3, Ui values have been computed for different fin densities and for different hi values. These are shown in Table A-6. Also shown are the ratio of Ui values between and the 5 and 2 fins/in. designs as well as their surface area. Following conclusions can be drawn. a. As the tube side coefficient reduces, the ratio of Ui values (between 5 and 2 fins/in.) decreases. With hi = 20, Ui ratio is only 1.11. With a hi of 2000, Ui ratio is 1.74. What this means is that as hi decreases the benefit of adding more external surface becomes less attractive. We have 2.325 time the surface area but only 1.11 times the improvement in Ui. With a higher hi of 2000, the increase is better, 1.74. b. A simple estimation of tube wall temperature can tell us that higher the fin density,higher will be tube wall temperature. For the case of hi = 100: With n = 2, Ui = 39.28, gas temperature = 900°F and fluid temperature of 600F, the heat flux qi = (900-600) x 39.28 = 11784 Btu/sq It h. The temperature drop across the tube side film (hi = 100) = 11784/100 =118°F. The wall temperature =600 + 118 =718°F. With n = 5, Ui = 53.55, qi =53.55 x 300 = 16065 Btu/sq It h. Tube wall temperature = 600 + (16065/100) = 761°F. Note that comparison is for the same height. The increase in wall temperature is 43°F.

292


c. The ratio of the gas pressure drop between the 5 and 2 fins/in. designs (after adjusting for the effect of Ui values and differences in surface area for the same energy transfer) increases as the tube side coefficient reduces. It is 1.6 for hi = 20 and 1.02 for hi = 2000. That is, when hi is smaller, it is prudent to use a lower fin surface .. Effect of fouling factors: The effect of inside and outside fouling factors fti and !fa are shown in Table A-7. Following observations can be made: a. With a smaller fin density, the effect of Iii is smaller. With .01 fouling and 2 fins/in., U o = 6.89 compared with 10.54 with .001 fouling. The ratio is .65. With 5 fins/in., the corresponding values are 4.01 and 7.56, the ratio being .53. That means with increased tube side fouling, it makes sense to use a lower fin density or lower ratio of external to internal surface area. The same conclusion was drawn with a lower tube side coefficient. b. The effect of ffo is less significant as it is not enhanced by the ratio of external to internal surface area.

PERFORMANCE AND SURFACE AREA Let us now study the performance of an evaporator and superheater with different fin configurations to bring out the fact that it is possible to transfer more energy with a lower surface area and at a lower gas pressure drop.HRSGs invariably use extended surfaces if the gas stream is clean as in gas turbine exhaust, air-tofluid heat transfer, and similar applications. With bare tubes, one could probably assume that the greater the surface, more energy will be transferred (for comparable velocities). With HRSGs using extended surfaces, however, one can fall into a trap by evaluating alternate designs or bids based on surface area alone. The reason is that different combinations of fin height, thickness, and density lead to different heat transfer coefficients, ffn efficiencies, and overall heat transfer coefficients. A large surface area does not necessarily mean more energy transfer. The energy

------

293


transfer capability depends on the product of the surface area and the overall heat transfer coefficient, not on surface area alone. We shall illustrate this with two examples, one for an evaporator and one for a superheater. Table A-6 , Effect of hi on Vi

[Calculations based on: 2. 0 x .105 tubes, 29 tubes/row, 6 ft long, 0.05 thick serrated fins; tubes on 4.0 in. square pitch; fin height = 0.75 in.; gas flow = 150,000 pph ; gas inlet temp = 10000 P] 1. hi 2 2 5 2. n, fins/in. 5 5591 5591 6366 3. G, 1b/sq ft h 6366 .01546 .00867 .01546 .00867 4. At/AI/1J hg 2.73 1.31 7.03 5. Uo 4.12 15.28 17.00 39.28 53.55 6: Ui 7. ratio 5 2.325 < 8. ratio Ui 9. ratio LlFg

2 5 5591 6366 .01546 .00867 11.21 8.38 62.66 109 >

(surface area of 2 fins/in. tube = 2.59 sq ft/ft and for 5 fins/in. = 6.02)

Table A-7a affect of fli, tube side fouling factor (tube side coefficient = 2000) 1.Fins/in., n 2. Uo clean

3·fti 4. Uo dirty 5. Uo as %

2 11.21 .001 10.54 100

2 11.21 .01 6.89 65

5 8.38 .001 7.56 100

5 8.38 .01 4.01 53

-.!

-----------~


294

TableA-7b affect of ff01 outside fouling factor (tube side coefficient = 2000) 1. fins/in. 2. Uo clean 3. 110 4. U o dirty 5. U o as %

2 11.21 .001 11.08 100

2 11.21 .01 10.08 91

5 8.38 .001 8.31 100

5 8.38 .01 7.73 93

Example A-5: Evaporator Let us consider a finned evaporator for a HRSG for gas turbine exhaust. Table 8-A-a sets forth the design data.

Table A-Sa Data for HRSG evaporator design Gas flow, pph Gas inlet temperature, F Steam Pressure, psig Feedwater temperature, F Gas analysis, % by volume Carbon dioxide Water Nitrogen Oxygen Nominal gas turbine power, kW

150,000 900 235 240 3 7

75 15 4,500

We shall evaluate fin densities ranging from 2 to 6 fins per in. and fin heights of 0.5 and 0.75 in. Assumptions include 6 ft long tubes, 2 by 0.105 in., 29 tubes per row, 20 rows deep, serrated fins, 0.05 i. fin thickness, and an inline arrangement with 4 in. square pitch. The methodology of design is presented elsewhere, and the results are shown in Figure A-4.

295


STUDY OF EVAPORATOR Tube Size: 2 x .105 in , 29 tubes/row, 6 ft long, 20 deep, 4 in square pitch Fins: 2 to 6 per inch, 0.5 to 0.75 in high, .05 thick, serrated

Tabkl1-oata for HRSG evaporator design.

1lIbIe 2-E1fect of fin geometry on performance forthe same evaporator duty (see Ag. 2 also).

Duly,Q. Gas pressure drop, 4 ~.in.

Duty,MM8ll1h 18.41 18.J8 Gas pressure ckop, ill. WG U4UZ 12,633 1W2 Surface area. sq It OvenH heat transfer coef· ficient Btu pe~sq It.flr·F 8.60 9.42 Case A: 3 Hns per in. 0.15 in. high by 0.05 in.lhiclI

19 18.5 -§

~ 18

::E

Case 8: 4 fins per in., 0.50 in. high by 0.05 in. thick

~ 1

Table 3 -Effect of fin density on performance for the same flri height (evaporator).

17.5

&rIns periL 18.41 19.24 U4 3.86 12,633 23,444 4.50 85.50

3 fins peril.

17

10

16.5...1.--..:;---r----r-....,..-.....,.---'3'

Fins~rin.

Effeci of fin geometry on performance.

6

Duty, MM81ut.

Gas pressure drop, in. WG Surface area, sq It Increase in duly, percent

Increase in area, percent OvtraH heat ltansfer coef·

ficient, ntu per sq ft.flr·F

8.60

6.60

Figure A-4. Effect of fin configuration on evaporator performance. [Heating, Piping, Air-Conditioning]


296

Figure A-4 shows the effect of fin density and fin height on energy transferred, Q; gas pressure drop, LiP; and surface area, $. As the fin density increases, so do Q, 5, and LiP. However, the rate of increase in Q decreases while that for LiP decreases. One can see that for a particular Q, the surface area and the gas pressure drop are lower for the configuration with the lower external surface area. Table A-8b shows the results for a duty of approximately 18.4 MMBtu h. The higher heat transfer coefficient associated with the lower surface/ft helps in transferring the same energy with less surface and also at a lower gas pressure drop. Hence, with finned tubes, one has to look at the product of surface area and overall heat transfer coefficient and not just the surface area. Evaluating alternates from spreadsheet giving only surface areas, which unfortunately is being done in a number of engineering organizations, may result in a poor selection that will also be more expensive to operate because of the higher gas pressure drop. (Greater fin density and fin height result in greater gas pressure drop.) Typically, a 4 in. WG additional gas pressure drop in a HRSG results in a 1 percent drop in electrical power output of the gas turbine. Greater fin density or fin height can also result in higher tube wall temperature and fin tip temperature. This will be discussed when we take up the example of the superheater. Table A-8b Effect of fin geometry on performance for the same evaporator duty (see Figure A-4 also)

Duty, MMBtu h Gas pressure drop, in. WG Surface area, sq ft Overall heat transfer coefficient Btu per sq ft-hr-F

Case A

Case B

18.41 2.24 12,633

18.38 1.82 11,432

8.60

9.42

Case A: 3 fins per in., 0.75 in. high by 0.05 in., thick Case B: 4 fins per in., 0.50 in. high by 0.05 in. thick

297


Table A-8c shows the effect of fin density on the duty and gas pressure drop for the same fin height. Because of the lower heat transfer coefficient, the additional surface area of 86 percent results in only a 4.5 percent increase in duty. Also, the gas pressure drop is much higher. Based on an electrical cost of $0.06 per kWh and a gas turbine power output of 4500 kW, the additional gas pressure drop is worth: C =(3.86-2.24) x 0.06 x 8000 x 4500/400 C =$8740 per yr The additional energy output at $3 per MMBtu h is worth: C =(19.24-18.41) x 3 x 8000 C =$19,920 per yr Based on cost and overall economics, one could arrive at either Option A or Option B. However, selecting an option because there is more surface is simplistic and can lead to wrong conclusions and improper HRSG selection. Table A-Be Effect of fin density on performance for the same fin height (evaporator)

Duty, MMBtu h Gas pressure drop, in. WG Surface area, sq ft Increase in duty, percent Increase in area, percent Overall heat transfer coefficient, BTu per sq ft-hr-F

3 fins per in.

6 fins per in.

18.41 2.24 12,633

19.24 3.86 23,444

4.50 85.50 8.60

6.60

Example A-6: Superheater Let us now consider a superheater for clean gas application. Table A-9 shows the design data. We shall evaluate fin densities of 2 and 5 fins per in., and fin heights of 0.5 and 0.75 in. We shall also

- ---- - - - - - - - - - - - -


298

vary the number of rows deep to obtain desired duties of approximately 14 and 17.5 MMBtu h. Assumptions include lO.it long tubes, 2 by 0.120 in., 0.075 in. thick solid fins, 22 tubes per row, square pitch of 4 in., 22 streams and counterflow arrangement, and fouling factors of 0.001 hr ft 2 F /Btu on both the gas and steam sides. Table A-9 Data for HRSG superheater design Gas flow, pph Gas inlet temperature, F Gas analysis, % by volume Carbon dioxide Water Nitrogen Oxygen Steam flow, pph Entering steam temperature, F Leaving steam pressure, psig

200,000 1,200 7 12 75 6 100,000 491 600

Table A-lO shows the results of calculations for four cases, two at each of the two desired duties. The following facts can be inferred from Table A-10. 1) The energy transferred is the same for both the 2 and 5 fin per in. designs (Cases 1 and 2). 2) Because of the higher heat transfer coefficient of 11.79 versus 5.5, however, the surface area for Case 2 is nearly 2.16 times that for Case 1. This clearly shows that through poor fin geometry, one can have excessive surface area and still transfer the same amount of energy as a well designed configuration with a much lower surface area. Also, the higher fin density combination results in higher operating costs. The same conclusion is drawn by comparing Cases 3 and 4 and Cases 2 and 3. In Case 3 there is less surface area but more energy is transferred than in Case 2! 3) The tube wall and fin tip temperatures are significantly higher for the higher fin density. In Case 1, the tube wall

299


temperature is 836°F versus 908°F in Case 2 while the fin tip temperature is 949°F versus 1033°F. Hence, one could be forced into selecting better grade materials (at higher cost) for the tubes and fins by using a high fin density design, particularly in superheaters with higher steam temperature requirements. Table A-lO Effect of fin geometry on superheater performance Case 1 Case 2 Case 3 Case 4 Duty, MMBtu h Leaving steam temperature, F Gas pressure drop, in. WG Leaving gas temperature, F Fins per in. Fin height, in. Fin thickness, in. Surface area, sq ft Max tube wall temperature, F Fin tip temperature, F Overall heat transfer coefficient, Btu per sq ft-hr-F Tube side pressure drop, psi Number of rows deep Fin efficiency, percent

14.14 14.18 17.43 17.39 747 747 689 691 0.65 1.20 1.15 1.37 951 950 892 893 2 2.5 4 5 0.50 0.75 0.75 0.75 0.075 0.075 0.075 0.075 2471 5342 5077 6549 836 908 931 905 949 1033 1064 1057 11.79 9.0 6 79

5.50 6.5 4 70.8

8.04 11.0 7 63

6.23 9.0 6 68

As a general rule, the lower the tube side coefficient (as in superheaters or air heaters), the lower the ratio of external to internal surface area should be. By calculating the heat transfer coefficient based on inside tube diameter, one can easily show that a high ratio of external to internal surface area does not improve performance. Superheaters using more than 3 fins per in. do not contribute to improved duty and have to be looked at carefully. Also, they can be counter-productive, leading to higher tube wall and fin tip temperatures or gas pressure drops.


300

The author's advice to potential users of finned surface is: Don't evaluate any finned surface based on surface area along; try to understand the effect of fin configuration on heat transfer coefficients, gas pressure drops, and metal temperatures.

SURFACE AREA OF FINNED TUBES Finned tubes are widely used in boilers, economizers, superheaters and other heat-transfer devices. Estimating their surface area involves solving the following complex equation. (Solid Fins): A =1to[4dh + 4hz + 2bd + 4bh] + l1td) (I-ob) 24 \12

(A-28)

This nomograph can be used to estimate tube surface area in ft2 / ft. It covers the most common configurations in boiler plants: Tubes of 1.25 in. O.D. to 4.0 in. O.D., fin heights of 0.5 in. to 1.0 in., and fin densities of 1 to 6 fins/in. For quick estimates, the effect of fin thickness on surface area can be neglected. For example, the surface area of a 2.0 in. tube with 3 fins/in., where each fin is 0.75 in. high and 0.102 in. thick, is 3.88 ft 2 /ft, whereas for the same configuration but with fins 0.036 in. thick, the area is 3.81 ft 2/ft. Example A-7: Estimate the surface area of a finned tube with an O.D. of 2.0 in. and a fin density of 3 fins/in., where the fins are 0.75 in. high and 0.05 in. thick. Solution: The line on the nomograph illustrates the steps. Start at n = 3 on the right horizontal axis. Draw a vertical line up to the h = O.75in. line, then a horizontal line left to the d = 2.0 in. line. From there, draw a vertical line down to the left horizontal axis to find that A = 3.85 ft 2/ ft.

301


2.0 1.75

6

1.25

3 A.1t21ft

o

6

Nomenclature A Surfat,e area flI. ~ tube, /t'/ft; (including area of liDs ad tube """""") b Fill tlUtbess, ilL Ii Tube ..... diaJIIeter, in.

., "

F!nlleicbt.isL Fill cleasity, finsIia.

~+-b

b Figure A-S. Chart for estimating surface area. [Chemical Engineering]

I

I

T

I- ' -

-

l-

-

-

l-

i-

i

~


302

FIN WEIGHT ESTIMATION To evaluate structural and handling problems, fin weight must be known. While tables are available for figuring weight of bare tubes, tables are not readily available for finned tubes because of he several variables involved in the estimation of fin weight, namely the fin density, tube outer diameter, fin thickness, etc. With the enclosed chart, however, one can rapidly determine the fin weight for commonly used fin configurations in industrial heat transfer equipment. Of course, to determine the total weight of the tube or pipe, one must add the fin weight to the bare tube weight. The chart is based on the formula for solid fins of carbon steel given below: W = 10.68 x (D + H) (H + 0.03) NB where B = fin thickness, in. D =outside diameter of tube or pipe, in. H = fin height, in. (see drawing on chart) N = fin density, fins per in.

(A-29)

Example A-8: A 3 in. schedule 40 pipe is used in a fired heater for recovering energy from flue gases. If fin density is 4 fins per in., fin height is 0.75 in., and fin thickness is 0.06 in., determine the fin weight if fins are of carbon steel. Solution: Go up from D = 3.5 (outer diameter of a 3 in. pipe) to cut H = 0.75. Move left to cut N = 4 fins per in. and move down to cut fin thickness at 0.06 in. Move right to cut fin weight scale at 8.5 lb per ft. Multiplication factors for other fin materials are given in the table within the chart. If fins were made from 316 stainless steel, the fin weight would have been 1.024 x 8.5 =8.71b per ft. The weight of bare 3 in. schedule 40 pipe is 7.58 lb per ft, and hence the total weight of the finned pipe is 7.58 + 8.5 = 16.08 lb per

ft.

~_a.at

6

7 10

11

,,~

-II-B

='

8 9

12 13 14 15

~

i~~

h

;~

01

Muttiplh::ation factors for.

11,

17 ,,~

18 19 20

Typ. 304. 310. 316. 32! Typ. 409.410. 430. 26·1 Nickel20D

Incone1600.625

1.024 0.978 1.133 1.073

IncoloySDO

Incolo), 825 C.""ntor2o.Cb3 Hls:telloyB

1.013 1.038 1.024 1.179

Figure A-6. Chart gives weight of finned tubes. [Heating, Piping, Air-Conditioning]

., 304

Waste Heat Boiler Dcskbook

IMPORTANT CONCLUSIONS Several important aspects of finned tubes were discussed in this Appendix: They are summarized below for convenience. 1. More surface area does not necessarily mean more energy transfer. So don't purchase a boiler or finned equipment assuming that it will do better simply because it has more surface area. The spread sheet approach to purchasing should be avoided even though the difference in surface areas could be 100 to 200 %. Fin configuration affects U significantly and one should look at the product of U x S and not at S alone. One can transfer more energy with lesser surface area as shown in the examples. 2. Lower the tube side coefficient, as in superheaters or air heaters lower should be the ratio of external to internal surface area. That is, don't use the same high density fin configuration for an evaporator as well as for the superheater. It should be much lower for the superheater. Else, the tube wall and fin tip temperature would be higher as also the gas pressure drop. 3. Staggered arrangement results in lesser surface area requirements for the same energy transfer and same pressure drop. However the impact on ligament or other mechanical considerations should be looked into. 4. Inside fouling factor affects the performance more than the .outside fouling factor; the effects are similar to a low tube side coefficient situation. The tube wall and fin tip temperatures are also influenced.

NOMENCLATURE a-factor defined in gas pressure drop equation Ao, At, Ai - obstruction area, total area, inside tube area, sq ft/ ft AI - area of fins, sq ft/ft b - fin thickness, in. C1 to C6 - factors defined in Table A-I Cp - gas specific heat, Btu/Lb F


305

d - tube OD, in. 11- gas viscosity, Lb/ft h E - fin efficiency, fraction F, FI - factors accounting for gas properties f - factor defined in gas pressure drop equation ffi, ffo - fouling factor, inside and outside, sq ft h F /Btu G - gas mass velocity, Lb/sq ft h h - fin height, in. hi - tube side heat transfer coefficient, Btu/ sq ft h F he, hg - convective gas heat transfer coefficient, Btu/sq ft h F h n - non-luminous heat transfer coefficient, Btu/ sq ft h F h 0 - total outside heat transfer coefficient, Btu/ sq ft h F k - gas thermal conductivity, Btu/ ft h F K - tube or fin metal thermal conductivity, Btu/ ft h F L - tube length, ft m - factor defined in equation (A-11) n - fin density, fins/in. N w , Nd - number of tubes wide and deep ~Pg - gas pressure drop, in. wc q - heat flux, Btu/ sq ft h; subscript i stands for inside and 0 for outside. Re - Reynolds number S - Surface area, sq ft s - fin spacing, in. St - transverse pitch of tubes, in. S1 -longitudinal pitch, in. ~ T - log-mean temperature difference, F ti, tb, tf -temperature of fluid, wall and fin tip, F tg - average gas temperature, 1 and 2 refer to inlet and exit conditions. U - overall heat transfer coefficient, Btu/ sq ft h F-subscript 0 refers to outside and i to inside. W g - gas flow, pph ws - width of serration, in. TI - effectiveness of fins 11- viscosity of gas, Lb / ft h Pg- gas density,lb/cu ft

306


REFERENCES 1. Escoa fin tube manual, published by Escoa Corp. Tulsa, 1979 2. V. Ganapathy, "Applied heat transfer," Pennwell Books, Tulsa, USA 3. V. Ganapathy, "HRSG heat transfer," Heating, Piping, Airconditioning, March 90, Pg 99 4. V. Ganapathy,"Understanding and evaluating extended surfaces," paper presented at the 12th National Industrial Energy Technology conference, June 20, 1990, Houston, Texas. 5. V. Ganapathy, "How fin configuration affects heat transfer," Chemical Engineering, March 90, Pg 147 6. V Ganapathy, "Charts simplify spiral finned tube calculations," Chemical Engineering, Apri125, 1977 7. V. Ganapathy, "Charts help evaluate finned tube alternatives," Oil and Gas Journal, Dec 3, 1979 8. V. Ganapathy, "Estimate surface area of finned tubes," Chemical Engineering, May 27,1985, Pg 156 9. V. Ganapathy, "Chart speeds up fin weight estimation," Heating, Piping, Air-conditioning, March 1988, Pg 107

)

b

AI

---------------~ .. - - - - -

-

Appendix B

Low Temperature Corrosion In Chapter 1 we discussed two main areas of concern in waste heat boiler design, namely those due to high and low temperature corrosion. High temperature corrosion concerns as discussed in Chapter 1 may be alleviated by use of proper materials and design and by keeping the tube surfaces clean so as to prevent the formation of deposits of salts responsible for corrosion. Low temperature corrosion problems may be found in boilers at the cold end, in equipment such as economizers water heaters, and air heaters. They are caused by the condensation of corrosive acids on the surfaces of the tubes or duct work which operate below the acid dew point. A few widely used methods of dealing with this problem will be addressed in this section.

CAUSES AND CURES Whenever fossil fuels containing sulfur are fired in heaters or boilers, sulfur dioxide, and to a small extent sulfur trioxide, are formed in addition to C02 and water vapor. The 503 combines with water vapor in the flue gas to form sulfuric acid and condenses on heat transfer surfaces, which could lead to corrosion and destruction of the surfaces. This condensation occurs on surfaces that are at or below the dew point of the acid gas. Also when cooled below the water vapor dew point, C02 can combine with water vapor to form carbonic acid, which though weak, can attack mild steel. While thermal efficiency of the equipment is increased with reduction in exit gas temperature (or enthalpy), lower temperatures 307


308

than the acid gas dew point are not advisable for metallic surfaces in contact with the gas. In municipal solid waste fired plants, in addition to sulfuric acid, one has to deal with hydrochloric and hydrobromic acid formation. This article deals with methods for solving cold, or back end corrosion as it is called, with the most commonly used heat recovery equipment, namely economizers or water heaters. These are used to preheat feed water entering the system (Figure B-l) and operate at low metal temperatures, thereby increasing their susceptibility to corrosion by sulfuric, hydrochloric, hydrobromic and carbonic acid.

t Economizer Deaerator

Gas

Feed water Evaporator

Figure B-1. Economizer in a heat recovery boiler system.

Estimating the dew point of these acid gases is the starting point in understanding the problem of back end corrosion. Table B-1 gives the dew points of the various acid gases as a function of their partial pressures in the flue gas. Figure B-2 gives the dew point for sulfuric acid. As an example consider a typical glue gas having the following analysis: CO = 8%, H20 = 12%, N2 = 73%,502 = 0.02%, HCL = 0.015%, 02 = 6%,HBR = 0.01%, all by volume.

309

Appendix B - Low Temperature Corrosion

170r-----------------------------~

6

160

100 7

60

8 9

u Oct

140

10

~

5

130

i

11 r{

Example = 32 ppm

12 13 14

PS03

PH20 = 10% TDP = 150°C

15

16 17

120

18

110

~

________________

~

__________

19 ~20

Figure B-2. Dew Point of sulfuric acid as a function of partial pressure of S0;3 and water vapor. [Hydrocarbon Processing]

To compute the sulfuric acid dew point, one should know the amount of 503 in the flue gases. The formation of 503 is primarily derived from two sources. 1. Reaction of 502 with atomic oxygen in the flame zone. It depends on the excess air used and the sulfur content. 2. Catalytic oxidation of 502 with the oxides of vanadium and iron, which are formed from the vanadium in the fuel oil. It is widely agreed that 1 to 5% of 502 converts to 503. Hence the % volume in our case would be 4 ppm, assuming a 2% conversion. Using these numbers and after proper conversion and substitution in the equations in Table B-1, we have: dew point of sulfuric acid = 267°P, dew point of hydrochloric acid = 128°P, dew point of hydrobromic acid = 134°P and dew point of water vapor = 121°P.

" 310


HCI, HBr, HN03and S02 correlations were derived from vapor-liquid equilibrium data. 4 The H2S04 correlation is from reference 5. Hydrobromic acid: 1,000/Top = 3.5639 - 0.1350 In (PH20) 0.0398 In (PHB,) + 0.00235 In (PH20) In (PHB,) Hydrochloric acid: 1,000/Top = 3.7368 - 0.1591 In (PH2o) 0.0326 In (PHC1) + 0.00269 In (PH20) In (PHCI) Nitric acid: 1,000/Top = 3.6614 - 0.1446 In (PH20 )0.0827 In (PHNo3) + 0.00756 In (PH20) In (PHN03) Sulfurous acid: 1,000/Top = 3.9526 - 0.1863 In (PH30) + 0.000867 In (PS02) - 0.000913 In (PH20) In (PS02) Sulfuric acid: 1 ,OOO/Top = 2.276 - 0.0294 In (PH20) 0.0858 In (PH3S04) + 0.0062 In (PH20 ) In (PH2S04) Where: Top is dew point temperature (K) and P is partial pressure (mmHg). Compared with published data, the predicted dew points are within about 6K of actual values except for H2S04 which is within about 9K. Table B-lo Dew points of acid gases.

Hence, it is apparent the limiting dew point is that due to sulfuric acid and any heat transfer surface should be above this temperature (267°F) if condensation is to be avoided. There is a misconception even among experienced engineers that the gas temperature dictates the metal temperature of surfaces such as economizers. It is not so. To explain this, an example will be worked to show the metal temperature of an economizer with two different gas temperatures. Figure B - 8 shows this calculation. It can be seen that the water side coefficient is so high that the tube wall temperature runs very close to the water temperature in spite of a large difference in gas temperatures. Thus, the tube wall temperature will be close to the water temperature and the water temperature fixes the wall temperature and hence, the dew point. Some engineers think that by increasing the flue gas temperature the economizer corrosion can be solved; not so! It should

311


be noted also that the maximum corrosion rate occurs at a temperature much below the dew point (Figure B-3),

Peak corrosion Q}

~ c: o 'iii

Dew point

e o o

100

I

I

110 Wall temperature, ·C

130

Figure B-3. Corrosion rate as a function of wall temperature [Hydrocarbon processing]

Steam for preheating To process

jll~[E:conomizer

Feed water

Condensate To drum

Figure B-4. Steam to water exchanger preheats feed water.

Methods of dealing with cold end corrosion:

Basically there are two approaches used by engineers to combat the problem of cold end corrosion:

-----------~------------

312


A. Avoid it by using protective measures such as maintaining a high cold end temperature so that condensation of any vapor does not occur. B. Permit condensation of acid vapor or both acid and water vapor, thereby increasing the duty of the heat transfer surface, and use corrosion resistant materials such as glass, etc.

Methods of avoiding cold end corrosion: 1. Maintain a reasonably high feed water inlet temperature. If the computed dew point is say 250°F, a feed water of 250°F should keep the minimum tube wall temperature above the dew point. With finned heat transfer surfaces, the wall temperature will be slightly higher than with bare tubes. The simplest way would be to operate the deaerator at a slightly higher pressure, if the feed water enters the economizer from a deaerator (Figure B-l). At 5 psig the saturation is 228°F and at 10 psig it is 240°F. 2. In case the deaerator pressure cannot be raised, a heat exchanger may be used ahead of the economizer (Figure B-4) to increase the feed water temperature. It may be steam or water heated. 3. Figure B-5 shows a method for using an exchanger to pre-heat the water. The same amount of water from the economizer exit preheats the incoming water. By controlling the flow of the hotter water, one can adjust the water temperature to the economizer so that a balance between corrosion criterion and efficiency of operation can be maintained. 4. Hot water from either the economizer exit or the steam drum (Figure B-6) can be recirculated and mixed with the incoming water. The economizer has to handle a higher flow, but the exchanger is eliminated and a pump is added. Note that some engineers have the misconception that bypassing a portion of the economizer (Figure B-7) would solve the problemi not so. While bypassing, the heat transfer surface reduces the duty on the economizer and increases the exit gas temperaturei it does not help to increase the wall temperature of the tubes, which is the most important variable. A higher exit gas temperature probably helps the down stream ductwork and equipment, but not the

-----------------.----------

-----------~-~~-~~

313


economizer. One benefit, however, from bypassing is that steaming possibilities in the economizer are minimized.

Economizer

il!IC

Feed water Water to water heat exchanger

t

Gas

To drum

Figure B-S. Water to water exchanger preheats feed water.

Feed water Economizer

To drum

Recirculating pump

Figure B-6. Recirculation pump mixes hot water with feed water.

Economizer 1

Feed water

Feed water

Economizer 2

t

t

Gas

Economizer 2

TO

drum

A

Gas

To drum B

Figure B-7. Bypass arrangement for economizer. In "B" eco 1 is by-passed. This increases exit gas temperature and avoids steaming but does not solve dew point corrosion in eco 2.

314


Permitting condensation on surfaces: By using proper materials one can protect the heating surfaces from corrosion attack,if condensation is likely. This concept has now been extended to recovering the sensible and latent heat from the flue gases, thereby increasing the thermal efficiency of the system by several percentage points in what are called condensing heat exchangers. If flue gases contain say 10% by volume water vapor, by condensing even half of it, approximately 30 Btu/lb of flue gas can be recovered. This is nearly equivalent to a 120°F drop in gas temperature if sensible heat alone is transferred. A large amount of sensible and latent heat in the flue gas can be recovered if the gas is cooled below the water dew point. This implies that sulfuric acid, if present in the gas stream, will condense on the heat transfer surfaces as its dew point is much higher than that of water vapor. Borosilicate glass and teflon coated tubes have been widely used as heat transfer surfaces for this service. Glass is suitable for low pressures and temperatures (less than 450°F and 20 to 100 psig). However, presence of fluorides and alkalis is harmful to the glass tubes. One manufacturer of condensing heat exchangers uses teflon coated tubes. A thin film (about 0.015 in.) is extruded onto carbon or alloy steel tubes, and the surface is resistant to corrosion of sulfuric acid. Finned tubes can not be used as teflon cannot be extruded onto these surfaces. Hence, these exchangers will be larger than those with extended surfaces, however, the higher heat transfer rates with condensation process, improves the overall heat transfer coefficients and partly compensates for the lower surface area per linear foot of bare tubes. The high initial investment associated with condensing heat exchangers has to be carefully reviewed along with the energy recovered, fuel costs, etc. If the fuel cost is not high, then the payback period for this type of equipment may be long. Materials such as cast iron and stainless steels probably have better corrosion resistance than carbon steel, but still they are not corrosion proof. It is also felt by some that the higher thickness of cast iron is responsible for the longer life.

Appendix B - Low Temperature Corrosion.

315

The above material outline the importance of the dew point of acid gas and methods for dealing with the problem of condensation on heating surfaces such as economizers. Similar methods could be used for air heaters. The basic difference lies in the fact that the back end temperature is a function of both the gas and air temperatures. Steam air heating or air ~ypassing have been used 0 combat the problem of corrosion. Replaceable matrices and corrosion resistant materials such as enamels have been used at the cold end of regenerative air heaters.

The average wall temperature of a bare tube economizer is given by the simple equation:

tw = 0.5[tl + tg - U(tg - tl) (lIh o - 1!hl») Where:

hi = heat transfer coefficient inside tubes, Btu! ft 2 h of ho = heat transfer coefficient outside tubes, Btu! ft 2 h of tl = temperature of water inside tubes, of tg = temperature of gas outside tubes, of tw = average tube wall temperature, of U = overall heat transfer coefficient, Btulft2h °F

1/u = 1!hl + lIho, neglecting fouling and metal resistance, which are much smaller. Typically hi

= 1,000, ho = 15 and hence U = 14.77

Case 1: Determine tw when tg = 750°F and tl = 250°F tw = 0.5 [250 + 750 - 14.77 (750 - 250) (0.066 - 0.001») = 260°F Case 2: tg = 350°F, ti = 250°F tw = 0.5 [250 + 350 - 14.77 (350 - 250) (0.066 - 0.001») = 252°F Thus, for a variation of 400° F gas temperature, the tube wall temperature hardly changes by 8°F. Thus, the water temperature fixes the tube wall temperature.

Figure B-8. Determining tube wall temperature of economizer. [Hydrocarbon Processing]

CORROSION IN STACKS, DUCTS If a boiler stack or ductwork leading to the stack is uninsulated

then the average wall temperature will be lower than the gas


316

temperature due to the loss of heat from the casing, which depends upon the ambient conditions and the wind velocity. At lower loads the gas exit temperature from boilers will be lower and hence the problem becomes more acute. By adding external insulation,the heat loss can be minimized and the stack wall temperature can be maintained close to the gas temperature. The following methodology illustrates the procedure to compute the wall temperature of stack or duct. Stacks, ducts and scrubbers are designed so that the inside'wall temperature always remains above the dewpoint of any acid vapors passing through these units. This is because the condensed acids will attack many materials of construction, which would require the equipment to be made from costly alloys, ceramics or plastics. Designing equipment to avoid condensation requires estimating the inside-wall and stack temperatures (the latter being the temperature of the gas exiting the stack), and any heat losses through these walls, whether they are insulated or not. Generally, engineers use rule-of-thumb estimates for these temperatures. However, such guesses are sometimes inaccurate and as a result, structures are built that suffer needless acid corrosion. Here is a simple method to calculate these temperatures, as well as the heat losses, based partly on some equations previously developed by the author. Deriving the equations: The figure shows a typical temperature profile in a wall that is insulated. The heat loss is given by:

The temperature drop across the gas film is: (B-2)

where the gas-film heat-transfer coefficient is given by h = 2.44

WO· 8 e/di l .8

(B-3)


317

Insulation

,

,

, I

Stack or duct wall Figure 8-9. Temperature profile in ducts, stacks.

where: C = (Cp/I1)OA/k o.6

(B-4)

The duct-wall temperature drop is given by: twi - two

= Qdoln(do/di/24Km

(B-5)

The temperature drop across the insulation is: (B-6)

two - tc = QL/Ki

The effect of curvature [the heat-transfer surfaces are curved, since they are part of a cylindrical shape] is neglected, since ducts, stacks another such structures are generally large in diameter. The total heat loss from the duct (or stack or other such structure) is:

Qz =nd o HQ/12

= 3.14 doHQ/12

(B-7)

The temperature of the exiting gas is: tg2 = tgl -.Qe/WCp

(B-8)

318


Solution method: To solve the above equations: 1. Assume a gas-exit temperature, tg2 and calculate the average gas temperature, tg = 0.5(tgl + tg2) (B-9) 2. Assume an insulation-casing temperature, te. 3. Calculate Q, the heat loss, using Equation (B-V. 4. Calculate h, the film coefficient, and the various temperature drops, using Equations (B-3), (B-2) and (B-4) - (B-6). 5. Assume a value of the casing temperature, te. 6. Set tolerance for the difference between the assumed and calculated values of te. A difference of between 1-2°P will usually be accurate enough. If the tolerance is exceeded, repeat Steps 2-5. 7. Calculate the total heat loss from Equation (B-7) and the exit-gas temperature, tg2, via Equation (B-8) 8. If the calculated and assumed values of tg2 are not within the tolerance, repeat the process, starting from Step 1; otherwise, the calculations are finished. Although this method may seem tedious, the calculations are easily done on a scientific calculator, and the assumed and calculated values quickly converge to within a reasonably accuracy. A few trials are all that is usually needed. A computer program can easily be created to perform the calculations; the author will supply his version upon request. To illustrate the method, here is an example: Example 1; Insulated stack: Flue gases at 110,000 lb/h and 423°P enter a 48-in.-I.D. stack, 50 ft long and 1 in. thick. If the ambient temperature is 67°P and the wind velocity is 125ft/min, determine the wall and casing temperatures if the stack is covered with 2 in. of mineral-fiber insulation. Por flue gases at 420 o P, Cp = 0.266 Btu/(lb)(OF), Jl = 0.058 Ib/(ft)(h), and k = 0.023 Btu/(ft)j(h)/( OF). The gas temperature drop in the insulated stack/ducts is relatively low, typically 2-5°P. Thus, assume that the exit-gas temperature is 420o P: 1. Assume that the casing temperature is 90 o P. Then: Q =0.173 x 0.9(5.5 4 - 5.27 4) + 0.296(550-527)1.25[(125 + 69)/6910.5 =47.3

-


- - - - - - - - - --

------

- - -_ _ _ _ _ _ _

__

~

319

Btu/(ft2)(h), and h= 2.44 x 110,000°.8(0.266/.058)°.40.023°.6/481.8 = 4.74 Btu/(ft2)(h)(OF). 2. The thermal conductivity of the fiber (from manufacture's tables) is 0.30 and 0.42 Btu-in./(ft2 )/(h)/( oF) at 200 and 400 o P, respectively. At 250 oP, the value is 0.33. 3. The temperature drop across the insulation = 47.3 x 2/0.33 = 287°P. The average insulation temperature is 232°P. Por the next trial, the thermal conductivity has to be estimated at this temperature. 4.Stack-wall temperature drop = 47.3 x 50 In(50/48)/24 x 25 = 0.16°F. Note: A thermal conductivity of 25 Btu/(ft)(h)(OF) is used for carbon steel. 5. Gas film drop = 47.3 x (50/48)/4.74 = 1O.4°F. 6. Thus, the corrected casing temperature = 422-10.4-0.16-287 = 124°F. Since 124°P is higher than the assumed value of 90o P, another iteration must be tried. It can be shown that assuming a value of 92°P gives good results. Using this value Q = 52.5 Btuj(ft2)(h) 7.The total heat loss = 3.14 x 50 x 52.5 x 50/12 = 34,300 Btu/h. 8. The gas temperature-drop in the duct = 34,300/110,000/0.266 = 1.2 of. The final results: Temperature drops: gas-film = 11.5 of, duetwall = 0.2 OF, and insulation = 31JOF. Example B-2: Let us now check the stack wall temperature if no insulation is used. Solution: 1. Let the gas temperature drop = 20 0 Pi as the heat loss is higher, the gas temperature drop will also be higher. The average gas temperature is = 413°P. 2. Let the casing temperature = 250o P. This will be checked and corrected later. 3. Heat loss from the casing is: Q = .173 x .9 x [7.1 L 5.27 4J + .296 x (710-527)1.25 x [(125 + 69)/69Jo.5 =610 J3tu/sq ft h. 4. Temperature drop across gas film = 610/4.74 = 129°F. 5. Temperature drop across the stack wall = .0034 x 610 = 2 OF. 6. Hence two = 413-129-2 = 282°F versus 250°F assumed. Hence another round of iteration is required.

7~~~~

_ _...(

320


It may be shown that a wall temperature of 266°F balances the equations. Heat loss =679 Btu/ sq ft h. The gas temperature drop =3.14 x 50 x 679 X 50/(110000 X .265) = 15°F. This agrees with what we assumed in step 1 and hence no further iterations are necessary. Since insulation calculations particularly those with multiple layers involve iterative calculations, the author has developed a program to solve for the heat loss and the temperature profile across multi-layer insulation. This is presented in Figure B-10, followed by two examples.

HEAT -LOSS CALCULATIONS THROUGH MULTI-LAYERED INSULATION The heat loss from the casing of any surface can be found from

Q = 0.173iTa4) + 0.296 (ts_ta)1.25 [(V + 69)169Jo.5 This equation is obviously difficult to handle manually, but the Basic program presented here provides a quick calculation of heat loss from refractory or insulation-lined pipes or flat surfaces. It also calculates intermediate temperatures and casing outer temperature. The program can handle any number of layers, with the heat loss between layers being given by Q

= t1T/L

(B-l0)

For flat surfaces, L is the insulation or refractory thickness. For cylindrical surfaces, L = 0.5Do In(Do/Di), where Do= outer diameter of the layer and Di = inner diameter of the layer. Thermal conductivity K for each layer is calculated at its mean temperature. K is input at two given temperatures, and the program interpolates for the actual mean K using a linear relationship. Quick converging logic is used to compute the final results. A trial value of casing temperature is assumed, and intermediate temperature and heat loss are calculated. The intermediate

--------~--

---------

---~-----------~~


321

CLS:KEY OFF:CLEAR DIM A(64) ,A$(IS) ,L(IS) L=O:DO=O 10 PRINT"INSULATION PERFORMANCE-PROGRAM ••• FLAT SURFACES OR PIPES":PRINT" "

20 PRINT"This program can handle any number of layers-information on conductivit y of each layer at any two temeperatures is needed":PRINT" " 30 INPUT"NAlIE OF PROJECT-OATE(up to 30 characters)~";PROJ$:PRINT" " 40 INPUT"pipe outer dia in inch-for flat surface input 0="/00: IF 00=0 THEN B$=" flat surface"ELSE B$=" pipe" 50 00(1)=00 60 PRINT" " 70 INPUT"number of layers,hot face temp, ambient temp, wind vel in fpm,surf emissa: ";P(A(1) ,H,V,J:HF=A(l) 80 PRINT" " 90 PRINT"Input Insulation data starting from the hot surface-thermal conduotivit y in BTU in/ft2 h":PRINT" " 100 FOR 1=1 TO P 110 INPUT"Name(up to 10 characters) ,thickness-in=" ;A$,L 120 A$(I)=A$:L(I-l)=L 130 PRINT" " 140 INPUT"templ,cond l,temp2,cond 2~"/M(I) ,K(I) ,N(I) ,0(1) 150 IF 00=0 THEN A(lS+I)=L :GOTO 170 160 A(lS+I)=.S*LOG( (DO+2*L) /00) :00=D0+2*L 170 A(4S+I)=(K(I)-Q(I» / (M(I)-N(I» :A(30+I)=K(I)-A(4S+I)*M(I) 180 PRINT" ":NEXT I 190 A(P+l)=H+200:U=A(1) 200 Z=A(P+l)-H 210 X=(U-A(P+l) )/p 220 W=.01*(A(P+l)+460) :Y~.01*(H+460) 230 Q=.173*J* (WA4-Y A4)+93.6*(W-Y)A1.2S*SQR( (V+69) /69) 240 XX-.Ol* (460+HF) :QMAX-.173*. 9* (XX A4_Y A4 )+93.6* (XX-Y) A1. 2S*SQR( (V+69) /69) 250 FOR 1=1 TO P 260 A(P+l-I)=A(P+2-I)+X 270 T=.S* (A(P+I-I)+A(P+2-I» :G=A(3l+P-I)+A(46+P-I)*T 280 IF 00=0 THEN FA=l ELSE FA=OO 290 300 300 310 320 330 340 350 360 370 380 390 400 410

R=A(P+2-I)+Q*FA*A(16+P-I) /G IF ABS(R-A(P+I-I» rn nr

0.161--_---,;L+-_ _ _+-_ _ _+-_ _ _+-_ _---l

0.14 t - - + - - - t - - - - t - - - - t - - - - t - - - - - j

::r: (I)

~ O:l 0

=:

... (I)

0.15 L...._~--J_ 200 o

___L_--L_

400

__'__

_ ' _ _........_

600

Temperature, of

......._..I.____J

600

1,000

0.13 "--___"'-----'_--L_--'-_ o 200 400

__'__........_-'-_-'--_.l..----l

600

Temperature. of

BOD

1,000

0(I)

g

"'

~

Figure E-l. Air

Figure E-2. Oxygen

0.22 ...-..--_......,_ _ _

~

_ _ _-.-_ _ _-.-_ _---,

2 . 4 , . . - - - - , - - - - . -_ _r--_ _-.._ _----,

0.21 1-_+_-+_ _ _--1f-_ _..:.P-=a"-lra~m..:.e"'t.::.er...:s.:..:.:..P",re=+ss..:.u:::re..::.:..,:::al:::m..:.s"""-l' 2.3 t----I----+---+---+--~

0.20 Hr--~d----+----+----~----..,."e:::2.2r-...:p...:a=r:::am~m=efrs=:..:.p..:.re=s=s=u=re~·fa=tm=s=.-~~~~-_+-_ _~

0.19 1-\-_-+_+-_ _ _+-_ __

2.1 t----h~h'~---+---+----I 0.18 1---~"+__-::..--+-_r__4-----1---___I

0.17 f----+--+-+----+----+-----1

0.18 '--_-'-_.1.-_-'--_.1.-_.1.-_-'--_-'--_.1.-_.1.-----1 1,000 o 200 800 400 600

2.0

I-T-I-.H4-----f---+---I-----I

1.9 E..-----I_......L_...L_L---1._-L_.L_L-L_-' o 200 400 600 BOO 1,000

Temperature, of

Temperature, of

Figure E-3. Nitrogen

Figure E-4. Hydrogen

O~~----~------~----~~-----r------' 1.0 0.9 0.8 0:;

-

0.22 H . . - - - / - - - - + - - - + - - - - + - - - - I

-

0.21 1 - - - - \ - - ; - - - _ + - - - - + - - - - + - - - - 1 Parameters: Pre8lllure, atms.

0.6

0.20 I---rl---_+---+----+----:::;~

0.5

... 0 ...0~

0.19 1----1--"'...::--1--0.4

Q.

()

'"

0

0.3

:.!

i\\

\\

II

0

0.2

0.17 1:-------II----:::;."......I--7.c:------i~---;---_'_I

~

\~ r---..... i-

0.1

0.18 1-~--;---_+-7""'--7I""'_7'c---+---_I Parameters: Pressure, alms.

-

V

a

V\

""

'~ ~

I

200

-

250

I

400

I I

600

I----"...'-----+-----+----+------l

0.151----1----1----1-------11----;

I

800

0.18

1,OqO

o

200

400

600

600

Temperature, of

Temperature, of

FiQ"ure E-S. Carbon dioxide

Figure E-6. Carbon monoxide

1,000

4.0 3.0 2.0

----

1.0

\

~

\

0.8

...ci

-l!:.

~

0

Co

'":;;" ""II 0

0.6 0.5

\ \ \ \

1\ \ \

Parameters: Pressure, alms.

0.4 r\ \ \ 0.3

-

\\

0.2

0.1 0.08

--

0.06 0.05 0.04

"'-."

\ '&"-av,~""-

-

-- --

I~ ~ ~...

'

0.30 hf----t----t-------1r--------1------1

-

0.03

o

I

I

200

I

400

I 600

Temperature, OF

Figure B-7. Sulfur dioxide

I 800

0.25

1,000

'--_'--_'--_'----.JL-.--.JL-.--.JL-.--.JL-.---l_---l_---l

o

200

400

600

Temperature, OF

Figure B-S. Methane

800

1,000

/ Charts give heat-transfer coefficients considering pressure effect / 2.5...-_ _--.._ _ _-.-_ _-;r-_ _,-_ _--,

0.36 0.34

1--\----1------+---- -- --- -.--

0.32 0.30

I+--\---\--i----+---=-.-~~~ -~--=t~~~

0.28

~

\..

,;

...

0.26

ci

0.24

Parameters: Pressure, alms .

Il

~ I

Paramelers: Pressure. alms.

~~ ci

0

.K

...~ -'"

0.22

•

()

o

0.40 I--+_ _~ --~+-----t------;0.35 \---''---+- ' - - - - ; . . - - - - t - -

0.20

0.18

0.30

\-----+--........--t--"'....- - t

0.25

I--_ _ _~.

:E I»

0.20 1-----+--'''=---+

fIl

;0.16

::r: It>

0.15 h-----I---:~""__l.

~

O:l

g. 0.14 L---I---L---L_-L_-L_...l.-_..L-_L-_L...---l o 200 400 600 800 1,000 Temperature, of

Figure E-9. Nitric oxide

t!i

o

200

400

600

800

Temperature, ° F

1,000

I ,I

~

~ 0

~

Figure E-IO. Nitrogen peroxide

II

1:1

!

j ';'

Figure E-13. Air

Figure E-14. Oxygen

/ Charts give heat-transfer coefficients considering pressure effect / 2.5...-_ _--.._ _ _-.-_ _-;r-_ _,-_ _--,

0.36 0.34

1--\----1------+---- -- --- -.--

0.32 0.30

I+--\---\--i----+---=-.-~~~ -~--=t~~~

0.28

~

\..

,;

...

0.26

ci

0.24

Parameters: Pressure, alms .

Il

~ I

Paramelers: Pressure. alms.

~~ ci

0

.K

...~ -'"

0.22

•

()

o

0.40 I--+_ _~ --~+-----t------;0.35 \---''---+- ' - - - - ; . . - - - - t - -

0.20

0.18

0.30

\-----+--........--t--"'....- - t

0.25

I--_ _ _~.

:E I»

0.20 1-----+--'''=---+

fIl

;0.16

::r: It>

0.15 h-----I---:~""__l.

~

O:l

g. 0.14 L---I---L---L_-L_-L_...l.-_..L-_L-_L...---l o 200 400 600 800 1,000 Temperature, of

Figure E-9. Nitric oxide

t!i

o

200

400

600

800

Temperature, ° F

1,000

I ,I

~

~ 0

~

Figure E-IO. Nitrogen peroxide

II

1:1

!

j ';'

Figure E-13. Air

Figure E-14. Oxygen

0.13 1.25

0.12

1.20

1.15

:;; d

0.11

"-

if d

~

;

C.

..."

d

~

.l

Waste heat boiler deskbook

Boiler Maker

American Indian Law Deskbook

Materials selection deskbook

Materials Selection Deskbook

Financing energy projects deskbook

Boiler operator's handbook

The Pot-Boiler

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Waste Treatment: Reducing Global Waste

The Battery and the Boiler

Boiler Operator's Guide, Fourth Edition

Waste Land

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Waste Management

Waste Land

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Waste Land

Boiler Operator's Exam Preparation Guide

Heat

Heat

Heat

Heat

Heat

Heat

Heat

Heat

Heat

Heat

Heat

Waste heat boiler deskbook

Boiler Maker

American Indian Law Deskbook

Materials selection deskbook

Materials Selection Deskbook

Financing energy projects deskbook

Boiler operator's handbook

The Pot-Boiler

Boiler Control Systems Engineering

Boiler operator's handbook

Waste Treatment: Reducing Global Waste

The Battery and the Boiler

Boiler Operator's Guide, Fourth Edition

Waste Land

Waste Materials

Waste Management

Waste Land

Waste Land

Waste Land

Boiler Operator's Exam Preparation Guide

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