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10 Selection of Heat Exchangers andTheir ComponentsAs described in Chapter 1, a variety of heat exchangers are available, and thequestion becomes which one to choose for a given application. In addition, foreach type (core construction), either a large number of geometrical variables (suchas those associated with each component of the shell-and-tube exchanger) or a largenumber of surface geometries (such as those for plate, extended surface, or regen-erative exchangers) are available for selection. Again the question involves which setof geometries/surfaces will be most appropriate for a given application. There is nosuch thing as a particular heat exchanger or the selection of a particular heattransfer surface that is best (i.e., optimum) for a given application. Near-optimumheat exchanger designs involve many trade-offs, since many geometrical and operat-ing variables are associated with heat exchangers during the selection process. Forexample, a cheaper exchanger may be obtained if one wants to give up some per-formance or durability. One can get higher performance if the exchanger is heavieror costs a little more. A heat exchanger can be made smaller if we accept a littlelower performance or provide more pumping power for higher fluid flow rates. Theheat exchanger design team must consider the trade-offs and arrive at the optimumexchanger for a given application to meet the design requirements and constraints.In this chapter we discuss qualitative and quantitative criteria/methods used to selectexchanger type and surface geometry for a given application for engineers not havingprior design/operational experience. If one o r more exchangers are already in service forsimilar applications, this prior experience is the best guide fo r the selection an d design ofa heat exchanger for a given application. We first describe impo rtan t qualitative selectioncriteria fo r heat exchangers in two categories: 1 ) criteria based on impor tant operatingvariables of exchangers in Section 10.1, an d (2) general guidelines on major heat exchan-ger types in Section 10.2. Next, we describe some quantitative criteria fo r selection ofextended heat exchanger surfaces (screening methods) in Section 10.3.1 an d for selectionof tubula r exchangers (performance evaluation criteria) in Section 10.3.2. Thesequantitative criteria are energy-based (the first law of thermodynamics) for a heatexchanger as a component. Criteria based on the second law of thermodynamics canalso be devised as indicated in Section 10.3.3, with the details presented in Section 11.7.Finally, selection based o n cost criteria is presented briefly in Section 10.3.4 an d discussedfurther in Section 11.6.6. Except for some qualitative discussion, it is no t easy to present amethod for system-based selection and optimization of a heat exchanger in a textbooksince there are many systems in which heat exchangers are used, and each is different,

673

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674 SELECTION O F HEAT EXCHANGERS A ND THE IR COMPONENTS

depending o n the process. Thus , the overall objective of this chap ter is to provide a g oodunderstanding of heat exchanger selection in general a s a compo nent based o n qualita-tive an d qu antitative criteria. However, system-based heat exch anger optimization is thecurrent industrial practice.

10.1 SELECTION CRITERIA BASED ON OPERATING PARAMETERSA large number of heat exchangers are described in Chapter 1, providing under-standing of their functions, the range of operating parameters, reasons why they areused in certain applications, and so on. With such thorough understanding, onewould have a good idea of which types of exchangers to use in given applications.Refer to Table 10.1 for operating conditions and principal features for many heatexchanger types. Here we highlight heat exchanger selection criteria based on majoroperating parameters.

10.1.1 Operating Pressures and TemperaturesThe exchanger in operation must withstand the stresses produced by the operatingpressure and the temperature differences between two fluids. These stresses dependon the inlet pressures and temperatures of the two fluids. The most versatile exchan-gers for a broad range of operating pressures and temperatures are shell-and-tubeexchangers for medium- to high-heat duties and double-pipe exchangers for lower-heat duties. They can handle from high vacuum to ultrahigh fluid pressures [gener-ally limited to 30MPa (4350psi) on the shell or annulus side and 140MPa (20,000psi) on the tube side]. Coupled with high pressures, shell-and-tube exchangers canwithstand high temperatures, limited only by the materials used; however, the inlettem per atur e difference is limited t o 50 C (1 20F) fr om the the rmal expa nsion pointof view when the exchanger design allows only limited thermal expansion, such as inthe E-shell design. These exchangers are used for gas, liquid, and phase-changeapplications.F o r liquid-liquid o r liquid-phase chan ge applications, if the opera ting pressures an dtemperatures ar e modest (less tha n a bo ut 2.5 M Pa an d 200 C), gasketed o r semiweldedplate exchangers should be considered. For somewhat higher pressures and tempera-tures, fully welded or brazed plate exchangers may be the choice, depending on otherdesign criteria.The plate-fin extended surface exchanger is designed for low-pressure applications,with the op eratin g pressures o n either side limited to a bo ut 1000 kP a (150 psig), exceptfor cryogenics applications, where the operating pressure is about 9000 kPa gauge(1 300 psig). T he max imum op eratin g temp erature for plate-fin excha ngers is below650 C (1200F) an d usually below 150C (300 F), to avo id the use of expensive materials.There is no limit on the minimum opera ting temperature; plate-fin exchangers are com-mon ly used in cryogenic ap plications. Fin s in a plate-fin ex chan ger act as a flow-mixingdevice for highly viscous liquids, and if properly designed, add surface area for heattransfer with a reason ably high fin efficiency. In a plate-fin e xcha nger, fins o n the liquidside are used primarily for pressure containme nt an d rigidity. Fins o n the gas side areused for added surface area for heat transfer, with fin efficiencies usually greater than8 0 % .

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SELECTION CRITERIA BASED ON OPERATING PARAMETERS 675

The tube-fin exchanger is used to contain the high-pressure fluid on the tube side ifonly one fluid is at a high pressure. Fins o n the liquid or phase-change side generally havelow heights, to provide reasonably high fin efficiencies. Tu rbu lato rs may be usedwithin tubes for flow mixing. Tube-fin exchangers with or without shells are designedto cover the operating temperature range from low cryogenic temperatures to about870C (1600F).For ultrahigh temperature [870 to 2000C (1600 to 3600 F)I and near-atmosphericpressures, as in high-temperature waste heat recovery, either rota ry regenerators (870 to1100C) or fixed-matrix regenerators (up to 2000C) are used.

10.1.2 c o s tCost is a very important factor in the selection of the heat exchanger constructiontype. The cost per unit of heat transfer surface area is higher for a gasketed plateexchanger than for a shell-and-tube exchanger. However, from the total cost (capital,installation, operation, maintenance, etc.) point of view, PHEs are less expensivethan shell-and-tube exchangers when stainless steel, titanium, and other higher qual-ity alloys are used. Since tubes are more expensive than extended surfaces or aregenerator matrix, shell-and-tube (or broadly, tubular) exchangers are in generalmore expensive per unit of heat transfer surface area. In addition, the heat transfersurface area density of a tubular core is generally much lower than that of anextended surface or regenerative exchanger. Rotary regenerators made of paper orplastic are in general the least expensive per unit of heat transfer surface area.10.1.3 Fouling and CleanabilityFouling and cleanability are among the most important design considerations forliquid-to-liquid or phase-change exchangers and for some gas-to-fluid exchangers.Fouling should be evaluated for both design and off-design points. Periodic cleaningand /or replacement of some exchanger com pon ents depend on the fouling propensityof the fluids employed. In applications involving moderate to severe fouling, either ashell-and-tube or a gasketed plate heat exchanger is used, depending on the otheroperating parameters. In a shell-and-tube exchanger, the tube fluid is generallyselected as the heavily fouling fluid since the tube side may be cleaned more easily.A plate heat exchanger is highly desirable in those relatively low temperature appli-cations [ < 300C (575 F)I where severe fouling occurs on one or both sides, as platedisassembly, cleaning, and reassembly is a relatively easy task. For highly corrosivefluid heating or cooling applications, shell-and-tube exchangers are used exclusively,regardless of opera ting pressure and temperature conditions. Plate-fin exchangersusually have small hydraulic diameter passages and hence are more susceptible tofouling. They are also relatively difficult to clean and are not employed in evenmoderate fouling applications unless they can be cleaned chemically or thermallyby baking (see Section 13.4).Th e fouling an d cleanability problem is no t a s severe for gas-to-gas exchangers as forliquid-to-liquid or phase-change exchangers, since in most applications gases are neithervery dirty nor have the fouling propensity of water. Regenerators have self-cleaningcharacteristics because the hot and cold gases flow periodically in opposite directionsthrough the same passage. Hence, they can tolerate m odera te fouling. If the applicationhas a po tential fo r heavy fouling, a large r flow passage size is chosen, as in a fixed-matrix

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9 TABLE 10.1 Principal Features of Several Types of Heat ExchangersFeatureType of

Heat Compactness Stream Temperature Maximum Exchanger (m2/m') Types Materialh Range ( C) Pressure (bar)' Plate-and-frame

(gaskets)

Partially weldedplate

Fully weldedplate(AlfaRex)

Brazed plateBavex plate

Platular plate

CompablocPackinox plate

plate

Spiral

up to 200

up to 200

up to 200

up to 200200-300

200

up to 300up to 300

up to 200

Liquid-liquid,gas-liquid,two-phase

Liquid-liquid.gas-liquid,two-phase

Liquid-liquid,gas-liquid,two-phase

Liquid-liquid,two-phaseGases, liquids,two-phaseGases, liquids,two-phaseLiquidsGases, liquids

two-phaseLiquid-liquid.two-phase

s/s Ti, Incoloy,Hastelloy,graphite,polymerIncoloy,HastelloyNi alloys

sjs Ti,

sjs Ti,

-35 to +200

-35 to +200

-50 to f350

sjs Ni, Cu,Ti, specialsteelssjs.Hostelloy,

Ni alloysIncoloyHastelloy,lnconelTi, Incoloy,Hastelloy

sjs Tisjs Ti,

CIS, SIS

-195 to +220-200 to +900

up to 700

up to 300-200 to +700

up to 400

25

25

40

3060

40

32300

25

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Brazed 8W1500 Gases, liquids, Al, s / s , Cryogenic 90Diffusion- 70 800 Gases, liquids, Ti, s / s up to 500 > 200

plate-fin two-phase Ni alloy to +650bonded two-phaseplate-fin

Printed-circuit 20 5000 Gases, liquids s / s , Ni, Ni -200 to +YO0 > 400Polymer (e.g. 450 Gas-liquidp PVDFq up to 150 6Plate-and-shell Liquids s/s, Ti, (shell up to 350 70Marhond up to 10,OOO Gases, liquids, S / s , Ni, Ni -200 to +YO0 > 400

two-phase alloys, Tichannel plate) P P

also in c/s)' two-phase alloys, Ti

Source : Data from Lancaster 1998).' s / s , stainless steel; Ti, titanium; Ni, nickel; Cu, copper. Alloys of these materials and other special alloys are'The maximum pressure capability is unlikely to occur at the higher operating temperatures, and assumes no

Two-phase includes boiling and condensing duties.

Can he dismantled.Function of gasket as well as plate material.Not common.On gasket side.On welded side.' Ensure compatibility with copper braze.' unction of braze as well as plate material.'N ot in a single unit.

'On tube side."'Only when flanged access provided; otherwise, chemical cleaning.' ive fluids maximum." On shell side.Condensing on gas side.

Polyvinylidene difluoride.' olypropylene.' PEEK (polyetheretherketone) can go to 250C'Shell may be composed of polymeric material.9 On plate side.

.l

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678 SELECTION O F HEAT EXCHANGERS AN D THEIR COMPONENTS

regenerator, so tha t the impact of fouling is reduced, or cleaning by o ne of the meth odsdescribed in Section 13.4 ma y be em ployed .10.1.4 Fluid Leakage and ContaminationWhereas in some applications, fluid leakage from one fluid side to the other fluidside is permissible within limits, in other applications fluid leakage is absolutely notallowed. Even in a good leak tight design, carryover and bypass leakages from thehot fluid to the cold fluid (or vice versa) occur in regenerators. Where these leakagesand subsequent fluid contamination is not permissible, regenerators are not used.The choices left are either a tubular, extended surface, or some plate type heatexchangers. Gssketed plate exchangers have more probability of flow leakage thando shell-and-tube exchangers. Plate-fin and tube-fin exchangers have potential leak-age problems at the joint between the corrugated fin passage and the header or atthe tube-to-header joint. Where absolutely no fluid contamination is allowed (as inthe processing of potable water), a double-wall tubular or shell-and-tube exchangeror a double-plate PHE is used.

10.1.5 Fluids and Material CompatibilityMaterials selection and compatibility between construction materials and workingfluids are important issues, in particular with regard to corrosion (see Section13.5) and/or operation at elevated temperatures. While a shell-and-tube heat exchan-ger may be designed using a variety of materials, compact heat exchangers oftenrequire preferred metals or ceramics. For example, a requirement for low cost, lightweight, high conductivity, and good joining characteristics for compact heat exchan-gers often leads to the selection of aluminum for the heat transfer surface. On theother side, plate exchangers require materials that are either used for food fluids orrequire corrosion resistance (e.g., stainless steel). In general, one of the selectioncriteria for exchanger material depends on the corrosiveness of the working fluid.In Table 10.2, a summary of some materials used for noncorrosive and/or corrosiveservices is presented. More details about the selection of materials are provided inspecialized literature of TEMA (1999) and the ASME (1998) codes.

10.1.6 Fluid TypeA gas-to-gas heat exchanger requires a significantly greater amount of surface areathan that for a liquid-to-liquid heat exchanger for a given heat transfer rate. This isbecause the heat transfer coefficient for the gas is to 6 that of a l iquid. Theincrease in surface area is achieved by employing surfaces that have a high heattransfer surface area density O F'or example, fins are employed in an extendedsurface heat exchanger, or a small hydraulic diameter surface is employed in aregenerator, or small-diameter tubss are used in a tubular heat exchanger. Plateheat exchangers (of the type described in Section 1.5.2) are generally not used ina gas-to-gas exchanger application because they produce excessively high pressuredrops. All prime surface heat exchangers with plain (uncorrugated) plates are used insome waste heat recovery applications. The fluid pumping power is generally signifi-cant and a controlling factor in designing gas-to-gas exchangers.

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SELECTION CRITERIA BASED ON OPERATING PARAMETERS 679

TABLE 10.2 Materials for Noncorrosive and Corrosive ServiceMaterial Heat Exchanger Type o r Typical Service

Noncorrosive ServiceAluminum and austenitic chromium-nickel Any heat exchanger type, < -100C3 $ Ni steel Any heat exchanger type, -100 < < -45CCarb on steel (impact tested) Any heat exchanger type, -45 < T < 0CCarbon steel Any type of heat exchanger, 0 < < 500CRefractory-lined steel Shell-and-tube, > 500C

steel

Corrosive ServiceCarbon steelFerritic carbon-molybdenum andchromium-molybdenum alloysFerritic chromium steel

Austenitic chromium-nickel steelAluminumCopper alloys: admiralty, aluminum brass,High nickekhromium-molybdenumcupronickel

Mildly corrosive fluids; tempered cooling waterSulfur-bearing oils at elevated temperatures (above300C); hydrogen at elevated temperaturesTubes for moderately corrosive service; cladd ing forshells or channels in contact with corrosivesulfur bearing oilCorrosion-resistant dutiesMildly corrosive fluidsFreshwater cooling in surface condensers; brackishand seawater coolingResistance to mineral acids and CI-containing acidsalloysTitaniumGlassCarb on Severely corrosive serviceCoatings: aluminum, epoxy resinLinings: lead and rubberLinings: austenitic chromium-nickel steel

Source: Data from Lancaster (1998).

Seawater coolers and condensers, including PH EsAir preheaters for large furnacesExposure to sea and brackish waterChann els for seawater coolersGeneral corrosion resistance

In liquid-to-liquid exchanger applications, regenerators are ruled ou t because of theassociated fluid leakage and carryover (contamination). Fluid pumping power is, how-ever, not as critical for a liquid-to-liquid heat exchanger as it is for a gas-to-gas heatexchanger.to j

of that o n the liquid side. Therefore, for a thermally balanced designt (i.e., having h Aof the same o rd er of magn itude o n each fluid side of the exchanger), fins are employed toincrease the gas-side surface area. Th us , the co m m on heat ex changer construc tions usedfor a liquid-to-gas heat exchanger are the extended surface and tubular; plate-type andregenerative constru ctions are n ot used.Fo r phase-change exchangers, the condensing o r evap orating fluid has a range of heattransfer coefficients th at vary f rom low values appro xima ting those f or ga s flows to highvalues approximating those for high liquid flows and higher. Therefore, the selection of

In a liqu id-to-gas heat e xchan ger, the heat tran sfer coefficient o n the gas side is

A thermally balanced design usually results in an op timu m design from the cost v iewpoint since the cost of theextended surface per unit surface area is less than that of the prime surface, either tubes or plates.

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680 SELECTION OF HEAT EXCHANGERS AN D THEIR COMPONENTSexchanger type fo r phase-change exchangers parallels the guidelines p rovided for the gasor liquid side of the exchanger.

10.2 GENERAL SELECTION GUIDELINES FOR MAJOR EXCHANGERTYPESA large number of heat exchangers are described in Chapter 1 . That information,complemented by the material presented in this section, will provide a good under-standing of the selection of heat exchanger types.10.2.1 Shell-and-Tube ExchangersMore than 65% of the market share (in the late 1990s) in process and petrochemicalindustry heat exchangers is held by the shell-and-tube heat exchanger, fo r the follow-ing reasons: its versatility for handling a wide range of operating conditions with avariety of materials, design experience of about 100 years, proven design methods,and design practice with codes and standards. The selection of an appropriate shell-and-tube heat exchanger is achieved by a judicious choice of exchanger configura-tion, geometrical parameters, materials, and the right design. Next we summarizesome guidelines on all these considerations qualitatively to provide the feel for theright design for a given application. The major components of a shell-and-tubeexchanger are tubes, baffles, shell, front-end head, read-end head, and tubesheets.Depending on the applications, a specific combination of geometrical variables ortypes associated with each component is selected. Some guidelines are providedbelow. For further details on geometrical dimensions and additional guidelines,refer to TEMA (1999).10.2.1.1 Tubes. Since the desired heat transfer in the exchanger takes place acrossthe tube surface, the selection of tube geometrical variables is important from aperformance point of view. In most applications, plain tubes are used. However,when additiona l surface area is required t o co mp ensate for low heat transfer coefficientson the shell side, low finned tubing with 250 to 1200 fins/m (6 to 30 fins/in.) and a finheight of up t o 6.35 mm a in.) is used. W hile m aintaining reason ably high fin efficiency,low-height fins increase surfac e ar ea by tw o to three times over plain tubes a nd decreasefouling on the fin side based o n the da ta reported.The m ost com mo n plain t ub e sizes have 15.88, 19.05, an d 25.40 m m ($,i,nd 1 in.)tube outside diameters. Fro m the he at transfer viewpoint, smaller-diameter tubes yieldhigher heat transfer coefficients and result in a more compact exchanger. However,larger-diameter tubes are easier to clean and more rugged. The foregoing commonsizes represent a compromise. For mechanical cleaning, the smallest practical size is19.05 mm (in.). For chemical cleaning, smaller sizes can be used provided that thetubes never plu g completely.The number of tubes in an exchanger depends on the fluid flow rates and availablepressure d ro p. The nu mb er of tub es is selected such th at th e tube-side velocity fo r waterand similar liquids ranges from 0.9 to 2.4 m/s (3 to 8 ft/sec) and the shell-side velocityfrom 0.6 to 1.5 m /s (2 to ft/sec). Th e lower velocity limit cor resp on ds to limiting thefouling, an d the uppe r velocity limit corresp onds t o limiting the rate of erosion. Wh ensand and silt are present, the velocity is kept high enough to prevent settling.

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G E N E R A L SELECTION GUIDELINES FOR MAJOR EXCHANGER TYPES 68The num ber of tube passes depends on the available pressure dro p. H igher velocitiesin the tube result in higher heat transfer coefficients, at the expense of increased pressuredro p. There fore, if a higher pressure d ro p is acceptable, it is desirable to have fewer butlonger tubes (reduced flow area and increased flow length). Long tubes are accommo-da ted in a sh or t shell exchanger by multiple tube passes. Th e number of tube passes in ashell generally range from 1 to 10 (see Fig. 1.61). The stand ard design has one, two, orfour tube passes. An odd number of passes is uncommon and may result in mechanicaland thermal problems in fabrication and operation.

20.2.2.2 Tube Pitch and Layou t. The selection of tube pitch is a comprom ise betweena close pitch (small values of p,/d, ,) for increased shell-side heat transfer and surfacecompactness, and an open pitch (large values of p , / d , ) for decreased shell-side pluggingand ease in shell-side cleaning. In most shell-and-tube exchangers, the ra tio of the tubepitch to tube outside diameter varies from 1.25 to 2 .00. The m inimum value is restrictedto 1.25 because the tubesheet ligamentt m ay become to o weak for proper rolling of thetubes and cause leaky joints. The recommended ligament width depends on the tubediameter and pitch; the values are provided by TEMA (1999).Two stand ard types of tube layouts ar e the square and the equilateral triangle, shownin Fig. 10.1. The equ ilateral pitch can be oriented a t 30 or 60 angle to the flow direction,and the square pitch at 45 and 90 .$ No te that the 30 , 45 nd 60 arrangements arestaggered, an d 90 is inline. Fo r the identical tube pitch a nd flow rates, the tube layouts indecreasing ord er of shell-side heat transfer coefficient and pressure dr op are: 30 , 45 , 0 ,and 90 . Th us the 90 layout will have the lowest heat transfer coefficient an d the lowestpressure drop.The square pitch (90' or 45 ) is used when jet or mechanical cleaning is necessary onthe shell side. In tha t case, a m inimum cleaning lane of in. (6.35 mm) is provided. Thesquare pitch is generally not used in the fixed tubesheet design because cleaning is notfeasible. The triangular pitch provides a more compact arrangemen t, usually resulting ina smaller shell, an d the strongest header sheet for a specified shell-side flow are a. Hence, itis preferred when

30

the operating pressure difference between the two fluids is large. If

Triangular Rotatedtriangularor paralleltriangularSquare

FIGURE 10.1 Tube layout arrangements.

Rotatedsquare

'The ligament is a portio n of material between two n eighboring tub e holes. The ligament width is defined as thetube pitch minus the tube hole diameter, such as the distance (I shown in Fig. 10.1.$Note hat the tube layout angle is defined in relation to the flow direction and is not related to the horizontal orvertical reference line. Refer to Tab le 8.1 for the definitions of tub e layou ts and associated geometrical variables.

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682 SELECTION O F HEAT EXCHANGERS AND THEIR COMPONENTSdesigned properly, it can be cleaned from six directions instead of four as in the squarepitch arran gem ent. W hen m echanical clea ning is required , the 45 layo ut is preferred forlaminar or turbulent flow of a single-phase fluid and for condensing fluid on the shellside. If the pressure dr o p is limited on the shell side, the 90 layo ut is used fo r turbulentflow. F o r boiling applications , the 90 layo ut is preferred , because it provid es vap orescape lanes. Ho wev er, if mech anical cleaning is no t req uired, the 30 layout is preferredfor single-phase laminar or turbulent flow and condensing applications involving a highA T ranget (a mixture of condensables). The 60" layout is preferred for condensingapplications involving a low A T range (generally, pure vapor condensation) and forboiling applications. Horizontal tube bundles are used for shell-side condensation orvaporization.10.2.1.3 Bafles. As presented in Section 1.5.1.1, baffles may be classified as eitherlongitudinal or transverse type. Longitudinal baffles are used to control the overallflow direction of the shell fluid. Transverse baffles may be classified as plate bafflesor grid baffles. Plate baffles are used to s up po rt the tubes, to direct the fluid in the tub ebundle at approximately right angles to the tubes, and to increase the turbulence andhence the heat transfer coefficient of the shell fluid. However, the window sectioncreated by the plate baffles results in excessive pressure drop with insignificant contri-bution to heat transfer; flow normal to the tubes in crossflow section may create flow-induced vibration problems. The rod baffles, a most common type of grid baffles,shown in Fig. 1 . 11 , are used to support the tubes and to increase the turbulence.Flow in a rod baffle heat exchanger is parallel to the tubes, and hence flow-inducedvibration is virtually eliminate d by the baffle su pp or t of the tub es. T he choice of baffletype, spacing, and cut are determined largely by the flow rate, required heat transfer,allowable pressure drop, tube support, and flow-induced vibration. The specificarrangements of baffles in various TEMA shells are shown in Fig. 10.3.

Plate Baflees. Tw o types of plate baffles, shown in Fig. 1.10 are segmental, and disk an ddoughnut. Single and double segmental baffles are used most frequently. The singlesegmental baffle is generally referred to simply as a segmental b@e. Th e practical rangeof single segmental baffle spac ing is f to 1 shell diameter, althoug h op timum could beto i. Th e minimum baffle spacing for cleaning the bu ndle is 50.8 m m ( 2 in.) o r shelldiam eter, whichever is larger. Spacin gs closer th an shell diam eter provid e add edleakage$ that nullifies the heat transfer advantage of closer spacings. If the foregoinglimits on the baffle spacing do not satisfy other design constraints, such as Ap,,, ortube vibration, no-tubes-in-window or pure crossflow design should be tried.The segmental baffle is a circular disk (with baffle holes) with one disk segmentremoved. The baffle cut varies from 20 to 49% (the height l in Fig. 8.9 given as apercentage of the shell inside diameter), with the most common being 20 to 2 5 % . Atlarger spacings, it is 45 to 50%, to avoid excessive pressure dro p across the windows ascompared to the bundle. Large or small spacings coupled with large baffle cuts areundesirable because of the increased potential of fouling associated with stagnant flow

'Here the A T range represents the difference in condensing tem peratu re at the inlet minus condensing te mp era-ture at the outlet of an exchanger.:These are tube to baffle hole, baffle to shell, bundle to shell, and the tu be pass pa rtition leakages or bypassesdescribed in Section 4.4.1.

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GENERAL SELECTION GUIDELINES FOR MAJOR EXCHANGER TYPES 683areas. If fouling is a primary concern, th e baffle cut should be kep t below 25 . The bafflecut and spacing should be designed such that the flow velocity has approximately thesame magnitude for the cross flow and window flow sections. Alte rnate segmental bafflesare arranged 180 to each o the r, which cause shell-side flow t o a ppr oac h crossflow in thecentra l bundle regiont a nd axial flow in the window zone. All segmental baffles shown inFig. 1.10 have horizontal baffle cuts. The direction of the baffle cut is selected as followsfo r shell-side fluids: Eithe r horizon tal or vertical fo r a single-phase fluid (liquid or gas),horizontal for be tter m ixing for very viscous liquids, and vertical for the following shell-side applications: condensation (for better drainage), evaporation/boiling (for no strati-fication and f or providing disengagement room ), entrained particulates in liquid (to pro-vide least interference fo r solids to fall ou t), and multishell pass exchanger, such a s those inFig. 1.62 an d the F shell.Since one of the principal functions of the plate baffle is to s up po rt the tubes, the termsb a s e and support plate are sometimes used interchangeably. However, a sup por t platedoes not direct the fluid normal to the tube bank, it may be thicker than a baffle, it hasless tube-to-baffle hole clearance, an d it provides greate r stiffness to the bundle. Su pp or tplates with single-segmental baffles are cut approximately at the centerline and spaced0.76 m (30 in.) apa rt. This results in an unsupp orted tube span of 1.52 m (60 in.) becauseeach plate supports half the number of tubes. The double-segmental baffle (Fig. l.lO),also referred to a s a strip bufle , provides lower shell-side pressure d ro p (and allows largerfluid flows) tha n t ha t f or the single segmental baffle for the same unsupported tube span.The baffle spacing for this case should not be too small; otherwise, it results in a moreparallel (longitudinal) flow (resulting in a lower heat transfer coefficient) with significantzones of flow stagnation . Triple-segmental baffles have flows with a strong parallel flowcomponent, provide lower pressure dro p, and permit closer tube sup po rt to prevent tubevibrations.Th e lower allowable pressure dro p results in a large baffle spacing. Since th e tubesin the window zone are suppo rted a t a distance of two or m ore times the baffle spacing,they are most susceptible to vibration. To eliminate the possibility of tube vibrationsand to reduce the shell-side pressure dr op , the tubes in the window zone are removed a ndsupport plates are used to reduce the unsupported span of the remaining tubes. Theresulting design is referred to as the segmental bafle with no-tubes-in-window, shown inFig. 1.10. Th e su pp ort plates in this case are circular an d s up por t all the tubes. T he bafflecut and number of tubes removed varies from 15 to 25%. Notice that low-velocityregions in the baffle corners d o no t exist, resulting in good flow characteristics and lessfouling. Thus the loss of heat transfer surface in the window section is partiallycompensated for. However, the shell size must be increased to compensate for the lossin the surface area in the window zone, which in turn may increase the cost of theexchanger. If the shell-side operating pressure is high, this no-tubes-in-window designis very expensive comp ared to a similar exchanger having tubes in the window zone.The disk-and-doughnut baffle is made up of alternate disks and doughnut-shapedbaffles, as shown in Fig. 1.10. Generally, the disk diame ter is somewhat g reater th an thehalf-shell diameter, and the diameter of the hole of the doughnut is somewhat smallerthan the half-shell diam eter. This baffle design provides a lower pressure d ro p comparedto th at in a single-segmental baffle for the same unsup ported tube span an d eliminates thetube bundle-to-shell bypass stream C.The d isadvantages of this design are th at (1) all theVarious allowable clearances required for construction of a tube bundle with plate baffles are provided by TEMA(1999).

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684 SELECTION OF HEAT EXCHANGERS AND THEIR COMPONENTS

tie rods to hold baffles are within the tube bundle, an d (2) the central tubes ar e suppo rtedby the disk baffles, which in turn ar e sup por ted o nly by tube s in the ov erlap of the larger-diameter disk over the dou gh nu t hole.Rod Baffles. Rod baffles are used to eliminate flow-induced vibration problems. Forcertain shell-and-tube exchanger applications, it is desirable to eliminate the cross flowand have pure axial (longitudinal) flow on the shell side. Fo r the ca se of high shell-sideflow rates a nd low-viscosity fluids, the rod baffle exchange r has several adva ntag es overthe segmental baffle excha nger: (1) It elim inates flow-induced tub e vib ratio ns since thetubes a re rigidly sup por ted a t four poi nts successively; (2) the pressu re d ro p on the shellside is ab ou t one-half th at with a doub le segmental baffle a t the sa me flow rate a nd heattransfer ra te. Th e shell-side heat transfer coefficient is also conside rably low er tha n th atfor the segmental baffle exchanger. In general, the rod baffle exchanger will result in asmaller-shell-diameter longer-tube unit having m ore surface ar ea for the same heattransfer and shell-side pressure drop; (3) there are no stagnant flow areas with therod baffles, resulting in reduced fouling and corrosion and improved heat transferover that for a plate baffle exchanger; (4) since the exchanger with a rod (grid) baffledesign has a counterflow arrangement of the two fluids, it can be designed for higherexchanger effectiveness and lower mean (or inlet) temperature differences than those ofan exchanger with a segmental baffle design; and (5) a rod baffle exchanger will gen-erally be a lower-cost unit and has a higher exchanger heat transfer rate to pressuredrop ratio overall than that of a segmental baffle exchanger. If the tube-side fluid iscontrolling and has a pressure drop limitation, a rod baffle exchanger may not beapplicable. Refer to Gentry (1990) for further details on this exchanger.Impingement Baffles. Impingement baffles or plates are generally used in the shell sidejust below the inlet nozzle. Their pu rpo se is to protect the tubes in the to p row nea r theinlet nozzle from erosion, cavitation, and/or vibration due to the impact of the high-velocity fluid jet from the nozzle to the tubes. One of the most common forms of thisbaffle is a solid square plate located under the inlet nozzle just in front of the first tuberow, as shown in Fig. 10.2. The location of this baffle is critical within the shell tominimize the associated pressure drop and high escape velocity of the shell fluid afterthe baffle. Fo r this purpose, adequ ate area s should be provided both between the nozzleand plate and between the plate and tube bundle. This can be achieved either byomitting some tubes from the circular bundle as shown in Fig. 10.2 or by modifyingthe nozzle so that it has an expanded section (not shown in Fig. 10.2). Also, properpositioning of this plate in the first baffle space is important for efficient heat transfer.

Good baffle design Bafle too close Baffle too bigImpingement baffles at the shell-side inlet nozzle. (From Bell, 1998.)I G U R E 10.2

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Shell Type

TEMA E

TEMA F

TEMA G

TEMA H

TEMA Jsingle nozzleentry

TEMA Jdouble nozzleentry

Llongitudinalflow

TEMA Xcross flow

Fixed Tubesheet andFloating Head Bundles U-Tube Bundles

FIGURE 10.3 Shell-side flow arrangement for various shell types (Courtesy of Heat TransferResearch, Inc., C ollege Station, T exas).

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686 SELECTION OF HEAT EXCHANGERS AND THEIR COMPONENTSEnough space should be provided between the tip of the plate and the tubesheet andbetween the tip of the plate and the first segmental baffle. The most common cause oftube failure is improper location and size of the impingement plate.10.2.1.4 Shells. Seven type s of shells, as classified by T E M A (1999), are shown in Fig.I .6; they ar e also shown in F ig. 10.3 with baffles. The E shell, the m ost c om m on d ue toits low co st an d relative simplicity, is used f or single-phase shell fluid a pp licatio ns an dfor small cond ensers with low v ap or volumes. Mu ltiple passes o n the tu be side increasethe heat transfer coefficient h (if corresponding more increased Ap is within allowedlimits). However, a multipass tube arrangement can reduce the exchanger effectivenesso r F factor compared to that for a single-pass arrangement (due to some tube passesbeing in parallelflow) if the increased h and NT U d o not compensate for the parallel -Row effect. Two E shells in series (in overall coun terflow configuration ) may be used toincrease the exchanger effectiveness E .

As an alternative, a co unterflow arra nge me nt is desirable (i.e., high E ) for a two-tube-pass exchanger. This is achieved by the use of an F shell having a loiigitudinal baffle,resulting in two shell passes. How ever, a T E M A F shell is rarely used in practice becauseof heat leakage across the longitudinal baffle an d potential flow leakage t ha t ca n o ccur ifthe area between the lon gitudinal baffle an d the shell is no t sealed properly. Also, the Fshell presents additional problems of fabrication and maintenance, and it is difficult toremove o r replace th e tu be bund le. If o ne needs to increase th e exchan ger effectiveness,multiple shells in series are preferred o ver a n F shell.T he T E M A G an d H shells are related to the F shell bu t h ave different long itudinalbaffles. Hence, when the shell-side Ap is a limiting factor, a G or H shell can be used;however, E or Fwill be lower th an t ha t of a counterflow exchanger. Th e split-flow G shellhas horizo ntal baffles with the ends remove d; the shell nozzles are 180 ap ar t at themidpoint of the tubes. The double-split-flow H shell is similar to the G shell, but withtwo inlet and two outlet nozzles and two longitudinal baffles. The G and H shells areseldom used for shell-side single-phase app lications, since there is no adv anta ge over Eo r X shells. They ar e used as horizonta l thermo siphon reboilers, condensers, and otherphase-change applications. The longitudinal baffle serves to prevent flashing of thelighter com pon ents of the shell fluid, helps flush o u t noncon densa bles, pro vides increasedmixing, an d helps distribu te the flow. Generally, A T and Ap across longitud inal bafflesare small in these application s, an d heat transfer acro ss the baffle and flow leakages a t thesides have insignificant influence o n the perfo rman ce. T he H shell app roa che s the cross-flow arrangem ent of the X shell, an d it usually has low shell-side Ap compared to the E,F , a n d G shells. F o r high-inlet-velocity application s, two nozzles are req uired a t the inlet,hence the H or shell is used.The divided-flow TEMA J shell has two inlets and one outlet or one inlet and twooutlet nozzles (a single nozzle a t the midpoin t of th e tubes an d tw o nozzles nea r the tub eends). The J shell has approximately one-eighth the pressure drop of a comparable Eshell an d is therefore used fo r low-pressure-drop applications such as in a con denser invacu um. Fo r a condensin g shell fluid, the J shell is used, with two inlets for the gas pha sean d on e central outlet [or the conden sate an d residue gases.T he T E M A K shell is used fo r partial ly vapo rizing the shell fluid. I t is used a s a kettlereboiler in the process industry and as a flooded chiller (hot liquid in tubes) in therefrigeration industry. Usually, it consists of an o verall circular-cross-section ho rizo ntalbundle of U tubes placed in an oversized shell with on e o r more v ap or nozzles on the topside of the shell (see on e vap or nozzle in Fig. 1.6) to reduce l iquid entrainment. Th e tube

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l H o t wellF I G U R E 10.4 An X shell exchanger with omission of top tube rows for better flow distributionfor a condensing application. (From Bell, 1998.)

bundle diameier ranges 50 to 70% of the shell diameter. The liquid (to be vaporized)enters from below near the tubesheet through the left-hand nozzle an d covers the tubebundle. Pool an d some convective boiling takes place on the shell side without forcedflow of the vaporizing fluid outside the tubes on the shell side. The vapor occupies theupper space in the shell without the tubes. The large empty space in the shell acts as avapor disengaging space; and if properly sized, almost dry vapor exits from the topnozzle, thus eliminating the need for an external vapor-liquid separator. Hence, it iscommonly used, although it is more expensive to fabricate, particularly, for high-pressure applications. Generally, the kettle reboiler is considered as a pool boiling device;however, convective (flow) boiling prevails in the tube bund le.For a given flow rate and surface area, the crossflow TEMA X shell has the lowestshell-side pressure d ro p of all (except for K) shell configurations. Hence, it is used for gasheating a nd cooling applications with o r without finned tubes and f or vacuum conden-sing applications. It is also used fo r applications having large shell flows.No transversebaffles are used in the X shell; however, supp ort plates a re used to suppress the flow-induced vib rations. Flow d istributions on the shell side could be a serious problem unlessproper provision has been made to feed the fluid uniformly at the inlet. This could beachieved by a bathtub nozzle, multiple nozzles, or by providing a clear lane along thelength of shell near the nozzle inlet as shown in Fig. 10.4.The type of shell described in Fig. 1.6 has either o ne o r two shell passes in on e shell.The cost of the shell is much more than the cost of tubes; hence, a designer tries toaccom moda te the required heat transfer surface in one shell. Three o r fou r shell passesin a shell could be made by the use of longitudina l baffles.+Multipassing on the shell sidewith long itudina l baffles will reduce the flow area per pass com pared to a single pass o nthe shell side in a single shell, resulting in a possibly higher shell-side pressure drop.Multiple shells in series are also used for a given application for the following reasons:

They increase the exchanger effectivenessE , or reduce the surface area for the sameE . Fo r the latter case, a subseq uent reduction in tubing cost may offset the cost of anadditional shell and other components.

'Positive or tight sealing between the longitu dinal baffles and the shell is essential to ma intain the high exchan gerefectivenesses p redicted.

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688 SELECTION OF HEAT EXCHANGERS A ND THEIR COM P ONENTS

0 F or an ex chang er requiring high effectiveness, multipass ing is the only alternative .Fo r part-load opera tion a nd where a spare bundle is essential, mu ltiple shells (may

0 Shipping an d handling m ay dictate restrictions on the overall size or weight of th ebe smaller in size) will result in a n econo mical o per ation .un it, resulting in multiple shells for a n app licatio n.

In heat recovery trains and some other applications, up to six shells in series arecommonly used. The limitation on the number of shells in such applications is thepressure drop limit on one of the fluid streams.10.2.1.5 Front-End Heads. Th e front- and rear-end head types, a s classified by TE M A(1999), ar e show n in Fig. 1.6. Th e front-end head is stationary, while the rear-end headcan be either stationary o r floating, depending on the allowed thermal stresses betweenthe tubes an d the shell. Th e ma jor criteria for the selection of front- and rear-end headsare the thermal stresses, o perating pressures, cleanability, hazards, an d cost.The front-end heads are primarily of two types, the channels and the bonnet. Thebonnet head B is cast in one piece and has either a side- or an end-entering nozzle.+Although the bo nnet head is less expensive, inspection and m aintenan ce requires break-ing the pipe joints a nd removing the bo nnet. Hence, the bon net head is generally used fo rclean tube-side fluids. The channel head c an be removable, as in the TE M A A head, o rcan be integral with the tubesheet, as in TEMA C and N heads. There is a removablechannel cover in these front-end heads for inspection an d maintenance withou t disturb-ing the piping. T he nozzles a re side entering in these types. N otice t ha t while th e shell iswelded o nto the TE M A N head, it is flanged to the TE M A C head. In the TE M A N head,no mechanical joint exists (all welded joints) between the channel and tubesheet andbetween the tubesheet an d the shell, thus eliminating leakage between th e shell an d thetubes. The TE M A D head ha s a special high-pressure closure an d is used for applicationsinvolving 2100 kPa gauge (300 psig) fo r amm onia service an d higher pressures for oth erapplications.10.2.1.6 Rear-End Heads. In a shell-and-tube exchanger, the shell is at a temp eraturedifferent from that of the tubes because of heat transfer between the shell and tubefluids. This results in a differential thermal expansion and stresses among the shell,tubes, and the tubesheet. If proper provisions are not made, the shell or tubes canbuckle, or tubes can be pulled apart or out of the tubesheet. Provision is made fordifferential thermal expansion in the rear-end heads. They may be categorized as fixedor floating rear-end heads, depending on whether there are no provisions or someprovisions for differential thermal expansion. A more commonly used third designthat allows tube expansion freely is the exchanger with U tubes having the front-and rear-end heads fixed; it is included in the floating rear-end-head category in thediscussion to follow. The design features of shell-and-tube exchangers with variousrear-end heads are sum marized in Table 10.3.A hea t exchan ger with a fixed rear-end head L, M , or N has a fixed tubesheet on thatside. Hence, the overall design is rigid. Th e tub e bundle-to-shell clearance is least am on gthe designs, thus minimizing the bundle-to-shell bypa ss stream C. Any number of tube+Notice hat the nozzles on the fron t- and rear-end heads are fo r the tube fluid. Th e nozzles for the shell fluid arelocated on the shell itself.

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GENERAL SELECTION GUIDELINES FOR MAJOR EXCHANGER TYPES 689

TABLE 10.3 Design Features of Shell-and-Tube He at ExchangersOu tside- Outsid e- InsideReturn packed Packe d Pull- Split

Design Fea ture Tubesheet (U-T ube) box Ring Bundle RingFixed Bend Stuffing Latern Thr oug h Backing

TEM A R ear-Head Type: L, M, N U P w T STu be bundle removable N o Yes Yes Yes Yes YesSpare bundles used N o Yes Yes Yes Yes YesProvides for differential Yes, with Tes Yes Yes Yes Yes

movem ent between bellows inshell and tube s shellreplacedchemically cleaned,both inside andoutsidemechanicall cleaned specialon inside toolsmechanically cleanedon outsidebolting ar e requiredpractical

Individual tubes can be Yes Yesa Yes Yes Yes YesTube s can be Yes Yes Yes Yes Yes Yes

Tub es can be Yes With Yes Yes Yes Yes

Tubes can be Yes Yesb Yesb Yesb Yesb Yesb

Internal gaskets and N o N o No No Yes YesDouble tubesheets are Yes Yes Yes N o N o N oNumber of tubesheet Any Any even Any' On e or Anye Any'passes available num ber twodApproximate diametral 1 1-18 11-18 25-50 15-35 95-160 35-50clearance (mm)(Shell ID, Do,,

ascending orde r,(leastexpensive= 1)

Relative costs in 2 1 4 3 5 6

Source: Data from Shah (1995).a Only those in outside rows can be replaced without special designs.'Ou tside mechanical cleaning possible with square or ro tated s qua re pitch, or wide triang ular pitch.'Axial nozzle required a t rear end for odd nu mber of passes.Tube-side nozzles must be at stationary end for two passes.'Odd number of passes requires packed or bellows at floating head.

passes can be employed. Th e TE M A L, M , a n d N rear-end heads ar e the counte rparts ofT E M A A , B, and N front-end heads. The major disadvantages of the fixed tubesheetexchang er are (1) no relief for thermal stresses between the tubes and the shell, (2) theimpossibility of cleaning the shell side mechanically (only chemical cleaning is possible),and (3) the impracticality of replacing th e tube bu ndle . Fixed tubesheet e xch ang ers arethus used for applications involving relatively low temperatures [315C (600F) andlower] coup led with low pressures [2100 kP a ga uge (300 psig) and lower]. As a rule

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690 SELECTION OF HEAT EXCHANGERS AND TH EIR COMPONENTSof thu mb , the fixed tubeshee t design is used for a n inlet tem per atur e difference betweenthe two fluids th at is less tha n a bo ut 50 to 60C (100F).If an expa nsion bellows is used,this temperature difference can be increased to about 80 to 90C (1 50F).Expansionbellows ar e generally un econo mical for high pressures [ > 41 50 kPa gauge (600 psig)]. Thefixed tubesheet exchanger is a low-cost unit ranked after the U-tube exchanger.Th e differential thermal expan sion can be accom mod ated by a floating rear-end headin which tub es expan d freely within the shell, thus elimin ating thermal stresses. Also, thetube bundle is removable for mechanical cleaning of the shell side. Basically, there arethree types of floating rear-end heads: U-tube heads, internal floating heads (pull-through/split-ring heads), an d o utside packed floating heads.In the U-tube bundle, the thermal stresses are significantly reduced, due to freeexpansion of the U-tubes, and the rear-end head has an integral cover which is theleast expensive among rear-end heads. The exchanger construction is simple, havingonly one tubesheet and no expansion joints, and hence it is the lowest-cost design,particularly a t high pressures. Th e tube bundle c an be removed for shell-side cleaning;however, it is difficult to remov e a U tub e from th e bun dle except in the ou te r row, an d i tis also difficult to clean the tu be-side bends mec hanically . So a U-tube exchanger is usedwith clean fluids inside the tubes unless the tube side can be cleaned chemically. Flow-induced vibration can also be a prob lem fo r the tubes in the outerm ost row because of along unsupp orted span, particularly in large-diameter bundles.Th e next-simplest floating head is the pull-throu gh hea d T shown in Fig. 10.5. O n thefloating-he ad side, the tub esheet is small, acts a s a flange, and fits in th e shell with its ow nbonn et head . Th e tube bundle can easily be removed from the shell by first removing the

FI G U RE 10.5 Two-pass exchanger (BET) ith a pull-through (T) rear-end h ead. (Courtesy ofPatternson-Kelley Co ., Division of HA RS CO Co rpora tion, East Strou dsburg, Pennsylvania.)

Baffcuts --a)

Shell-Tubefieldealingstrips 3 7Dummytubes

Sealing strip or dumm y tube

F I G U R E 10.6 a)Sealing strips; (b)dummy tubes or ties rods; (c) sealing strips, dummy tubes, ortie rods covering the en tire length of the tube bundle.

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GENERAL SELECTION GUIDELINES FOR MAJOR EXCHANGER TYPES 691

front-end head. Individual tubes or the tube bundle can also be replaced if required.Due to the floating-head bonnet flange an d bolt circle, many tubes are omitted from thetube bundle near the shell. This results in the largest bundle-to-shell circumferentialclearance o r a significant bundle-to-shell bypass stream C. So as not to reduce exchangerperformance, sealing strips (or dummy tubes o r tie rods with spacers) in the bypass a reaare essential, as shown in Fig. 10.6. They are placed in pairs every five to seven tubepitches between the baffle cuts. They force the fluid from the bypass stream back in to thebundle. However, localized high velocities near the sealing strips could cause Row-induced tube vibration; hence, proper care must be exercised for the design. Since thisdesign has the least number of tubes in a bundle for a given shell diameter compared toother floating-head designs, the shell diameter is somewhat larger, to accommodate arequired amount of surface area. One of the ideal applications of the TEMA T headdesign is in the kettle reboiler, for which there is ample space on the shell side an d the flowbypass stream C is of no concern.The large bundle-to-shell clearance can be minimized by bolting the floating-headbonnet to a split backing ring (flange) as shown in Fig. 10.7. It is referred to a s the TE M AS rear-end head. Th e shell cover over the tube floating head has a diameter larger thanthat of the shell. As a result, the bundle-to-shell clearances are reasonable and sealingstrips are generally not required. However, both ends of the exchanger must be disas-sembled for cleaning and maintenance. In both TE M A S and T heads, the shell fluid isheld tightly to prevent leakage to the outside. However, internal leakage is possible due tothe failure of an internal hidden gasket an d is not easily detectable. Th e TEM A T headhas m ore positive gasketing between the two streams than does the S ead. Both TEM AS and T head configurations are used for the tube-side multipass exchangers; the single-pass construction is not feasible if the advantages of the positive sealing of TE M A S andT heads are to be retained. The cost of TEMA S and T head designs is relatively high

Pass partitionrib or plate Floating head1FIGURE 10.7 Two-pass exchanger (AES) with a split-ring (S) floating head. (Courtesy of

Patternson-KelleyCO ., Division of HAR SCO Corporation, East Stroudsburg, Pennsylvania.)Split ringflange

Stuffing box,

SkirtFIGURE 10.8 Two-pass exchanger (AEP) with an outside packed (P ) floating head. (Courtesy ofPatternson-Kelley Co., Division of HAR SCO Corporation, East Stroudsburg, Pennsylvania.)

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692 SELECTION O F HEAT EXCHANGERS AND THEIR COMPONENTS

c I

F I G U R E 10.9 Two-pass exchanger (BEW) having a packed floating head (W ) with lantern rings.(Courtesy of Patternson-Kelley Co., Division of HARSCO Corporation, East Stroudsburg,Pennsylvania.)

compared to U-tu be o r fixed-tubesheet un its. Th e cost for the T E M A S head is higherthan for the TEM A T hea d. T he split backing ring floa ting head is used extensively in thepetroleum industry for moderate operating pressures and temperatures. For very highoperating pressures an d temperatures, the T E M A head design has a special test ring(TEMA, 1999).In the outside-packed floating-head TEM A P design of Fig. 10.8, the stuffing boxprovid es a seal against the skirt of the floating head an d preve nts shell-side fluid leak ageto the ou tside. This skirt (and th e tube b undle ) is free to m ove axially again st the seal totake thermal expansion i nto acc oun t. A split-ring flange near the en d of the skirt seals theback end of the chamber. Because of the specific design of this floating head, any leak(from either the shell side o r the tube side) at the gaskets is to the outside. Hence, th eT E M A P head is generally not used with very toxic fluids. Also, the inlet and outletnozzles mus t be located at th e statio nar y end; hence, this design could hav e only a n evennum ber of tu be passes. In this design, the bundle-to-shell clea rance is large [abou t 38 m m(1.5 in.)]; as a result, sealing strips are required. The TEMA P head exchanger is moreexpensive than the TE M A W head exchanger.Th e packed floating head with lantern ring or TE M A W head is shown in Fig. 10.9.Here a lantern ring rests o n the machined surface of the tubesheet and provides aneffective seal between the shell- and tube-side flanges. Vents are usually provided inthe lantern ring to help locate any leaks in the seals before the shell-side and tube-sidefluids mix. Although a single-pass design is possible on the tu be side, generally an evennum ber of tub e passes is used. Th e TE M A W head ex changer is the lowest-cost design ofall floating heads. Althou gh its cost is higher t ha n th at of the U-tub e bun dle, this highercost is offset by the accessibility to the tube ends (by opening both rear- and front-endheads) for cleaning and repair; consequently, this design is sometimes used in thepetrochemical and process industries.A large nu mb er of comb inatio ns of front- an d rear-end h eads with different shell typesof Fig. 1.6 are possible, depending on the application and the manufacturer. Somecom mon types of com binations result in the following shell-and-tube heat exchangers:AEL, AES , AEW, BEM, AEP , CFU , AK T, and AJW.In light of the availability of different types of front- and rear-end heads, the tubebundle of a shell-and-tube exchanger may simply be classified as a straight-tube o r U -tube bundle. Bo th have a fixed tubesheet at the fron t end. Th e U-tube bund le has a shellwith a welded shell cover on the U-bend end. T he straight tub e bundle has either a fixed

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Need to increase heat transfer

Increase surface areatransfer coefficientShell side Increase Increase shell EmployI tube leneth diameter with multiple shells

Increase number Decrease theof tubes baffle spacing ordecrease tube baffle cutoutside diameter Increase For EUse counterflow onfigurationsUse multiole shells confimuntion

Need to reduce pressure drop

appropriate in series ornumber of paralleltubes

Tube side Shell side

ncrease Increase Increase Use doubleDecrease Increase Decrease tube the baftle the baffle tube pitch or triplenumberof tube length andtube diameter increase shellpasses diameter andcut spacing segmentalbaffles

number oftubesFIGURE 10.10 Influence of various geometrical parameters of a shell-and-tube exchanger onheat transfer and pressure drop.

tubesheet or a floating h ead a t the rear end . Th e former is referred t o a s a fixed-tubesheetbundle; the latter is referred to a s a floating-head bundle.With this background, two flowcharts are presented in Fig. 10.10 to increase theoverall heat transfer rate o r decrease the pressure dr o p on e ither the tube or shell sidedurin g various stages of designing a sh ell-and-tube exchang er.10.2.2 Plate Heat ExchangersThe chev ron plate is the most co mm on in P HE s. H ence, we will not discuss the reason-ing behind why various oth er plate geometries have been used in PH Es . As described inSection 1.5.2.1, PHEs have a number of advantages over shell-and-tube heat exchan-gers, such a s com pactn ess, low to tal cost, less fouling, accessibility, flexibility in c ha n-ging the number of plates in an exchanger, high q and E , and low fluid residence time.Because of these advantages, they are in second place to shell-and-tube heat exchangersfor ma rket sh are in liquid-to-liquid an d phase-change applications. The main reaso n fortheir limited versatility involves the pressure and temperature restrictions imposed bythe gaskets. Replacing gaskets on one or both sides by laser welding of the plates (as in

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694 SELECTION O F HEAT EXCHANGERS AND TH EIR COMPONENTS

a welded PHE) increases both the operating pressure and temperature limits of thegasketed PHEs, and it allows the PHE to handle corrosive fluids compatible with theplate material. Fo r low-heat duties (th at translate into a total surface area up to 10 m2),a more compact brazed PHE can replace the welded PHE, thus eliminating the frame,guide bars, bolts, and so on , of welded o r gasketed PH Es. A variety of other P HE s havebeen developed fo r niche application s to cover specific operating co nditions th at ca nn otbe handled by the above-described PHE s. Som e of these PH Es ar e described briefly a tthe end of Section 1.5.2.2.Although the shell-and-tube heat exchanger is versatile an d c an han dle all kinds ofopera ting conditions, it is not c om pact, no t flexible in design, requires a large fo otprint,an d is costly (total cost) com pared to PH E s an d other co mp act heat exchangers. Hence,PH Es an d m any o ther heat ex changer designs have been invented to replace shell-and-tube heat exchan gers in individual narrow opera ting ranges. Refer t o S ections 1.5.2 an d1.5.3 for details of these exchangers. Also, a n excellent source of in forma tion on com pactheat exchangers for liquid-to-liquid and phase-change applications is a recent mono-graph by Reay (1999).

10.2.3 Extended Surface Exchangers10.2.3.1 Plate-Fin Exchanger Surfaces. The plate-fin construction is commonly usedin gas-to-gas or gas-to-phase chang e excha nger application s where either the hea ttransfer coefficients are low or an ultrahigh exchanger effectiveness is desired. It offershigh surface area densities [u p to a bo ut 6000 m 2/m 3 (1800 ft2/ft3)] an d a considerableamount of flexibility. The passage height on each fluid side could easily be varied anddifferent fins can be used between plates for different applications. On each fluid side,the fin thickness and number of fins can be varied independently. If a corrugated fin(such as the plain triangular, louver, perforated, or wavy fin) is used, the fin can besqueezed o r stretched to vary the fin pitch, thus providing added flexibility. The fins oneach fluid side could easily be arranged such that the overall flow arrangement of thetwo fluids can result in crossflow, coun terflow, o r parallelflow. Even the constru ction ofa multistream plate-fin exch ange r is relatively straig htfo rwa rd with th e pro per design ofinlet and outlet headers for each fluid (ALPEMA, 2000).Plate-fin exch angers are generally designed f or low -pressure application s, with op er-ating pressures limited t o ab ou t 1000 kP a g auge (150 psig). Ho wev er, cryogenic plate-finexchangers are designed fo r opera ting pressures of 8300 kPa (1200 psig). W ith m odernman ufacturin g technology, they can be designed fo r very high pressures; f or examp le, thegas cooler for an autom otive air-conditioning system with C 0 2 as the refrigerant has anopera ting pressure of 12.5 to 15 .0 M Pa (1800 to 2200 psig). T he m aximum operatin gtemperatures a re limited by the type of fin-to-plate bond ing a nd the materials employed.Plate-fin exchangers have been designed from low cryogenic operating temperatures[-200C (-400 F)] to ab ou t 800C (1500F). Fou ling is generally no t as severe a pro-blem with gases as it is with liquids. A plate-fin exchanger is generally not designed forapplications involving heavy fouling since there is no easy method of cleaning theexchanger unless chemical cleaning ca n be used. If an exchanger is made of small mod -ules (stacked in the height, width, a nd length directions), and if it can be cleaned with adetergent, a high-pressure air jet, or by baking it in an oven (as in a paper industryexchanger), it could be designed for applications having considerable fouling. Fluid

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contam ination (mixing) is generally no t a problem in plate-fin exchangers since there ispractically zero fluid leakage from one fluid side of the exchanger t o the other.Selection of a fin surface depends on the operating temperature, with references tobonding of the fins to plates or tubes and a choice of material. For low-temperatureapplications, a mechanical joint, or soldering or brazing may be adequate. Fins can bemade from copper, brass, or aluminum, a nd thus maintain high fin efficiency. Fo r high-temperature applications, only special brazing techniques and welding may be used;stainless steel and other expensive alloys may be used to make fins but with a possiblylower fin efficiency, due to their relatively lower thermal conductivities unless properlower fin height is selected. Consequently, suitable high-performance surfaces may beselected to offset the potential reduction in fin efficiency unless the proper fin height isselected. Brazing would require added capital an d maintenance cost of a brazing furnace,cost of brazing, and process expertise (SekuliC et al., 2003).Cost is a very important factor in the selection of exchanger construction type andsurface. Th e plate-fin surface in general is less expensive tha n a tube-fin surface per unitof heat transfer surface area . In most app lications, one does no t choose a high-perform-ing surface, bu t rathe r, the least expensive surface, if it can meet the performance criteriawithin specified constraints. For example, if a plain fin surface can do the job for aspecified application, the higher-performance louver or offset strip fin surface is notused because it is more expensive to manufacture.We now discuss qualitatively the construction and performance behavior of plain,wavy, offset strip, louver, perforated, and pin fin plate-fin surfaces.Plain Fin Surfaces. These surfaces are straight fins tha t ar e uninterrup ted (uncut) in thefluid flow direction. Although triangular an d rectangular passages are more comm on,any complex shape desired can be formed, depending on how the fin material is folded.Although the triangu lar (corruga ted) fin (e.g., Fig. 1.29a, e an d f ) is less expensive, canbe manufactured at a faster rate, and has the added flexibility of having an adjustablefin pitch, it is generally not structurally as st ron g as the rectangular fin (e.g., Fig. 1.29band d) for the same passage size and fin thickness. The triangular fins can be made invery low to ultrahigh fin densities [40 to 2400 fins/m 1 to 60 finlin.)].Plain fins are used in applications where the allowed pressure drop is low and theaugm ented interrupted surfaces can no t meet the design requirement of allowed p for adesired fixed fron tal area. Also, plain fins are preferred for very low Reynolds numbersapplications. This is because with interrupted fins, when the flow approaches the fullydeveloped state at such low Re, the advantage of the high h value of the interrup ted fins isdiminished while cost remains high, due to making interruptions. Plain fins are alsopreferred for high-Reynolds-number applications where the A p for interrupted finsbecome excessively high.W a v y Fin Surfaces. These surfaces also have uncut surfaces in the flow direction, andhave cross-sectional shapes similar to those of plain surfaces (see Fig. 1.29~). owever,they are wavy in the flow direction, whereas the plain fins are straight in the flowdirection. The waveform in the flow direction provides effective interruptions to theflow and induces very complex flows. The augmentation is due to Gortler vortices,which form as the fluid passes over the concave wave surfaces. These are counterrotat-ing vortices, which produce a corkscrewlike pattern. The heat transfer coefficient for awavy fin is higher than that for an equivalent plain fin. However, the heat transfercoefficient for wavy fins is lower than that for interrupted fins such as offset or louver

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696 SELECTION OF HEAT EXCHANGERS AND THEIR COMPONENTSfins. Since there are no cuts in the surface, wavy fins are used in those applicationswhere a n interrupted fin might be subject to a potential fouling or clogging problem du eto particulates, freezing moisture, bridging louvers due to condensate, and so on.Offset Strip Fins. This is one of the most widely used enhanced fin geometries in plate-fin hea t exchang ers (see Fig. 1.29d). Th e fin h as a rectangular cross section, an d is cutinto small strips of length Every altern ate strip is displaced (offset) by ab o ut 50% ofthe fin pitch in the transv erse direction. In add ition t o the fin spacing an d fin height, themajor variables are the fin thickness and strip length in the flow direction. The heattransfer coefficients for the offset strip fins are 1.5 to 4 times higher th an those of plainfin geometries. The c orrespo nding friction factors are also high. T he ratio of j / f for anoffset strip fin to j / f for a plain fin is about 80%. Designed properly, the offset strip finexchanger would require a substantially lower heat transfer surface area than that ofplain fins a t the same A p . The heat transfer enhancement for an offset strip fin is causedmainly by the redeveloping laminar boundary layers for Re 5 10,000. However, athigher Re, it acts as a rough surface (a constan t value o f f with decreasing j forincreasing Re).Offset strip fins are used in the ap pro xim ate Re rang e 500 to 10,000, where enhance-ment over the plain fins is substantially higher. For specified heat transfer and pressuredrop requirements, the offset strip fin requires a somewhat higher frontal area than aplain fin, but a sh orte r flow length an d overall lower volu me. Offset strip fins ar e usedextensively by aerospace, cryogenic, and many other industries where higher heat trans-fer performance is required.Louver Fins. Louvers are formed by cutting the metal and either turning, bending, orpushing o ut the cut elements from the p lane of the base metal (see Fig. 1.29e). Louverscan be made in many different forms and shapes. The louver fin gauge is generallythinner th an tha t of a n offset strip fin. Louver pitch (also referred to as louver width)and louver angle (in addition, the fin spacing and fin height) are the most importantgeometrical parameters for the surface heat transfer and flow friction characteristics.On an absolute level, j factors are higher for louver fins than for the offset strip fin atthe same Reynolds number, but th ef fa ct or s are even higher than those for the offsetstrip fin geometry. Since the louver fin is trian gula r (or co rrug ated ), it is generally n ot a sstron g as an offset strip fin; the latte r fin has a relatively large flat are a fo r brazing, th usproviding strength. The louver fins may have a slightly higher potential for fouling tha noffset strip fins. Louver fins are amen able to high-speed m ass prod uction man ufacturingtechnology, an d as a result, a re less expensive th an offset strip fins an d o ther inter rup tedfins when p roduc ed in very large quantities. T he fin spac ing desired ca n be achieved bysqueezing or stretching the fin; hence it allows some flexibility in fin spacing withoutchanges in tools and dies. This flexibility is not possible with the offset strip fin.A wide range in performan ce c an be achieved by varying the louve r angle, width, a ndform. The operating Reynolds number range is 100 to 5000, depending o n the type oflouver geometry employed. Modern multilouver fins have higher heat transfer coeffi-cients th at those for offset strip fins, but with som ewh at low er jlf ratios. However, theperformance of a well-designed multilouver fin exchanger can a pp ro ac h th at of a n offsetstrip exchanger, possibly with increased surface compactness an d reduced man ufacturingcost. M ultilouver fins (see Figs. 1.27, 1.28, an d 1.29e) ar e used extensively by the au to -motive industry.

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Perforated Fins. A perforated fin has either round or rectangular perforations with thesize, shape, and longitudinal and transverse spacings as major perforation variables (seeFig. 1.29f). Th e perforated fin has either triangular or rectangular flow passages. Whenused as a plate-fin surface, it is generally brazed. The holes interrupt the flow and mayincrease h somewhat, but considerable surface area may also be lost, thus nullifying theadvantage. Perforated fins are now used only in a limited number of applications. Theyare used as turbulators in oil coolers for mixing viscous oils, or as a high-Ap fin toimprove flow distribution. Perforated fins were once used in vaporizing cryogenic fluidsin air separation exchangers, but offset strip fins have now replaced them.Pin Fins. These can be manufactured at very high speed continuously from a wire ofproper diameter. After the wire is formed into rectangular passages (e.g., rectangularplain fins), the top and bottom horizontal wire portions are flattened for brazing orsoldering with the plates. Pins can be circular or elliptical in shape. Pin fin exchangerperformance is considerably lower, due to the parasitic losses associated with roundpins in particular an d to the inline arrangem ent o f the pins (which results from the high-speed manufacturing techniques). The surface compactness achieved by pin fin geo-metry is much lower than that of offset strip or louver fin surfaces. Due to vortexshedding behind the rou nd pins, noise- an d flow-induced vibration may be a problem.Finally, the cost of a round wire is generally more than the cost of a flat sheet, so theremay not be a material cost advantage. The potential application for pin fins is at verylow Reynolds number (Re < 500), for which the pressure d ro p is of no major concern.Pin fins are used in electronic cooling devices with generally free convective flows overthe pin fins.10.2.3.2 Tube-Fin Surfaces. When an extended surface is needed on only one fluidside (such as in a gas-to-liquid exchanger) or when the operating pressure needs to becontained on one fluid side, a tube-fin exchanger (see Section 8.2) may be selected, withthe tubes being round, flat, or elliptical in shape. Also, when minimum cost is essential,a tube-fin exchanger is selected over a plate-fin exchanger since the fins are not brazedbu t are joined m echanically to the tubes by mechanical expansion. F lat o r ellipticaltubes, instead of round tubes, are used for increased heat transfer in the tube andreduced pressure drop outside the tubes; however, the operating pressure is limitedcompared to that for round tubes. Tube-fin exchangers usually have lower heat transfersurface compactness than a plate-fin unit, with a maximum heat transfer surface areadensity of about 3300 m2/m3(1000 ft2/ft3).A tube-fin exchanger may be designed for a wide range of tube fluid operating pres-sures [up to abo ut 3000 kPa gauge (450 psig) o r higher] with the oth er fluid being at lowpressure [up to about 100 kPa (15 psig)]. The highest operating temperature is againlimited by the type of bonding and the materials employed. Tube-fin exchangers aredesigned to cover the operating temperature range from low cryogenic temperatures toabout 870C (1600F). Reasonable fouling can be tolerated on the tube side if the tubescan be cleaned. Fouling is generally not a problem on the gas side (fin side) in manyapplications; plain uninterrupted fins are used when moderate fouling is expected.Fluid contamination (mixing) of the two fluids is generally not a problem since thereis essentially no fluid leakage between them. Since tubes are generally more expensivetha n extended surfaces, the tube-fin exchanger is in general more expensive. In addition,the heat transfer surface area density of a tube-fin core is generally lower than tha t of aplate-fin exchanger, as mentioned earlier.

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Th e tube-fin c ons truc tion is generally used in liquid-to-gas or ph ase-cha nge fluid-to-gas heat exchang er applications with liquid, condensing fluid, o r evapora ting fluid o n thetube side. Fins are generally used on the outside of tubes (on the gas side), althoughdepending on the application, fins or turbulators may also be used inside the tubes.Ro und and flat tubes (rectangular tubes with rounded or s harp corners) are most com-mo n; however, elliptical tubes are also used. R ou nd tubes ar e used f or higher-pressureapplications and also when considerable fouling is anticipated. Parasitic form drag isassociated with flow normal to round tubes. In contrast, the flat tubes yield a lowerpressure drop for flow normal to the tubes, due to lower form drag, and thus avoidthe low-performance wake region behind the tubes. Also, the heat transfer coefficient ishigher for flow inside flat tubes t ha n for circular tubes, particularly at low R e. Th e use offlat tubes is limited to low-pressure applications, such as au tomo tive ra diators, unless thetubes ar e extrude d with ribs inside (see the m ultip ort tu be in Fig. 1.27, also referred to asmicrochannels) or with integral fins outside.Flat Fins on a Tube Array . Thi s type of tube-fin geo metry (show n in Fig. 1.31b) is mo stcommonly used in air-conditioning and refrigeration exchangers in which high pressureneeds to be contained on the refrigerant side. As men tioned earlier, this type of tube-fingeometry is no t as com pac t (in terms of surface ar ea density) as the plate-fin geometries,but its use is becoming widespread due to its lower cost. This is because the bondbetween the fin and tube is made by mechanically or hydraulically expanding thetube against the fin instead of soldering, brazing, or welding the fin to the tube.Because of the mechanical bond, the applications are restricted to those cases inwhich the differential thermal expansion between the tube and fin material is small,and preferably, the tube expansion is greater than the fin expansion. Otherwise, theloosened bond may have a significant thermal resistance.M any different types of flat fins are availab le (see some exam ples in Fig. 1.33). Th emost co mm on a re the plain, wavy, an d interrup ted. Th e plain flat fins are used in thoseapplications in which the pressure dr op is critical (quite low), although a larger am ou ntof surface area is required on the tube outside for the heat transfer specified than withwavy o r interrupted fins. Plain flat fins have the lowest pressure dr op tha n th at of anyoth er tube-fin surfaces a t the sam e fin density. Wav y fins are supe rior in performan ce toplain fins and are more rugged. Wavy fins are used m ost com mo nly for air-conditioningcondensers and other commercial heat exchangers. A variety of louver geometries arepossible on interrupted flat fins. A well-designed inter rup ted fin would h ave even betterperformance t ha n a wavy fin; however, it may be less rugged, m ore expensive to m an-ufacture, and may have a propensity to clog.Individually Finned Tubes. This tube-fin geometry (shown in Fig. 1.31~)s generallymuch more rugged than continuous fin geometry but has lower compactness (surfacearea density). Plain circular fins are the simplest and most common. They are manu-factured by tension wrapping the fin material around a tube, forming a continuoushelical fin or by mounting circular disks on the tube. To enhance the heat transfercoefficient on the fins, a variety of enhancement techniques have been used (see Fig.1.32). Segmented or spine fins are the counterpart of the strip fins used in plate-finexchangers. A segmented fin is generally rugged, has h eavy-g auge metal, a nd is usuallyless com pact tha n a spine fin. A studded fin is similar to a segmented fin, but individualstuds are welded to the tubes. A slotted fin has slots in the radial direction; whenradially slitted material is wound on a tube, the slits open, forming slots whose width

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SOME QUANTITATIVE CONSIDERATIONS 699

increases in the radial direction. This fin geometry offers an enhancement over tension-wound plain fins; however, segmented or spine fins would yield a better performance.The wire loop fin is formed by spirally wrapping a flattened helix of wire around thetube. The wire loops are held to the tube by a tensioned wire within the helix or bysoldering. Th e enhancem ent characteristic of small-diameter wires is im por tan t at lowflows, where the enhancement of other interrupted fins diminishes.10.2 .4 Regenerator SurfacesRegenerators, used exclusively in gas-to-gas heat exchanger applications, can have ahigher surface area density (a more compact surface) than that of plate-fin or tube-fin surfaces. While rotary regenerators have been designed for a surface area density/3 of up to ab out 8800 m2 /m 3 (2700 ft2/ft3) , he fixed-matrix regenerators have beendesigned for /3 of up to ab out 16,000 m 2/m 3 (5000 ft2/ft3). Regenerators are usuallydesigned for low-pressure applications, with operating pressures limited to near-atmospheric pressures for rotary and fixed-matrix regenerators; an exception is thegas turbine rotary regenerator, having an inlet pressure of 615 kPa gauge or 90 psigon the air side. The regenerators are designed to cover an operating temperaturerange from low cryogenic to very high temperatures. Metal regenerators are used foroperating temperatures up to abo ut 870C (1 600F); ceramic regen erators a re usedfor higher temperatures, up to 2000C (3600F). The maximum inlet temperature forpaper and plastic regenerators is 50C (120F).Regenerators have self-cleaning characteristics because hot and cold gases flow inopposite directions periodically through the same passage. As a result, compact regen-era tors have minimal fouling problems a nd usually have very small hydraulic diame terpassages. If severe fouling is anticipated, rotary regenerators are not used; fixed-matrixregenerators with large hydraulic diam eter flow passages [50 mm (2 in.)] could be used fo rvery corrosive/fouled gases a t ultrahigh temperatu res [925 to 1600C (1900 to 2900 F)I.Carryover an d bypass leakages from the ho t fluid t o the cold fluid (or vice versa) occur inthe regenerator. Where this leakage and subsequent fluid contamination is not permis-sible, regenerators ar e not used. Hence, they are no t used with liquids. Th e cost of therotary regenerator surface per unit of heat transfe r surface area is generally substantiallylower than tha t of a plate-fin o r tube-fin surface.

10.3 SOME QUANTITATIVE CONSIDERATIONSAs presented in Fig. 1.1, heat exchangers can be broadly classified according toconstruction as tubular, plate type, extended surface, and regenerative. A large vari-ety of high-performance surfaces are used on the gas side of extended surface andregenerative exchangers. A large number of enhanced tube geometries are availablefor selection in tubular exchangers. For the general category of enhanced tubes,internally finned tubes, and surface roughness, Webb and Bergles (1983) have pro-posed a number of performance evaluation criteria (PEC) to assess the performanceof enhanced surfaces compared to similar plain (smooth) surfaces. In plate-typeexchangers (used primarily with liquids), although many different types of construc-tion are available, the number of surface geometries used in modern exchangers islimited to high-performance chevron plate geometry, which is most commonly usedin PHEs. As a result, in this chapter we focus on quantitative screening methods for

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ReFIGURE 10.11(From Shah, 1983.)Comparison of surface basic characteristics of two heat exchanger surfaces.

gas flows in compact heat exchangers and performance evaluation criteria (PECs) fortubular surfaces. For extended surfaces, particularly the plate-fin type, selection ofthe surfaces on both fluid sides are independent of each other, and generally, onefluid side is critical from the pressure drop requirement. Hence, we consider only onefluid side for the surface selection for plate-fin surfaces.

10.3.1 Screening MethodsSurface selection is made by comparing the performance of various heat exchangersurfaces and choosing the best under some specified criteria (objective function andconstraints) for a given heat exchanger application. Consider the j and character-istics of surfaces A and B in Fig. 10.11. Surface A has both j and higher thanthose for surface B. Which is a better surface? This question is meaningless unlessone specifies the criteria for surface comparison. If the pressure drop is of lessconcern, surface A will transfer more heat than surface B for the same heat transfersurface area for a given application. If the pressure drop is critical, one cannot ingeneral say that surface A is better than surface B. One may need to determine, forexample, the volume goodness factors for comparison (see Section 10.3.1.2); or onemay even need to carry out a complete exchanger optimization after selecting thesurface on the other fluid side of a two-fluid exchanger.A variety of methods have been proposed in the literature for surface performancecom parison s. These me tho ds could be categorized as follows: (1) direct com par ison s of jand A (2) a comparison of heat transfer as a function of fluid pumping power, (3)miscellaneous direct comparison methods, and (4) performance comparisons with areference surface. Over 30 such dimensional or nondimensional comparison methodshave been reviewed critically by Shah (1978), and many more methods have beenpublished since then.It should be emphasized that mo st of these c omparisons are for the surfaces only onone fluid side of a h eat exchanger. When a com plete exchanger design is considered th atdo es no t lend itself to h aving o ne fluid side as a stro ng side (i.e., having high vohA) , hebest surface selected by the foregoing methods may not be an optimum surface for agiven applica tion. Th is is because th e selection o f the surfac e for the o ther fluid side an dits thermal resistance, flow arrange men t, overall exchanger envelope, an d oth er criteria

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SOME QUANTITATIVE CONSIDERATIONS 701(not necessarily related to the su rf ac ej and vs. Re characteristics) influence the overallperformance of a heat exchanger.In addition, if the exchanger is considered as pa rt of an (open or closed system), theexchanger surface (and/or other variables) may be selected based on the system as awhole rather tha n based on the optimum exchanger as a component. C urre nt methods ofsurface selection for an optimum heat exchanger fo r a system include the use of sophis-ticated computer program s th at take into account many possible effects. Such selection isnot possible in simplified approaches presented in the open literature. We focus onconsidering simple but important quantitative screening methods for surface selectionon the gas side of compact heat exchangers since these exchangers employ a large varietyof high-performance surfaces.The selection of a surface for a given application depends on exchanger design cri-teria. F or a specified heat transfer rate an d pressure d rop on one fluid side, two im por tantdesign criteria for compact exchangers (which may also be applicable to other exchan-gers) are the minimum heat transfer surface area requirement and the minimum frontalarea requirement. Let us first discuss the significance of these criteria.

To understand the minimum frontal area requirement, let us first review how the fluidpressure drop and heat transfer are related to the flow area requirement, the exchangerflow length, and the fluid velocity. The fluid pressure drop on one fluid side of anexchanger, neglecting the en trance/exit and flow acceleration/deceleration losses, isgiven from Eq . (6.29) as

(10.1)

Since predominantly developed and/or developing laminar flows prevail in compactheat exchangers, the friction factor is related to the Reynolds number as follows (seeSections 7.4.1.1 and 7.4.2.1):

(10.2)1 Re-' fo r fully developed laminar flow= { C . Re-'.' fo r developing lam inar flowwhere C1 and C2 are constants. Substituting Eq. (10.2) into Eq. (10.1) and notingthat Re = GDh/p, we get

LG for fully developed laminar flowfor developing laminar flowP { LGl.5 (10.3)Here G = m / A o . Therefore, for a specified constant flow rate m , the pressure drop isprop