practical thermal design of air-cooled heat exchangers

8/10/2019 Practical Thermal Design of Air-Cooled Heat Exchangers

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About the Author

Rajiv Mukherjee is a consultant in unfired heat transfer based in New Delhi, India. Hehas 36 years of experience in the thermal design, revamping, and troubleshooting of air-cooled and shell-and-tube heat exchangers, and considerable experience in the design ofheat exchanger networks. He has written several articles in reputed journals and pre-sented many papers at technical symposia. Rajiv has also served as faculty for numerouscourses on heat exchanger design and operation, energy conservation, and heat exchanger

networks, and presently teaches an intensive in-house refresher course on the design andoperation of heat exchangers that can be offered at any plant site or office location aroundthe world. He is an honors graduate in chemical engineering from Jadavpur University,Kolkata, India.

In his spare time, Rajiv enjoys reading (Swami Vivekananda and Kahlil Gibran are bigfavorites), writing, and listening to music. He lives in New Delhi with his wife, Kalpana.Their daughter, Shilpi, and her husband, Bappa, presently live with their sons, Sohum andShivum, in Tokyo, Japan.

Rajiv may be contacted by e-mail at [email protected] or by telephone at0091-11-2551 8281 or 0091-98711-20126.

R. Mukherjee

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P REFACE

My desire to write this book was precipitated by the absence of such a practical book.Recent heat exchanger design literature has been predominantly occupied by proceedingsof conferences. There is no book in the market that explains the logic of heat exchangerthermal design and gives practical suggestions and recommendations for actually design-ing industrial heat exchangers. So, having written my earlier book, Practical Thermal

Design of Shell-and-Tube Heat Exchangers , which received a fairly good response, I de-cided to write a sequelone on air-cooled heat exchangers.

The theoretical aspects of single-phase heat transfer and condensation have been verywell presented in several books. So, what was really required was a practical how-to-design book with numerous worked-out examples or case studies to embellish or illustrate a

particular technique, facet, or style of design. The thousands of air-cooled heat exchangerdesigns that I have been associated with over the last three decades have provided numerousexamples. They say that one picture is more eloquent than a thousand words. If you extendthis logic, one appropriate illustration by a case study is eminently more didactic than a long

dissertation on a particular subject.This book has been written in the same style, language, and format as the one on shell-

and-tube heat exchangers. For the sake of convenience, both English and metric units have been used throughout the book. There are 26 case studies, all aimed at embellishing,illustrating, reinforcing, or demonstrating a feature, rationale, or methodology of designelaborated or advocated in the text. Not only are the case studies based on the HTRIsoftware, the entire book is founded on the platform of HTRI know-how, which has becomea way of life for me for almost three decades.

Being a practical book, theory is limited to a bare minimum, and the accent is onfundamentals, on design logic, on the interplay of parameters, on cause and effect, onunderstanding why things happen the way they do. For example, why does a lighthydrocarbon condenser tend to have only four rows of tubes, whereas a heavy hydrocarbonliquid cooler tends to have more rows of tubes? Or why do we choose 1/2 in. (12.7 mm)high fins in certain situations but 5/8 in. (15.875 mm) high fins in others? Or why is the

process fluid break point between an air-cooled heat exchanger and its downstream trimcooler related to the design ambient temperature? And many, many others.

This book has been written primarily for the heat exchanger thermal designer. However,I think it will also be useful to process engineers, a significant part of whose routine job is tospecify heat exchangers. This book has not been written in an esoteric style for this veryreason. Since operating aspects are also often discussed, I trust it will be of interest to plantoperation specialists as well.

It is my fond hope that even B.S. and M.S. chemical and mechanical engineering

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students will find the book interesting, informative, and useful. I still remember when I wasan undergraduate studentI used to long for more practical, real-life information aboutindustrial practice. If one considers that many engineering graduates end up working in thechemical process industries, there may be a lot of merit in adding such a flavor to heattransfer in the university curriculum, as indeed it is to all other fields of human learning. The

juxtaposition of industrial equipment design practice with basic theory will go a long way in

making the subject more interesting and meaningful.The thermal design of air-cooled heat exchangers is a fascinating activitysometimeseven more so than that of shell-and-tube heat exchangersfor the simple reason that thereare more variables: even the coolant (air) flow rate is a variable! This book will have servedits purpose if it can inspire the reader to consider the thermal design of air-cooled heatexchangers as a joyous activity rather than a mundane chore.

I will be grateful for any feedback regarding any aspect of this book, and the same may be sent to [email protected] or [email protected].

R. Mukherjee Heat Transfer Consultant

New Delhi, India

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Dedication

This book is dedicated to my parents, who would have been proud to seethis work. And that is an understatement. To my dear wife, Kalpana, whohas been supporting and inspiring me for over three decades now; to ourdaughter, Shilpi, our son (-in-law), Bappa, and their sons Sohum and

Shivum; but most importantly, it is dedicated to the reader, whose ap- probation and appreciation would make all the toil worthwhile.

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Acknowledgments

I am indebted to Almighty God for having given me the education, intelligence, opportunity,strength, and fervor to write this book.

I am also indebted to all those from whom I learned the design of air-cooled heatexchangers over the years, especially to Wim Bos and Peter van der Broek of Lummus

Nederland B.V., who in 1971, initiated me on the path of air cooler design.I will always be grateful to Cindy Mascone, ex-technical editor at Chemical Engineering

Progress, and her one-time compatriot, Gail Nalven, who led me to believe that I could write

a book. I will never forget Gails words, You do not know how prolific you are! As forCindy, she is the finest editor I have ever worked with.This book might not have been possible without the wonderful exposition of air-cooled

heat exchanger technology by Heat Transfer Research, Inc. (HTRI). My long experience inthe field of air-cooled heat exchangers has been very largely honed on the platform of HTRI,whose software I have been using since 1974.

I am grateful to HTRI, Begell House, Hudson Products Corp., Moore Fans Ltd., TEMA(Tubular Heat Exchanger Manufacturers Association), and Cal Gavin Ltd. for permission touse some of their diagrams and photographs in this book. These have been dulyacknowledged where they appear.

I am indebted to Bill Begell who decided to publish this book, and to all the people atBegell House who were responsible for its production. Special thanks go to DonnaThompson who did a splendid job of copyediting this book, as she did with the previous oneon shell-and-tube heat exchangers. Donna, I have enjoyed working with you again.

How can I forget my good friend Graham Polley in the UK? It was he who led me toBill in the first place.

I am thankful to Geoff Hewitt, who is the editor of the present series of books, forhaving readily accepted this book into his fold.

Thanks are also due to my wonderful friend Sam Chapple of Edmonton, Canada, whoguided me on some important issues in the text.

I must also express my gratitude to another good friend, Lalit Shingal, who helped mewith the reproduction of many diagrams that appear in the book.

What we are able to accomplish in our lives, whether professionally or otherwise, is theresult of the Lords grace and the encouragement and support we receive from myriadsources. This book is therefore truly a collaborative effort, and the credit belongs to thehuman fraternity at large, rather than to any individual.

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C HAPTER 1

Introduction

Although air is much more freely available than water and costs nothing, process coolinghas historically been accomplished by cooling water. This is attributable to the muchlower cost of cooling by water, thanks to its substantially higher thermal conductivity andlower temperature. However, with increasing shortages of cooling water and a conse-

quent increase in its cost, air cooling has become more and more popular. Today, air-cooled heat exchangers (ACHEs) are a common sight in the chemical process industries(CPIs).

The first cost of an ACHE is much greater than that of a water-cooled heat exchangerfor the same heat duty, but its operating cost is usually much less. The operating cost withwater cooling comprises the cost of the initial raw water itself, makeup water, treatmentchemicals, apportioned cost of the cooling tower, and of course the pumping cost. For air-cooled heat exchangers, the operating cost is only the cost of the power required to make theair flow across the tube bundles. Thus, on an overall cost basis, ACHEs often compare quitefavorably with water-cooled heat exchangers.

The design of ACHEs comprises two distinct activities, namely, thermal design andmechanical design. In thermal design the basic sizing of the heat exchanger is accomplished,whereas in mechanical design the thicknesses and precise dimensions of the variouscomponents are determined and a bill of materials is produced. Detailed engineeringdrawings are then prepared based on which actual fabrication drawings are made. In this

book, as the title suggests, we shall talk principally about thermal design.With the availability of sophisticated software, there has been an undue dependence on

them as black boxes, without the designer being truly in control of the design process andunderstanding the nuances of design. A proper and sound understanding of the fundamental

principles and interplay of parameters is essential in order to produce an optimum design.The principal purpose of writing this book is to help the heat exchanger thermal designerattain such an understanding.

Presently, there is no book available on practical ACHE thermal design. This book is based on the authors experience of over 36 years in the thermal design of ACHEs for thechemical process industry, and reflects many real-life situations that were far fromstraightforward. This book has been written in a structured, logical, and didactic manner, andspecial effort has been made at bringing out the interplay of parameters for a thoroughunderstanding of basic issues.

As Example is better than precept , several case studies are presented in this book in

order to vividly bring out a particular methodology, principle, or practice that has beenadvocated. The reader is invited to run these examples with further variations in the parameters being examined, in order to develop a comprehensive understanding.

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It is well known that the thermal design of ACHEs is still largely an enigma, with farfewer engineering and fabricating companies practicing the trade than the thermal designshell-and-tube heat exchangers. This is really quite surprising, considering that thermaldesign of ACHEs is simpler and more straightforward than that of shell-and-tube heatexchangers! This book will have served its purpose if it encourages more companies toovercome this diffidence and take up the thermal design of ACHEs.

Now, coming to the individual chapters themselves, Chapter 2 dwells on the advantagesand disadvantages of air cooling, while Chapter 3 discusses the optimization of air and watercooling. In some instances, only cooling by air need be employed, whereas in others onlycooling by water is adequate. However, in the vast majority of cases that fall between thesetwo extremes, cooling by both air and water is favorable.

Chapter 4 gives a detailed rundown of the various components and constructionalfeatures of ACHEs, since a good understanding of the same is vital to the thermal design ofthis equipment. This chapter will also be of considerable interest to mechanical designers ofACHEs, since it explains the implications of several constructional features on thermaldesign.

Chapter 5 discusses various basic concepts that form much of the foundation ofknowledge for ACHE design. The simultaneous optimization of airside and tubesidecalculations is certainly not an easy task. However, with the help of logical explanations,arguments, and case studies, the design methodology is made easy to understand and apply.

Chapter 6 is on the thermal design of condensing ACHEs. After a brief classification ofcondensers and a brief account of the mechanisms of condensation, practical guidelines forthermal design are discussed. These include isothermal, narrow-range and wide-rangecondensation, the effect of pressure, the handling of desuperheating and subcooling, nozzlesizing, and the handling of condensing profiles and physical property profiles.

In Chapter 7, with the help of numerous case studies, optimization of ACHEs isdemonstrated vis--vis tube OD, fin height, fin spacing, number of tube rows, fan powerconsumption, tube pitch, and the number of tube passes.

In Chapter 8, physical properties and heat release profiles are discusses at length. Thereader is offered guidance on how to feed heat release profiles, a matter that is not as simpleas it may appear.

Chapter 9 explains why overdesign is provided, and elaborates on the modalities ofoverdesign for single-phase and condensing services.

After reviewing the various categories of fouling and the parameters that affect it,suggestions are offered in Chapter 10 on how to specify fouling resistance. Comprehensive

guidelines are then suggested and analyzed in order to minimize fouling.Chapter 11 is on the control of ACHEs, where various methods of control are discussedin detail. Unlike water-cooled shell-and-tube heat exchangers, ACHEs offer very goodcontrol on the process.

Chapter 12 deals with operating problems in air-cooled heat exchangers. Various potential problems and ways to avoid them are discussed for both the tubeside and theairside cases.

In Chapter 13, many special applications are elaborated on, including combinedservices, recirculation ACHEs, humidified ACHEs, tube inserts, variable finning density,natural convection, and vacuum steam condensers.

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C HAPTER 2

Advantages and Disadvantages of Air Cooling

Let us take a look at the advantages and disadvantages of air cooling as compared to watercooling

2.1 Advantages of Air Cooling

Air cooling offers many advantages over water cooling. We have already discussed thecost advantage of air cooling over water cooling. Besides this advantage, the use of air asa cooling medium eliminates certain inherent disadvantages associated with water cool-ing:

a ) The location of the cooler and thereby a plant is independent of a source of watersupply such as a river or a lake, or even a sea; hence, the plant can be located inany geographic area. To use water as a cooling medium, however, the plant hasto be located at a site close to a large natural body of water such as a lake, river,or sea. This could very easily entail a penalty in terms of transportation of rawmaterials or finished products.

b) Air coolers are far more environment friendly since thermal and chemical pollu-tion of the source of water are eliminated. In once-through cooling water, such aswith sea water, warmer water is returned to the body from which the water isdrawn, thereby leading to a rise in temperature of that body of water. This has adirect adverse effect on the life and longevity of the aquatic plant and animalspecies inhabiting the body of water. In recirculating cooling water systems(which are the norm), the outlet warm water is cooled by a cooling tower so as to

eliminate this increase in temperature of the discharge water with its associatedadverse effect on aquatic life.

c) Maintenance costs are lowered considerably since frequent cleaning of the waterside of coolers (necessitated by fouling such as scaling, biofouling, sedimenta-tion, etc.) is eliminated.

d ) The installation is simpler since water piping and water pumps are eliminated.Another advantage with air coolers is that they continue to operate (although at a reducedcapacity) by natural convection even when there is a power failure. In some cases, thiscan be as much as 6070% of the design duty. In the case of water cooling, however, a

power outage usually means a plant shutdown, which results in direct loss in production.Yet another advantage with air cooling is that air-cooled heat exchangers offer very

effective control of the process fluid outlet temperature (and thereby the heat duty) through

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the following various means:a) switching fans on/off

b) use of two-speed motorsc) use of autovariable fansd) use of louverse) use of variable-speed drives

These will be discussed in detail in Chapter 11.On the other hand, water cooling does not render an effective means of control of the

process fluid outlet temperature (and thereby the heat duty). This is because of two reasons:(a) the MTD is predominantly controlled by the cold end temperature difference which doesnot change with a reduction in the cooling water flow rate and (b) the cooling water filmresistance is a very small percentage of the overall resistance to heat transfer. Consequently,a reduction in the cooling water flow rate has a negligible effect on the performance of awater-cooled cooler.

In an air-cooled heat exchanger, however, a reduction in the air flow rate has a muchmore pronounced effect on the performance of the cooler because both the MTD and theoverall heat transfer coefficient change significantly. This is because the airside heat transfercoefficient controls the overall heat transfer resistance quite strongly, and the MTD alsovaries significantly with a change in the air flow rate, and thereby the outlet air temperature.This is illustrated in the following case study.

C ASE STUDY 2.1: E FFECT OF R EDUCTION OF AIR F LOW R ATE

A quantity of 1,700,000 lb/h (771,115 kg/h) of hot water is to be cooled from 174F(78.9C) to 140F (60C), representing a heat duty of 57.665 M Btu/h (14.53 M kcal/h).The allowable pressure drop of hot water is 10 psi (0.7 kg/cm 2) and its fouling resistanceis 0.001 h ft 2F/Btu. For cooling by air, the design air temperature is 107F (41.7C). Forcooling by water, the cooling water inlet temperature is 93F (33.9C), its allowable pres-sure drop is 10 psi (0.7 kg/cm 2), and its fouling resistance is 0.002 h ft 2F/Btu (0.0004 hm2C/kcal).

The air-cooled heat exchanger design was prepared first and its principal construction parameters are indicated in Table 2.1a. The total air flow rate was 7,700,000 lb/h (3,493,700kg/h). To demonstrate the effect of a reduction in the total air flow rate, the same waschanged to 6,900,000 lb/h (3,130,000 kg/h) and then to 6,100,000 lb/h (2,767,000 kg/h). The

principal performance parameters for all three total air flow rates are shown in Table 2.1b. Itwill be seen that there is a significant reduction in the overdesign with a lowering in the airflow rate, from 6.2% in the first case to 2.9% in the second case, and finally to 13.7% inthe third case. This is due to an appreciable reduction in both the MTD [from 33.2F(18.4C) to 28.6F (15.9C)] and the airside heat transfer coefficient [from 178.8 Btu/h ft 2F(873 kcal/h m 2C) to 163.2 Btu/h ft 2F (797 kcal/h m 2C)].

Next, the water-cooled hot water cooler design was then prepared, the principalconstruction parameters of which are shown in Table 2.1c. The cooling water flow rate is3,400,000 lb/h (1,542,200 kg/h). In order to demonstrate the effect of a reduction in thecoolant flow rate, the same was reduced by the same amount as the air flow rate in the case

of the air-cooled heat exchanger design. The principal performance parameters of all threedesigns are shown in Table 2.1d. It will be seen that the reduction in overdesign is far lessthan that of the air-cooled heat exchanger, from 7.3% in the first case to 3.5% in the second

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Table 2.1a: Principal construction parameters of air-cooled hot water cooler

No. of bays in parallel 4 No. of bundles per bay 2 No. of tubes per row 46 No. of tube rows no. of tube passes 5 2Tube/fin material CS/AluminumTube OD thickness 1 in (25.4 mm) 12 BWG (2.77 mm)

Tube length, ft (m) 34 (10.36)Fin height fin thickness 5/8 in (15.875 mm) 0.016 in (0.4 mm)Fin density 11 per in (433 per meter)Transverse pitch, in (mm) 2.625 (67)Bundle width, ft (m) 10.2 (3.11)Total bare tube area, ft 2 (m2) 16,047 (1491)Total extended area, ft 2 (m2) 381,080 (35,416)

No. of fans per section fan dia., ft (m) 2 14 (4.27)Motor power, HP (kW) 30 (22.4)

Table 2.1b: Effect of variation in air flow rate on performance of air-cooled hot water cooler7,700,000

(3,492,700)6,900,000

(3,130,000)6,100,000

(2,767,000)Total air flow rate, lb/h (kg/h)

Air outlet temperature, F (C) 138.1 (58.9) 141.7 (60.9) 146.3 (63.5)Static pressure, in. WC (mm WC) 0.43 (10.9) 0.35 (8.9) 0.3 (7.6)Tubeside pressure drop, psi (kg/cm 2) 2.1 (0.15) 2.1 (0.15) 2.1 (0.15)

Tubeside 838 (4092) 838 (4092) 838 (4092)Airside (bare tube) 178.8 (873) 171.1 (835) 163.2 (797)

Heat transfer coefficient,Btu/h ft 2F (kcal/h m 2C)

Overall 114.9 (561) 111.7 (545) 108.3 (529)MTD, F (C) 33.2 (18.4) 31.2 (17.3) 28.6 (15.9)Overdesign, % 6.2 2.9 13.7

Absorbed power, HP (kW) 26.5 (19.7) 19.7 (14.7) 14.2 (10.6)

Table 2.1c: Principal construction parameters of water-cooled hot water cooler

TEMA Type AEL (Fixed tubesheet) No. of shells 1Shell ID, in. (mm) 45 (1143)

No. of tubes no. of tube passes 1578 1Tube OD thickness, in. (mm) 0.75 (19.05) 14 BWG (2.108 mm)Tube length, ft (m) 20 (6.1)Tube pitch, in. (mm) 1.0 (25.4) triangularType of baffles baffle cut orientation Double segmental horizontalBaffle spacing, in (mm) no. of tube rows overlap 19 (483) 6Connections: shellside/tubeside, nominal, in (mm) 16 (400)/20 (500)Heat transfer area, ft 2 (m2) 6083 (565)

case, and finally to 0.1% in the third case. This is because although there is a similarreduction in the overall heat transfer coefficient as in the case of the air-cooled heatexchanger, the drop in the MTD is far less.

2.2 Disadvantages of Air Cooling

Let us now consider the limitations of air-cooled heat exchangers as compared to water-cooled heat exchangers. This is a comprehensive list and only some of them will be pre-sented for a particular situation or application.

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Table 2.1d: Effect of variation in cooling water flow rate on performance of water-cooledhot water cooler

3,400,000(1,542,000)

3,000,000(1,360,800)

2,700,000(1,224,700)

Cooling water flow rate, lb/h (kg/h)

Cooling water outlet temperature, F (C) 110 43.3) 1123 (44.6) 114.4 (45.8)Shellside 3.8 (0.27) 3.8 (0.27) 3.8 (0.27)Pressure drop, psi (kg/cm2)Tubeside 3.0 (0.21) 2.3 (0.16) 1.9 (0.134)Shellside 1539 (7514) 1539 (7514) 1539 (7514)Tubeside 1065 (5200) 970 (4736) 897 (4380)

Heat transfer coefficient,Btu/h ft 2F (kcal/h m 2C)

Overall 185.6 (906) 182.5 (891) 179.8 (878)MTD, F (C) 54.8 (30.4) 53.8 (29.9) 52.8 (29.3)Overdesign, % 7.3 3.5 0.1

High initial cost

Since air has a much lower thermal conductivity and specific heat than water, an air-cooled heat exchanger has a much lower overall heat transfer coefficient and conse-

quently a much larger bare tube heat transfer area than a water-cooled heat exchanger forthe same heat duty. Thus, the initial cost of an air-cooled heat exchanger (cooler hard-ware only) is considerably more than that of a water-cooled heat exchanger. However,should the process fluid require ordinary carbon steel and the cooling water require admi-ralty brass or copper-nickel or duplex special stainless steel, the cost disadvantage of theair-cooled heat exchanger reduces appreciably. The more superior the required coolingwater metallurgy, the lower the cost disadvantage of the air-cooled heat exchanger.

Costly winterization

In cold climates, extensive winterization arrangements have to be incorporated to negoti-ate subzero temperatures, thereby increasing the first cost even further. Winterizing anair-cooled heat exchanger means rendering it operable even under winter conditions; thatis, the process stream will not freeze or congeal. These arrangements include heating bysteam coils, incorporating louvers to reduce air flow rate, reversing air flow, and even anelaborate recirculation system that will be discussed in detail in Chapter 13.

Lower economical approach temperature

Because of the very low overall heat transfer coefficient, an economical approach temperature between the outlet process fluid and the ambient air is generally in the range of

1820F (1012C) whereas in water-cooled heat exchangers, this approach temperaturecan be as low as 57F (34C). The logic here is that a very low MTD, coupled with thetypically low overall heat transfer coefficient of an air-cooled heat exchanger, will lead toan enormous heat transfer area and thereby first cost. The relatively large MTD for aircooling as compared to that for water cooling will neutralize to a large extent the inher-ently higher first cost of an air-cooled heat exchanger. This disadvantage is overcome byhaving air cooling followed by trim cooling with water; that is, by off-loading the last

part of the cooling duty to water cooling. However, should cooling water be unavailableat a particular site, this could represent a major limitation.

Larger plot areaBecause of the larger heat transfer area, an air-cooled heat exchanger requires a consid-erably larger plot area than a water-cooled heat exchanger. However, this disadvantage is

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overcome by locating an air-cooled heat exchanger on a pipe rack, so that no valuable plot area is wasted. Process equipment such as shell-and-tube exchangers, accumulators,and pumps are usually located at grade level under the pipe rack.

Fan noise

Due to the low specific heat and density of air, air-cooled heat exchangers have to force

large quantities of air across the tube bundles, resulting in a high noise level. However,with the improvement in fan technology, including the development of low-noise fans,the noise level can usually be restricted to permissible levels.

Limitations in plant layout

Air-cooled heat exchangers cannot be located near large obstructions, such as buildings,since air recirculation can set in and affect the cooling performance adversely. However, thisneed not represent a major limitation since a judicious plant layout can usually circumventthis problem.

Availability of fewer vendorsSince air-cooled heat exchangers employ relatively sophisticated design technology, thenumber of vendors who can offer air-cooled heat exchangers is far smaller than the numberof vendors who can offer water-cooled heat exchangers, which are shell-and-tube heat ex-changers.

Problems associated with laminar flow

It is well known that viscous liquids yield laminar flow and a rather poor heat transfer coef-ficient when flowing inside tubes due to the boundary layer separation. Such liquids yield

considerably higher heat transfer coefficients when flowing on the outside of heat exchangertubes, due to the much higher turbulence with a staggered tube arrangement. Therefore, forcooling viscous liquids, air-cooled heat exchangers become even more expensive due to theextremely low tubeside heat transfer coefficient and thereby the higher heat transfer area.Besides, such liquids often have a high pour point, which necessitates the use of an air recir-culation arrangement, thereby increasing the first cost even more. However, this situationcan be remedied to a large extent by the use of wire-fin tube inserts. This is addressed laterin Chapter 13.

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C HAPTER 3

Optimization of Air and Water Cooling

In applications where the process outlet temperature is relatively low, cooling by only airmay not be feasible. For example, cooling a light hydrocarbon liquid to 113F (45C) byair may not be feasible at a site where the design ambient temperature is 107.6F (42C).In such cases, a combination of cooling by air followed by trim cooling (cooling by wa-

ter) has to be adopted.Occasionally there are services where cooling by air may not be economically viable.Thus, at a site where the design ambient temperature is 107.6F (42C), cooling by air maynot be viable for a naphtha stabilizer condenser service wherein the inlet/outlet temperaturesare 122/113F (50/45C), due to the very low temperature difference between the processstream and the air. Here, only cooling by water should be employed.

In other services, where both the inlet and the outlet temperatures of the process streamare relatively high, e.g., 248F (120C) and 140F (60C), only air cooling can be employed.

However, such services as described above are quite rare and invariably the process inlettemperature will be fairly high, such as 212F (100C), whereas the process outlet

temperature will be rather low, such as 113F (45C). In such services, the most optimumarrangement is to have an air-cooled heat exchanger, followed by a trim cooler employingcooling water in a shell-and-tube heat exchanger.

Thus, for some services, only air cooling should be employed; for some others, acombination of air and water cooling should be employed; while for the balance, only watercooling should be employed. The vast majority of services in the chemical services fall inthe middle category where both air and water cooling should be used.

The optimum temperature break point between air and water cooling (that is, thetemperature at which a process fluid should leave an air-cooled heat exchanger and enter awater-cooled heat exchanger) should be established by overall economics for every project,

because it will depend on equipment cost (air-cooled and water-cooled heat exchangers),cost of water (total cost, as discussed in Chapter 1), and the cost of power. Generallyspeaking, this optimum temperature is about 2732F (1518C) greater than the designambient temperature.

It is important to realize here that for a combination of air and water cooling, the air-cooled heat exchanger will handle the major heat duty (75% or more of the total heat duty),thereby resulting in a considerable reduction in the cooling water flow rate.

When using a combination of air and water cooling, it is usually best to design the air-cooled heat exchanger for a somewhat lower ambient temperature (than what would be used

if there were no trim cooling), and then to design the trim cooler for the process fluidtemperature, which would be discharged by the air-cooled heat exchanger at the higherambient temperature. This is because with the higher ambient temperature, the decrease in

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Table 3.1a: Heat exchanger service for Case Study 3.1Fluid Stabilizer overheadFlow rate, lb/h (kg/h) 127,870 (58,000)Temperature in/out, F (C) 153.3 (67.4)/116.6 (47)Operating pressure, psia (kg/cm 2a) 172 (12.1)Total allowable pressure drop, psi (kg/cm 2) 8.5 (0.6)Fouling resistance, h ft 2 F/Btu(h m2 C /kcal) 0.00195 (0.0004)

Heat duty, MM Btu/h (MM kcal/h) 18.72 (4.717)Density in/out, lb/ft 3 (kg/m 3)

the cost of the trim cooler will be less than the increase in the cost of the air-cooled heatexchanger, thereby resulting in a higher overall cost. This is illustrated by the following casestudy.

C ASE STUDY 3.1: SELECTION OF DESIGN AMBIENT T EMPERATUREFOR AIR AND W ATER C OOLING

Consider the stabilizer condensing duty specified in Table 3.1a. The allowable pressuredrop of 8.5 psi (0.6 kg/cm 2) is for both the air-cooled heat exchanger and the trim cooler.The maximum and minimum ambient temperatures at the site were 100.4F (38C) and64.4F (18C), respectively. The inlet temperature of the cooling water was 89.6F(32C), its fouling resistance was 0.002 h ft 2F/ Btu (0.0004 h m 2oC/kcal), and its permit-ted pressure drop was 10.7 psi (0.75 kg/cm 2). The optimum break temperature betweenair and water was established as 131F (55C). Thus, an air-cooled heat exchanger was to

be designed for condensing the stabilizer overhead from 153.3F (67.4C) to 131F(55C) and a trim cooler (using water) was to be designed to condense the balance vaporfrom 131F (55C) to 116.6F (47C). Since there had to be a trim cooler, it was decided

to consider the design ambient temperature as 95F (35C).An air-cooled heat exchanger was designed, its principal construction and performance parameters being as per Table 3.1b.

By carrying a performance run, it was established that if the ambient temperature were100.4F (38C), the stabilizer overhead outlet temperature would be 133.3F (56.3C).Therefore, the trim cooler was designed for this heat duty; that is, for condensing theuncondensed vapor from 133.3F (56.3C) to 116.6F (47C). A single TEMA (TubularHeat Exchanger Manufacturers Association) type AES shell having a heat transfer area of2233 ft 2 (207.5 m 2) was found to be adequate. The principal construction and performance

parameters of this design are indicated in Table 3.1c. Now, let us consider what would happen if the air-cooled heat exchanger were to be

designed for an ambient temperature of 100.4F (38C) and the trim cooler for condensingthe stabilizer overhead from 131F (55C) to 116.6F (47C). The bare tube area of the air-

1.643 (26.31)/1.382 (22.14)Vapor properties Viscosity in/out, cp 0.01/0.0095

Specific heat in/out, Btu/lb F (kcal/kg C) 0.45/0.43Thermal conductivity in/out, Btu/h ft F (kcal/h m C) 0.0134(0.02)/0.0128 (0.019)Density in/out, lb/ft 3 (kg/m 3) 31.28 (501)/31.84 (510)Viscosity in/out, cp 0.1/0.131Specific heat in/out, Btu/lb F (kcal/kg C)

Liquid properties

0.71/0.7

Thermal conductivity in/out, Btu/h ft F (kcal/h m C) 0.0605 (0.9)/0.0659 (0.098)

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Table 3.1b: Principal construction and performance parameters of air-cooled heatexchanger for Case Study 3.1

Air inlet temperature, F (C) 95 (35) 100.4 (38)1. No. of bays 1 12. No. of bundles per bay 2 23. No. of tubes per bundle per row no. of tube rows 38 6 44 4

4. No. of tube passes 2 25. Tube OD thickness, in. (mm) 0.984 (25) 0.098 (2.5)6. Tube length, ft (m) 41.0 (12.5)7. Fin OD thk. tube pitch, in. (mm) 2.24 (57) 0.016 (0.4) 2.638 (67)8. Fin density, no./in. (mm) 11 (433)9. Total bare tube area, ft 2 (m2) 4714 (438.1) 5459 (507.3)

Tubeside 364.7 (1781) 336.7 (1644)Airside 188.9 (922.3) 184.8 (902.4)

10. Heat transfercoefficient, Btu/h ft 2F (kcal/h m 2

C) Overall 98.3 (479.9) 95.1 (464.1) ()

Tubeside, psi (kg/cm 2) 5.3 (0.37) 4.17 (0.29)11. Pressure drop

Airside, in. (mm) WC 0.5 (12.64) 0.45 (11.5)12. Airflow rate, lb/h (kg/h) 1,984,100

(900,000)2,160,500(980,000)

13. MTD, F ( C) 33.1 (18.4) 29.0 (16.1)14. Power per fan, HP (kW) 28.7 (21.4) 28.7 (21.4)15. Total fan power, HP (kW) 57.4 (42.8) 57.4 (42.8)16. Fan diameter, ft (m) 14 (4.267) 15 (4.571)17. Overdesign, % 11.0 8.6

cooled heat exchanger increases from 4714 ft 2 (438.1 m 2) to 5459 ft2 (507.3 m 2). The principal construction and performance parameters of this design are shown in Table 3.1b. Notice that the fan diameter also goes up from 14 ft (4.267 m) to 15 ft (4.57 m) in order to provide the minimum 40% bundle coverage. However, since its heat duty is lower, the trimcooler heat transfer area decreases from 2233 ft 2 (207.5 m 2) to 1891 ft 2 (175.7 m 2).

The above results are summarized in Table 3.1d. It will be seen that when the air-cooledheat exchanger is designed for 100.4F (38C), (a) the total heat transfer area of the air-cooled heat exchanger and the trim cooler is significantly higher and (b) the heat transferarea of the air-cooled heat exchanger too is higher. Therefore, since air-cooled heatexchangers cost significantly more than shell-and-tube heat exchangers for the same bare

tube heat transfer area, the installed cost of this option is considerably higher. Consequently,it will be more economical to design the air-cooled heat exchanger for an ambienttemperature of 95F (35C) and the trim cooler for a stabilizer overhead temperature thatwould be the outlet from the air-cooled heat exchanger when the ambient temperature is themaximum (100.4F or 38C).

From the above example, we see that is more economical to design an air-cooled heatexchanger for an ambient temperature that is somewhat lower than the maximum ambienttemperature, and to pass on the shortfall in the heat duty of the air-cooled heat exchanger tothe trim cooler when the ambient temperature is the maximum expected.

Evidently, the above approach is possible only for air-cooled heat exchangers that are

followed by a trim cooler. If an air-cooled heat exchanger is not followed by a trim cooler, itwill obviously have to be designed for the maximum expected ambient temperature.

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Table 3.1c: Principal construction and performance parameters of water-cooled heatexchanger for Case Study 3.1

1. Shellside inlet temperature, F (C) 133.3 (56.3) 131 (55)2. Heat duty, Btu/h (kcal/h) 5.85 (1.475) 4.89 (1.232)3. Type of exchanger Floating head (TEMA AES)4. Shell ID, in. (mm) 30.9 (785) 28.7 (730)

5. Number of tubes 560 4746. Total heat transfer area, ft 2 (m 2) 2233 (207.5) 1891 (175.7)7. Tube pitch, in. (mm) layout angle 1.024 (26) square8. Number of tube passes 4 4

Baffle spacing, in (mm) 10.8 (275) 10.8 (275)9. BafflingBaffle cut (dia), % 25 25Cross-flow 5.61 (1.71)

Further Reading

1. Maze, R.W., 1975, Air Cooler or Water Tower: Which for Heat Disposal, Chem. Eng .,Jan. 6.

5.45 (1.66)10. Shellsidevelocity, ft/s (m/s) Window flow 5.09 (1.55) 5.35 (1.63)11. Shellside pressure drop, psi (kg/cm 2) 2.13 (0.15) 2.42 (0.17)12. Tubeside velocity, ft/s (m/s) 4.82 (1.47) 4.76 (1.45)

13. Tubeside pressure drop, psi (kg/cm 2) 1.9 (0.58) 1.84 (0.56)Shellside 241 (1177) 259 (1263)Tubeside 996 (4865) 996 (4865)

14. Heat transfercoefficient, Btu/h ft 2

F (kcal/h m 2 C) Overall 109 (533) 112.4 (549)15. Overdesign, % 8.3 7.9

Table 3.1d: Overall comparison of the two cases for Case Study 3.1

Air-cooled heat exchanger designed for95F (35C), trim cooler designed for duty

corresponding to air-cooled heatexchanger performance at 100.4F (38C)

Air-cooled heat exchangerdesigned for 100.4F

(38C), trim cooler designedfor corresponding duty

1. Bare tube area ofair-cooled heatexchanger, ft 2 (m 2)

4714 (438.1) 5459 (507.3)

2. Bare tube area oftrim cooler, ft 2 (m 2)

2233 (207.5) 1891 (175.7)

3. Total bare tubearea, ft 2 (m 2)

6947 (645.6) 7350 (683)

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C HAPTER 4

Construction Features of Air-Cooled Heat Exchangers

4.1 Introduction

Before we start discussing the thermal design of air-cooled heat exchangers, it will be

necessary to have a detailed look at the constructional features.The principal components of an air-cooled heat exchanger are:Tube bundleFans and drivePlenum chamberStructure

Before discussing these principal components, let us first consider some important termsin air-cooled heat exchanger parlance.

Tube bundle: A tube bundle is an assembly of headers, tubes, tube supports, andframes (Fig. 4.1).

Bay or section: A bay or a section is composed of one or more tube bundles served bytwo or more fans, complete with structure, plenum, and other attendant equipment(Fig. 4.2). Thus, a bay is the smallest independent part of an air-cooled heat ex-changer that is repeated for multibay or multisection units.

Unit: A unit is composed of one or more tube bundles in one or more sections for an in-dividual service (Fig. 4.3).

Bank: A bank or battery of air-cooled heat exchangers comprises one or more sectionsor units arranged on a continuous structure (Fig. 4.4).

Fig. 4.1 Tube bundle Fig. 4.2 A bay or a section

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Front elevation

Fan

Bundle

Plan view

Bay Bundle

Unit A Unit B

Bank

4.2 Tube Bundle

As defined above, a tube bundle is an assembly of tubes, headers, tube supports, and sideframes. We shall now take these up one by one.

Fig. 4.3 A unit

Fig. 4.4 A bank or battery of air-cooled heat exchangers

4.2.1 Finned tube

Because of the extremelylow heat transfer coeffi-cient of air, which is adirect result of its lowthermal conductivity, it

becomes imperative toemploy extended surfaceon the airside. The com-monest and cheapest formof extended surface is thefinned tube (Fig. 4.5).

The base tube may be

of any commercially avail-able material suitable forthe process, based on

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considerations of corrosion, pressure, and temperature, but it is usually of carbon steel orstainless steel. The fins are invariably of circular cross section, although fins of ellipticalcross section have also been used. Although occasionally of steel for high-temperatureservices, fins are normally of aluminum because it has the most favorable thermalconductivity-to-cost ratio, besides having good cold-working properties. Since these fins can

be as thin as 0.016 in. (0.4 mm), it is common to pack in 11 fins per inch (433 fins per

meter). However, several vendors prefer to limit the fin density to 10 per inch (394 permeter) on account of airside fouling, while others are known to employ even 12 per inch(472 per meter). Limiting the fin density to 11 per inch (433 per meter) appears to be a verysensible practice. The manufacturers of finned tubes find it economically practical to limitthe fin density to between 7 and 11 per inch (276 to 433 per meter).

Since the coefficient of linear expansion of aluminum is about twice that of carbon steel,a gap resistance between the tube and the fin material develops and this increases as theoperating temperature is increased, due to the increased difference of the coefficients oflinear expansion. Thus, depending on the type of bond between the tube and the fin,maximum operating temperatures have been established. These are specified for the various

types of finned tubes in Sections 4.2.1.1, 4.2.1.2, 4.2.1.3, and 4.2.1.4.The standard tube OD is 1 in., although 1-1/4 in., 1-1/2 in., and even 2 in. are employed

where the tubeside pressure drop is controlling. The corresponding metric values are 25 mm,32 mm, 38 mm, and 50 mm. The standard fin heights are 3/8 in. (10 mm), 1/2 in. (12 mm),and 5/8 in. (16 mm), with the latter two being far more popular than the former. Such astandardization becomes necessary, considering an economical production program offinning machines.

There are various types of finned tubes, as described below.

4.2.1.1 Single L-footed finned tube

This is a circular fin wrapped helically around the tube under tension (Fig. 4.5a). Fullcoverage of the base tube by the L-foot offers good protection against atmospheric corro-sion. However, these fins tend to become loose with time, thereby causing an appreciabledeterioration in the airside performance due to the air gap between the tube and the fins(see Ref. [1]). Consequently, their use is generally restricted to applications where the

process inlet temperature is less than 248F (120C).However, even for such applications, their use is generally not very popular because the

airside performance of these finned tubes is more likely to deteriorate with time due to theloosening of the fins. An exception is in corrosive marine atmospheres, such as on offshore

platforms, where the good protection provided by these tubes against atmospheric corrosionof the base tubes makes them superior to conventional grooved fins. However, in suchapplications, it would be advisable to derate the effectiveness of these tubes to cater to the

probability of loosening of the fins from the base tube.

4.2.1.2 Double L-footed finned tube

These tubes offer an even better coverage of the base tube (Fig. 4.5b) but since they aremuch more costly, may be preferred only in extremely corrosive atmospheres. The upperlimit of process fluid inlet temperature for these finned tubes is 338F (170C). However,claims of bond improvement and improved corrosion protection have been questioned bysome users. The use of these finned tubes is quite rare in the chemical process industries.

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4.2.1.3 Grooved or embedded or G type finned tubes

In these tubes, the fin is embedded in the tube by first plowing a groove in the tube wall,and then stretching the fin material into the groove under sufficient tension to achievespecified bond strength (Fig. 4.5c). Evidently, G-finned tubes require a thicker tubewall than L-footed tubes. As per clause 4.1.11.3 of API (American Petroleum Institute)

661, the minimum tube wall thickness is 0.083 in. (2.1 mm) for carbon steel and low-alloy steel, and 0.065 in (1.65 mm) for stainless steel. For embedded fin tubes, however,this thickness is reckoned from the bottom of the groove. Hence, embedded fin tubeshave to be thicker than L-footed fin tubes by the groove depth (which is usually theequivalent of one gauge).

Grooved finned tubes can tolerate process fluid temperatures of up to 752F (400C)due to their strong fin bond (usually no fin-bond resistance penalty is applied because it is

(e)

(c) (d)

(a) (b)

Fig. 4.5 Various types of finned tubes: (a) single L-footed; (b) double L-footed; (c) grooved (G-type); (d) bimetallic; (e) extruded

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considered negligible) and they are a very commonly used type of finned tubes. They canalso withstand cyclic operation without any loss of fin-tube contact. The disadvantage ofthese tubes is that their base tube material is exposed to the atmosphere and, therefore, theiruse in aggressive atmospheres (such as marine applications) is not recommended.

4.2.1.4 Bimetallic finned tubes

Bimetallic finned tubes have G-fins embedded in an outer tube of aluminum, which isstretched over the base tube (Fig. 4.5d). In applications where the process fluid is at high pressure and/or is corrosive, thereby requiring the use of an expensive alloy, it may becheaper to use bimetallic finned tubes with a thin-wall inner tube, than to use a heavy

base tube of high alloy with G-fins. The upper temperature limit for these finned tubes is554F (290C).

4.2.1.5 Extruded finned tubes

These are basically a double tube construction. An outer tube with a large wall thickness,usually aluminum, is swaged over an inner base tube and extruded into high fins, all in oneoperation. An uninterrupted bond over the entire tube length is formed, thereby protectingthe base tube from the outside (Fig. 4.5e). The fin bond is thus considered superior to that ofany other interference fit fin type.

Due to the high fin material usage and the more expensive manufacturing process, thesefinned tubes have the highest cost. Their applications are similar to those of bimetallic finnedtubes.

For a detailed discussion of fin bond resistance and the maximum permissible operatingtemperature for the above types of high-finned tubes, the reader is referred to [1]. For adetailed discussion on specifying the right fin type for different applications, please see [2].

4.2.2 Headers

Headers serve to introduce the hot fluid into the tubes and collect the cooled fluid at theend of the flow passage. They carry the inlet and the outlet nozzles and other connections(such as vents and drains), as well as the pass partition plates required for multipass ex-changers.

Headers are so arranged that movement within the side frame is possible to containthermal expansion. As per API 661, clause 6.1.6.1.2, if the temperature difference betweenthe inlet to one pass and the outlet from the adjacent pass is greater than 200F (111C), splitheaders (Fig. 4.6.) or other means of restraint relief are to be employed. This condition also

applies when the maximum operating temperature is greater than 350F (177C).There are various types of header

construction, each having specificadvantages and disadvantages. Themost common header types are asfollows.

Fig. 4.6 Split header

4.2.2.1 Plug type header

This is the most commonly used type

of construction (Fig. 4.7a) and comprises a rectangular, welded box withinlet nozzles in the top plate and outlet

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nozzles in the bottom plate. The tubes are either welded to the tubesheet or expanded intotube holes in the tubesheet. The plug hole opposite each tube in the plug sheet allows me-chanical cleaning of each tube and plugging in case of leakage.

This type of header is relatively cheap and can be used for pressures up to 3000 psig(211 bar g). The disadvantage of plug headers is that for frequent cleaning of tubes for dirtyservices (fouling resistance greater than 0.00195 h ft 2F/Btu or 0.0004 h m 2C/kcal),

removal of the large number of plugs becomes time consuming and therefore costly.Thus, plug headers are preferred for clean and moderate-to-high pressure applications.

4.2.2.2 Cover-plate type header

For fouling services, cover-plate type headers (Fig. 4.7b) are preferred since it is mucheasier to remove cover plates than the numerous plugs of plug type headers. However, athigher pressures, this header type becomes expensive since the cover plate becomes verythick. Hence, cover-plate type headers are usually not used for pressures in excess of 569

psig (40 bar g). For easy removal, cover-plate type headers are equipped with jackscrewsand lifting lugs.

4.2.2.3 Manifold headers

Round manifold headers (Fig. 4.7c) are used in very-high-pressure applications. Thetubes are welded to the manifold by means of stubs and will usually have return bends,i.e., the tubes will be U-tubes.

Due to manufacturing limitations, the number of tube rows per manifold is restricted to

(c)

(a)

(d)

(b)

Fig. 4.7 Various types of headers: (a) plug; (b) cover plate; (c) manifold; (d) bonnet

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Before expansion After expansion

(a) Expanded (b) Welded

Fig. 4.9 Tube-to-tubesheet joint: (a) expanded; (b) welded

fan manufacturer. Hence, the designer should have access to fan curves giving informa-tion regarding volume (or mass) of air, static pressure, absorbed power, and noise. Somefan manufacturers even furnish computer software to aid the designer in proper fan selec-tion. Also, some contemporary software packages incorporate fan performance curves ofthe leading fan manufacturers in their software.

The fan diameter for air-cooled heat exchangers usually varies between 6 ft (1.83 m)and 18 ft (5.49 m), although fans having smaller and larger diameters are employed inspecial circumstances. A fan consists of two basic components: the hub and the blades.

4.3.1.1 The hub

The hub (see Fig. 4.10) is the component that is mounted on the fan shaft and the bladesare mounted on the hub. Hub material may be cast iron, cast aluminum or fabricatedsteel. Manufacturers usually conduct static and dynamic balancing of the hub in the shop.

The hub is usually of two types:a ) Manually adjustable, where the blade angle can be altered only when the fan is

stationary.b) Autovariable, where the hub carries a device (usually a pneumatic controller) that

can alter the blade angle even while the fan is in motion in order to control airflow. Control is usually effected by means of a signal from a TIC (temperature

indicator controller) responding to theoutlet temperature of the process fluid.

4.3.1.2 Blades

Blades (Fig. 4.10) can either be of metal(usually aluminum) or FRP (fibreglassreinforced plastic). Plastic blades aresuitable only for temperatures of up to158F (70C); thereby representing alimitation for induced-draft air-cooledheat exchangers.

Fan performance (air flow rate andstatic pressure) is determined by theFig. 4.10 Axial fan (Courtesy Moore Fans Limited)

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number of blades, tip speed, blade angle, and blade width. The effect of a change in the tipspeed of a fan on its performance is dramatic. The volume of air flow varies directly with thetip speed; the pressure varies as the square of the tip speed and the horsepower varies as thecube of the tip speed. The tip speed is normally limited to 200 ft/s (61 m/s) since noise

becomes excessive at higher values.An increase in the number of blades of a fan increases its ability to work under pressure.

Thus, the tip speed of a six-blade fan can be reduced to deliver the same volume of air, ascompared to that of a four-blade fan. However, this can be carried too far in that as thenumber of blades increases beyond six, multiblade interference may actually reduce theefficiency of the fan, since each blade works in the disturbed wake of the preceding blade.Therefore, the number of blades has to be selected carefully by the fan vendor.

All the blades of a fan should be set at the same angle for smooth operation. Usually, the blade angle is set between 12 and 27. This is because performance deteriorates at lowangles and becomes unstable at high angles. The volumetric flow rate varies as the bladeangle tangent ratio (BART) to the 1/3 power and pressure to the 2/3 power. Therefore, HPvaries directly with BART. The blade angle should also be carefully selected by the fan

vendor.A fan with a wider blade width can be operated at a lower tip speed to achieve the same

performance. Consequently, fans with wider blades operate less noisily. This feature isexploited by fan vendors who offer special low-noise fans.

4.3.1.3 Fan laws

Fans of the same basic design and dimensions operate theoretically in accordance withcertain fan laws. In practice, these laws do not apply exactly because of design considera-tions and manufacturing tolerances, but they are useful in estimating the approximateoutputs of similar fans of different diameters and speeds, as applied to normal air deliv-ery. These laws can be expressed as follows:

a ) The volume of air flow varies as rpm and as (fan diameter)b) The pressure developed varies as (rpm) and as (fan diameter)c) The power absorbed by the fan varies as (rpm) and as (fan diameter) 5

It is important to note, however, that these laws apply to the same operating point on afan characteristic. They cannot be used to predict other points on the fans curve.

These laws are most often used to calculate changes in the flow rate, pressure, and power of a fan when the size or rotational speed is altered. They assume no change in fanefficiency for any given point on the fan curve, when there is a change in speed.

4.3.1.4 API specifications

Section 4.2.3 of API 661 stipulates the following for fans and fan hubs:1) There should be at least two fans along the tube length; however, a single-fan de-

sign can be agreed to in exceptional circumstances (such as very small units) be-tween purchaser and vendor. This two-fan requirement is apparently based onconsiderations of reliabilityshould one fan stop functioning due to a belt

breakage or other reason, the other fan will be running so that the unit will con-tinue to run, albeit at a somewhat lower heat duty. Furthermore, at lower loads

and at cooler ambient temperatures, one fan may be stopped for better control ofthe process outlet temperature as well as for power savings. Another reason isthat even when autovariable fans are used, both fans need not be autovariable;

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one autovariable fan and one manually adjustable fan can achieve the necessarycontrol.

2) Fans should be of the axial-flow type and each fan should occupy at least 40% ofthe tube bundle face area served by it. This is to ensure a reasonably good distri-

bution of air across the face of the tube bundles.It may be added here that when the width of a bay is about one-half of the

tube length, it represents an ideal situation because the fans then see a squarecross section of the bay since there are two fans in a bay. For example, if thewidth of a bay is 20 ft (6.095 m) and the tube length 40 ft (12.191 m), each of thetwo fans will deliver air across a 20 ft (6.095 m) 20 ft (6.095 m) cross section.A 14 ft (4.27 m) diameter fan with a cross-sectional area of 153.9 sq. ft (14.3 sq.m) will just fall short of the 40% requirement specified above, so that the use of a15 ft (4.57 m) diameter fan will be necessary. The reader should note that a tubelength-to-bay width ratio of 2:1 is not a must, but it is desirable that it does notexceed this ratio substantially. Furthermore, when this ratio tends toward 3:1, adesign with three fans per section will perform much better than one with twofans.

3) The fan dispersion angle (see Fig. 4.11) should not exceed 45 at the centerline ofthe tube bundle. This requirement is also based on proper air distribution.

4) The radial clearance between the fan ring and the fan tip should not exceed 0.5%of the fan diameter or 0.75 in. (19 mm), whichever is less; fan stalling may occurat larger clearances. Compliance with this requirement is strongly recommendedsince it yields significantly better fan performance but is not expensive to imple-ment. Table 4.1 gives a clearer picture of the effect of tip clearance on the per-formance of a 5 ft diameter fan. Notice that until a 0.5% tip clearance, the fans

4 5

m a x .

4 5

angledispersionMax.

Fig. 4.11 Fan dispersion angle

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Table 4.1: Effect of fan tip clearance on its performance

Tip clearancein. (mm)

Tip clearance as %of fan diameter

Fan performanceefficiency (%)

0.074 (1.9) 0.12 99.50.15 (3.8) 0.25 98.50.2 (5.1) 0.33 97.40.3 (7.6) 0.5 95.0

0.4 (10.2) 0.67 91.00.464 (11.8) 0.77 88.0

perform quite well; but thereafter, there is a rapid deterioration in efficiency.5) Fan tip speed should not exceed 200 ft/s (61 m/s) unless approved by the pur-

chaser, and in no case exceed 266 ft/s (81 m/s). As stated earlier, the noise levelof a fan increases sharply beyond this tip speed.

Some additional recommended design guidelines are as follows:1) The minimum distance between the plane of the fan and the tube bundle (that is,

the plenum height) should be one-half the fan diameter for forced draft units andone-third the fan diameter for induced draft units. These requirements are formaintaining favorable aerodynamics and thereby superior performance of fans.

2) For both forced draft and induced draft units, the height of the fan ring should beat least one-sixth of the fan diameter.

3) Air seals should be provided between tube bundles and between tube bundles andthe plenum chamber in order to minimize air bypassing. Any gap wider than 0.4in. (10 mm) should be considered excessive and therefore sealed.

4.3.2 Fan drivesThe power required by the fans to move air across the tube bundles is provided by anelectric motor, steam turbine, gas or gasoline engine, or hydraulic motor, with the electricmotor being the overwhelming choice. Polyphase, squirrel-cage totally enclosed fan-cooled (TEFC) induction type motors are usually used. Steam turbine drivers are requiredto be as per API Standard 611.

The power is transmitted from the motor (or the turbine) to the fans through direct drive,V-belt drive, HTD (high-torque drive), or gear drive. A direct drive (Fig. 4.12a) has the fanshaft directly connected to the driver and is usually used with fan diameters of 5 ft (1.53 m)

or less, as well as drives of 5 HP (3.73 kW) or less.A V-belt drive (Fig. 4.12b) is used when the rpm of the fan is less than the rpm of thedriver. V-belt drives may be used with motor drives rated 30 HP (22 kW) or less, as per API661 clause 4.2.8.2.10.

HTD may be used with motor drives rated 50 HP (37 kW) or less, as per API 661 clause4.2.8.2.11. Unlike flat and V-belts, HTD belts do not rely on friction for its pulling power.HTD belts utilize a revolutionary new tooth design that substantially improves stressdistribution and higher overall loading.

HTD belts do not stretch due to wear, are corrosion resistant, and operate at reducednoise levels. The belts are capable of transmitting higher torque at lower speed, thus

improving the horsepower capacity of toothed belts. The belts do not depend on thickness todevelop great tensile strength. There is no loss of speed caused by belt creep or slippage aswith flat and V-belts.

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HTD systems have unusually high mechanical efficiency. Further, transmissionefficiency is not lost with use. The belt construction ensures very little heat buildup sincefriction is not required to pull the load. Since belt tension is reduced, significant powersavings are thus achievable, especially on larger HP installations.

Maintenance is simple. No adjustments are required due to stretch or wear. HTD beltsare ideal where proper maintenance is difficult or where downtime could prove to be

extremely expensive.For electric motors rated above 50 HP (37 kW), right-angle gear drives (Fig. 4.12c)must be used (API 661 clause 4.2.8.3.1). All steam turbine drivers must employ right-anglegear drives (API 661 clause 4.2.8.3.3)

4.3.3 Plenum chamber

The air delivered by a fan isdistributed to (forced draft) orcollected from (induced draft)the tube bundle by a plenumchamber that consists of duct-work in the form of a rectangu-lar box (Fig. 4.13) or acone/rectangle transition piece(Fig. 4.14). For forced draftunits, the plenum chamber can

be square (or rectangular) orconical whereas for induceddraft units they are invariably

conical. A partition is provided between fans, and the gap be-tween the tube bundle and the

partition plate should not exceed20 mm.

For forced draft units, the plenum chamber has a conicalinlet at the bottom to reduce inletlosses. When low-noise fans areemployed, the conical inlet isreplaced by a bell mouth.

4.4 Configurationof ACHEs

Horizontal configuration

Air-cooled heat exchangers(ACHEs) are usually configured

in the horizontal dispositionsince maintenance becomes eas-ier (see Fig. 4.15).

(c) Right-angle gear dri ve with f an support

(b) Belt Drive

(a) Direct Motor Drive

Fig. 4.12 Fan drives: (a) direct drive; (b) V-belt drive; (c)right-angle gear drive

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A-frame configuration

This design is almost exclusively employed in power plants for condensing turbine ex-haust steam (see Fig. 4.16). The tube bundles are mounted on a triangular frame with thefans located below. The incli-nation from the horizontal isusually between 45 and 60.

The A-frame configura-tions permit a 3040% reduc-tion in the plot area ascompared to a horizontal con-figuration. Additionally, and noless importantly, the A-frame isideally adapted for condensingsince it facilitates condensatedrainage. The common headerat the top of the unit allowsuniform steam distribution withminimum pressure loss, whichis important for the efficientoperation of vacuum steamcondensers. The A-frame con-figuration is in fact the basis ofseveral patented freeze-proofdesigns.

Plenum

Fan ring

Bundle

Fig. 4.13 Box type plenum (redrawn with permission fromHTRI)

Vertical configuration

Vertical configurations aregenerally employed for pack-aged units such as compres-sors with their intercoolers(see Fig. 4.17). Evidently,they are used where floorspace is at a premium andcould thus be used advanta-geously in offshore platforms

as well. They are much more prone to deterioration in per-formance due to crosswinds.Furthermore, multipass de-signs are not feasible for con-densing services.

Plenum

Fan ring

Bundle

Fig. 4.14 Transition plenum (redrawn with permission fromHTRI)

Fig. 4.15 Horizontal configuration of an air-cooled heat ex-changer (redrawn with permission from HTRI)

4.5 Natural Draft versusMechanical Draft

In natural draft there are nofans, so that the flow of air is

by natural convection due to

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the stack effect across the tube bundle.An external stack is sometimes incorpo-rated to increase the draft and therebythe cooling. The principal application isin dry cooling towers in power plantswhere a large chimney (dry cooling

tower) establishes an appreciable draft.Some process licensors specify externalstacks for process condensers so as toachieve 65 or 70% plant throughputeven under power failure conditions.

Most air-cooled heat exchangers areof the mechanical draft type. Vastamounts of air are moved across finned-tube bundles by axial fans driven byelectric motors. There are two principal

categories of mechanical draftforceddraft and induced draft. In forced draft(see Fig. 4.18), the fans are mounted

below the tube bundles and blow airacross the finned tubes. In induced draft(see Fig. 4.19), the fans are located abovethe tube bundles and suck air across thefinned tubes. Each type has its advantagesand disadvantagesand therefore pre-ferred application, as elaborated below.

Fig. 4.16 A-frame configuration of an air-cooledheat exchanger (redrawn with permission fromHTRI)

Fig. 4.17 Vertical configuration of an air-cooledheat exchanger (redrawn with permission fromHTRI

4.6 Forced Draft versusInduced Draft

Forced draft advantages

The principal advantages are as follows:1) Since both fans and motors/drivetransmissions are located below the tube

bundles, accessibility of the same for

maintenance is far better.2) Since the fans are located below thetube bundles and handle the colder in-coming air, the air pressure drop andtherefore the fan power consumption aresomewhat lower.3) Fan blade life is longer since expo-sure is to cold inlet air.4) It is possible to have a recirculation

air-cooled heat exchanger system toavoid freezing and other solidification problems only with forced draft fans.Fig. 4.18 Forced draft (redrawn with permission

from HTRI)

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Forced draft disadvantages

The principal disadvantages are as follows:1) Poorer distribution of air across the tube bundles, since the air leaves the tube

bundles at a much lower velocity.2) Greater possibility of hot air recirculation as a result of the lower discharge ve-

locity and the absence of a stack. Hot air recirculation results in an increase of the

air inlet temperature and consequently a decrease in the MTD. In low-MTD applications, the deterioration in performance can be significant. Consequently, theforced draft type is not preferred where the cold end temperature approach (thedifference in temperature between the process outlet and the inlet air) is less than914F (58C).

3) Exposure to the elements (sunlight, rain, hail, and snow), unless louvers or roofsare provided at the top of the tube bundles. This results in poorer stability and

process control, as well as possible damage to the finned tubes.4) Due to a very small stack effect, natural draft capability in the event of fan failure

is rather low.

Induced draft advantages

The principal advantages are as follows:1) Better air distribution across the tube bundles with better cooling.2) The probability of hot air recirculation is considerably lower. The air velocity at

the discharge is usually over twice that at the entrance.3) Due to the much higher stack effect, natural draft capability under fan failure

conditions is much higher.

4) Better process control and stability from effects of rain, snow, hail, or sunlight, aswell as protection from the damaging/negative effects from the same.

5) No possibility of damage to fans and/or drive due to leaking products, where cor-rosive.

Induced draft disadvantages

The main disadvantages are as follows:1) The fans and drives are less accessible for maintenance, being located above the

plenum chamber. Further, maintenance work may even have to be carried out in

the hot air caused by natural convection.

Fig. 4.19 Induced draft (redrawn with permission from HTRI)

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2) Higher air pressure drop and thereby motor power because of handling hotter air,which is lighter.

3) In order to prevent damage to fan blades, V-belts, bearings, and other mechanicalcomponents, the exit air temperature has to be limited to about 194F (90C). Fi-

ber-reinforced plastic blades (which give superior performance) cannot be used atan air temperature higher than 158F (70C).

References

[1] Taborek, J., 1987, Bond Resistance and Design Temperatures for High-Finned Tubes - AReappraisal, Heat Transfer Eng. , 8(7), pp. 2634.

[2] McHugh, S., and Chappell, S.E. 1999, Specify the right fin type for air-cooled heat exchang-ers, Hydrocarbon Process. , Sept., pp. 6772.

[3] API, 1992, Air-cooled Heat Exchangers for General Refinery Services , API Standard 661, 3rdEd., April, American Petroleum Institute, Washington, DC.

Further Reading1. Monroe, R.C., 1979, Improving Cooling Tower Fan System Efficiencies, Combustion , May,

pp. 2026.2. Gardner, K.A., 1945, Fin Efficiency of Several Types of Straight Fins, Trans. ASME , 67 , pp.

621631.

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C HAPTER 5

Thermal Design of Single-Phase Air-cooled Heat Exchangers

5.1 Introduction

The design of air-cooled heat exchangers comprises two distinct activities: thermal de-

sign and mechanical design. In thermal design, the heat exchanger is sized, which meansthat all the principal construction parameters such as number of bays, number of bun-dles/bay, number of tube rows, number of tubes per row, tube OD and thickness, fin ODand density, tube length, tube pitch, number of tube passes, and nozzle sizes are deter-mined. In mechanical design, detailed calculations are carried out to determine the di-mensions of various components such as tubesheets or plugsheets, header boxes, flanges,etc. and a complete bill of materials and set of engineering drawings are generated. In this

book, we shall talk predominantly about thermal design.The basic equations for tubeside and airside heat transfer and pressure drop are well

known and are presented in several books (see references). This chapter will dwell on theapplication of these and other correlations for the optimum thermal design of air-cooled heatexchangers. Before we proceed any further, let us see what the broad objectives of a thermaldesigner are when he or she sets out to produce a thermal design.

An air-cooled heat exchanger, like any other heat exchanger, must satisfy the following basic equation:

(MTD)Q

AU

= (5.1)

where

A = heat transfer areaU = overall heat transfer coefficientMTD = mean temperature difference

The overall heat transfer coefficient is determined as follows:

air

1(tubeside) (air) f f w

t

I lr r

U h hr = + + + + (5.2)

where

hair = airside heat transfer coefficient

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ht = tubeside heat transfer coefficientr f = fouling resistancer w = tube wall resistance

Since the airside heat transfer coefficient is generally much lower than the tubeside heattransfer coefficient, it becomes necessary to use extended the surface (finned tubes) so asto make the airside heat transfer coefficient compatible with the tubeside heat transfercoefficient. The fin height and fin density (fins per in. or m) can be varied so as to incor-

porate the optimum extent of the extended surface. The more the airside heat transfer co-efficient is controlling, the greater will be the extended surface required. Thus, for air-cooled steam condensers or hot water coolers, which have a very high tubeside heattransfer coefficient, the airside will be highly controlling and the designer should there-fore use the highest fin density (11 fins per in. or 433 fins per m) and 5/8 in. (15.875 mm)fin height. For viscous liquid coolers, however, where the tubeside heat transfer coeffi-cient is much lower so that the airside heat transfer coefficient is not controlling, it will

be prudent to use a lower fin density such as 57 fins per in. (197276 fins per m) and a

fin height of 0.5 in. (12.7 mm). In an extreme situation with a very low airside heat trans-fer coefficient, even bare tubes may be considered.

5.2 Broad Objectives of Thermal Design

The basic aims of a thermal designer are as follows:a ) Produce a thermal design that has a low overall cost; the lower, the better. The

overall cost of a heat exchanger is the sum of the initial cost and the operatingcost. The initial cost is evidently the fixed cost or the first cost of the heat ex-changer. The operating cost is the sum of the pumping cost, the maintenance

cost, and the downtime cost. The maintenance cost is the sum of the cost of peri-odically cleaning the exchanger, the cost of any antifoulant treatment, and thecost of any repair or replacement.

Thus, it is not enough to produce a design having a very low fixed cost if itsoperating cost is high due to, say, frequent fouling and thereby the requirementfor cleaning. Designers often lose sight of the operating cost of a heat exchangerand should always attempt to minimize fouling and also minimize pressure drop.

Now, this represents a direct conflict because as we shall see later on in the book,the best way to minimize fouling is to maximize velocity (within limits of ero-sion, of course) that will directly maximize pressure drop and thereby power con-

sumption. Obviously, then, the designer has to optimize the design so that whilethe velocity is not low enough to exacerbate fouling, the pressure drop is not ex-cessively high.

In this context, the selection of the materials of construction is very impor-tant. The materials should be good enough to permit the heat exchanger to func-tion for the lifetime of the plant (typically 2025 years) without major repairs andwithout replacement of components (such as tubes). However, the materialsshould not be inordinately expensive because then the first cost of the heat ex-changer will become unnecessarily high.

b) Utilize allowable pressure drops as fully as possible.It will be easily appreciated that the higher the velocity of a given stream, thehigher will be its heat transfer coefficient. However, accompanying the high heat

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transfer coefficient will be a high pressure drop. So, while the former (high heattransfer coefficient) will tend to reduce the first cost of the heat exchanger, thelatter (high pressure drop) will tend to increase the operating cost of the heat ex-changer. Thus, a very important goal for a good thermal design is the best utiliza-tion of the allowable pressure drop. This is discussed in more detail later in thischapter.

It sometimes so happens that the permitted pressure drop is unnecessarilyhigh to produce a good design and if, in such cases, the pump specifications havenot been frozen, they can be revised to take advantage of the lower (than antici-

pated) pressure drop. However, if the pump specifications have already been fro-zen, the possible saving in pumping power cannot be realized and the differential

pressure drop will just have be let down (typically) across a control valve.c) Maintain adequate tubeside velocity to minimize fouling.

This has just been discussed above, and is treated in much more detail in Chapter 10.

5.3 Data to be Furnished for Thermal Design

Before coming to the actual thermal design of an air-cooled heat exchanger, let us take alook at the data required for the same.

The following information must be furnished by the process licensor for the hot processstream (wherever applicable) before thermal design can be taken up:

Flow rate

The complete requirements of vapor, liquid, steam, water, and noncondensable flow ratesmust be furnished, as applicable, at both the inlet and outlet of the heat exchanger.

Inlet and outlet temperaturesEvidently, the inlet and outlet temperatures have to be specified, as they will go towardthe determination of the heat duty, and also toward the calculation of the mean tempera-ture difference.

Heat duty

It is a good idea for the thermal engineer to corroborate the heat duty since licensors oc-casionally slip in this aspect. For sensible cooling services, the heat duty is simply the

product of the mass flow rate, the average specific heat, and the difference between the

inlet and the outlet temperatures. For condensing services, the total heat duty is the sumof the sensible vapor cooling duty, the sensible liquid cooling duty, and the condensingduty, which is the product of the amount of vapor condensed per hour and the latent heatof condensation.

Heat release profiles

By heat release profiles are meant plots of the following variables with temperature,wherever applicable: heat duty, weight fraction vapor, and vapor molecular weight. Evi-dently, for single-phase services, the last two are not applicable. Besides, the plot of heatduty versus temperature is essentially linear so that no heat release profile is really re-quired.

However, for any service involving phase change, heat release profiles as defined aboveare a must. If the temperature difference between the inlet and the outlet is rather small, such

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as 9 oF (5 oC) or 18 oF (10 oC), a straight-line heat duty versus temperature may be specifiedsince the curvature will be minimal. Heat release and other profiles will be discussed indetail in Chapter 8.

Operating pressure

This is not really required for liquids since their properties do not vary with pressure to

any significant extent. However, it is required for gases and vapors since their properties, particularly gas density, vary with pressure. However, if the physical properties are fur-nished, the operating pressure is no longer required for single-phase gas and condensingvapor streams.

Allowable tubeside pressure drop

This is a very important parameter for air-cooled heat exchanger design and the processlicensor should be aware of the significance of the same for thermal design. The higherthe pressure drop, the higher will be the heat transfer coefficient and thereby the lowerthe heat transfer area and fixed cost. However, the operating cost will be higher. Conse-quently, the allowable pressure drop represents the optimum balance between fixed costand operating cost of a heat exchanger such that the total cost is minimal. Generally, forliquids, a value of 710 psi (0.50.7 kg/cm 2) is permitted per shell. A higher value is usu-ally warranted for viscous liquids, especially if routed through the tubeside. For gases, theusually allowed value is 0.72.8 psi (0.050.2 kg/cm 2), a very typical value being 1.4 psi(0.1 kg/cm 2).

It must be stated here that whereas typical values are generally applicable, specificinstances must be investigated more thoroughly. For example, if it is found that theallowable pressure drop for a particular stream represents a severe constraint in producing asatisfactory thermal design, the effect of a higher allowable pressure must be examined toarrive at the optimum design based on minimum total cost.

It may be stated here that this aspect is very important for good thermal design of heatexchangers and indeed for any good design: the designer must not follow the beaten path butalways question the various parameters specified and examine alternatives. It should beremembered that some of the parameters specified are not really sacrosanct but are only

based on hereditary engineering practice. A special situation may call for special measures.The author has found that it always helps to keep asking oneself: Is there not a better wayof doing this? Why don't I see what happens if I change this parameter?

Fouling resistanceThis is another extremely important parameter and one that is unfortunately based moreon experience than fundamental understanding, thanks to the complexity of the phe-nomenon. If the fouling resistance of a particular stream is not furnished, the heat ex-changer designer should adopt the same from TEMA standards or from past operatingexperience. This subject is discussed in far greater detail in Chapter 10.

Physical properties

Principally viscosity, thermal conductivity, density, and specific heat, preferably at bothinlet and outlet temperatures. Viscosity data must be supplied at inlet and outlet tempera-tures, especially for liquids, since the variation with temperature is considerable and ir-regular (neither linear nor semilog nor log-log). Additional properties required are latent

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heat and surface tension for condensing services. Physical properties are discussed atlength in Chapter 8.

Line sizes

It is desirable to match nozzle sizes with line sizes since no expander or reducer will then be required. However, criteria for nozzle sizing (velocity and v2) are usually more strin-

gent for heat exchanger nozzles than for lines. Nozzle sizing is based on pressure drop,which in turn is based on expansion and contraction losses, whereas line sizing is basedon line pressure drop, which is dependent on velocity and the length of the line. Conse-quently, nozzle sizes are sometimes required to be one size (or even more in exceptionalcircumstances) larger than the corresponding line sizes. This is especially true for smallline sizes where the change in flow area from one pipe size to the next is quite consider-able.

Tube size

By tube size is meant tube OD, thickness, and length. As per the API 661 Standards, 1 in.(25.4 mm) is the smallest OD of tubes to be used in air-cooled heat exchangers. This issomewhat surprising, considering that shell-and-tube heat exchangers in the same chemi-cal process industries can be designed and built with 0.75 in (19.05 mm) tubes. This isone of the mysteries that the author has never been able to unravel!

Since air-cooled heat exchangers are rather large and occupy large plot areas, they areinvariably located over pipe racks. In such situations, the tube length is usually 1.64 ft (0.5m) greater than the pipe-rack width for reasons of mechanical convenience. Thus, the tubelength of an air-cooled heat exchanger gets fixed by the pipe-rack width of the unit in whichit is going to be located.

Occasionally, an air-cooled heat exchanger is located on a technological platform andin such situations its tube length can be optimized so as to yield the most cost-effectivedesign.

The selection of the tube length of an air-cooled heat exchanger is discussed in moredetail in Section 7.1.

Materials of construction (MOC)

The materials of construction of the tubes, tubesheets, and headers should be specified bythe process licensor. Since a process or other stream entering an air-cooled heat ex-changer is usually not at an elevated temperature, the most common tube material is car-

bon steel. In a shell-and-tube heat exchanger, the materials of construction of the compo-nents that face both the shellside and tubeside fluids (for example, tubes and floating-head covers) have to be so selected so that they can withstand both the shellside and tube-side fluids. Often, a material that is suitable for the tubeside fluid is not suitable for theshellside fluid. In such cases, material selection can become difficult and finally end witha very expensive material. In an air-cooled heat exchanger, however, the situation ismuch simpler since air is not corrosive and material selection is solely on the basis of thetubeside fluid. The fin material universally used for air-cooled heat exchangers is alumi-num since it exhibits the most favorable thermal conductivity-to-cost ratio. Copper has athermal conductivity far greater than that of aluminum but its cost is disproportionatelyhigher.

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Corrosion allowance

Corrosion allowance for the various pressure parts have to be specified by the processlicensor. No discrete corrosion allowance is applied on tubes since the standard tubethicknesses recommended by TEMA already incorporate a corrosion allowance.

Special considerations

All pertinent requirements, such as cycling, upset conditions, alternate cases o

practical thermal design of air-cooled heat exchangers

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