Unit II Mechanical Design

UNIT II Thermal design of heat exchangers

The mechanical design is done by the mechanical engineers on the inputs of chemical engineers and an d us usin ing g th thee co codes des.. Th Thee mo most st wi wide dely ly us used ed co code de in Tu Tubul bular ar Ex Exch chang anger er Ma Manu nufa fact ctur ures es Associations (TEMA). This USA code along with ASME selection !!! (unfired p ressure "essel) code is used together for the mechanical design of the heat exchanger. The !ndian code for the heat exchanger design !S #$%&. 'ere we would discuss about the process design (or thermal design) leading to the siing of the heat exchanger. efore understanding design steps* it is necessary to understand the following for the heat exchanger. 8.2.1 Overall heat transfer coefficient

As understood by the pre"ious discussion that generally heat exchangers are tubular in nature (+ote, we are not discussing about plate type heat exchangers). Thus we can easily find out the o"erall heat transfer coefficient based on our pre"ious -nowledge. igure /.0 shows a simplest form (double pipe heat exchang exchanger) er) of tubul tubular ar heat exchang exchanger* er* where fluid A is being heated by fluid  in a co1current flow pattern. The inside and outside radii of the inner tube is represented r i and ro. The length of the exchanger exc hanger for heat transfer is considered as 2 for section 0 to /.

Fig.2,1: (a !chematic of a do"#le $i$e $ i$e heat exchanger (# thermal resistance net%or& for overall heat transfer

Thus the rate of heat transfer from the hot fluid to the cold fluid will be represented by e3.4.0*

The o"erall heat transfer coefficient5 ased on inside area of the inner pipe (e3.4./)

ased on outer side area of the outer pipe (e3.4.&)

 'h m"lti)$ass exchangers*

The simplest type of heat exchangers is double pipe heat exchangers* which is inade3uate for  flow rates that cannot readily be handled in a few tubes. !f se"eral double pipes are used in  parallel* the metal weight re3uired for the outer tubes becomes so large that the shell and tube construction* such as 010 exchanger will be helpful. !n that one shell ser"es for many tubes* is economical. The heat transfer coefficient of tube side and shell side fluid is "ery important and the indi"idual heat transfer coefficients must be high enough to attain high o"erall heat transfer  coefficient. As the shell would be 3uite large as compared to the tubes* the "elocity and the turbulence of the shell side fluid is important. !n contrast* the 010 exchanger has limitations also. 6hen the tube side flow is di"ided e"enly among all the tubes* the "elocity may be 3uite low* resulting in low heat transfer coefficient. There it may be re3uired to increase the area to ha"e the desired heat exchange for this low heat transfer coefficient. The area may be increased b y increasing the length of the tube. 'owe"er* the tube length re3uirement may be impractical for a gi"en situation. Thus the number of tubes should be increased without increased the tube length. The increased number of tubes would also  pro"ide the increased "elocity in the shell side resulting in the higher heat transfer coefficient. Therefore* multi1pass construction is needed* which would permit to use the practical and standard tube lengths. 'owe"er* the disad"antages are that* 0. /.

The construction of the exchangers become complex. 7arallel flow cannot be a"oided.

&.

Additional friction losses may occur.

!t should be noted that generally e"en number of tube passes are used in multi pass exchanger.

+T- correction factor

!n the earlier chapter* we ha"e seen for co1current or counter current flow system. The a"erage dri"ing force for heat transfer was defined by log mean temperature difference (2MT8). Thus the 2MT8 can be used for 010 exchangers for co1current and counter current. 'owe"er* for multi  pass exchangers (01/* /1#* etc.) the fluids are not always in co1current or counter current flow. The de"iation for co1current or counter current flow causes a change in the a"erage dri"ing force. Therefore* in order to use true heat transfer dri"ing force* a correction factor is re3uired into the 2MT8. Thus* the heat transfer rate can be written as (e3.4. 4)* q = U d A(F    T ΔT    m ) where*

Ud 9 o"erall heat transfer coefficient including fouling:dirt A 9 heat transfer area T ;Tm 9 true a"erage temperature difference. T 9 2MT8 correction factor 

 !t is to be noted that the following assumption ha"e been considered for de"eloping 2MT8* 0. The o"erall heat transfer coefficient is constant throughout the exchanger  /. !n case any fluid undergoes for phase change (e.g.* in condenser)* the phase change occurs throughout the heat exchanger and the constant fluid temperature pre"ails throughout the exchanger. &. The specific heat and mass flow rate and hence the heat capacity rate* of each fluid is constant. #. +o heat is lost in to the surroundings. $. There is no conduction in the direction of flow neither in the fluids nor in the tube or shell walls. <. Each of the fluids may be characteried by a single temperature* at any cross section in the heat exchanger that is ideal trans"erse mixing in each fluid is presumed. T* the 2MT8 correction factor can be directly obtained from a"ailable charts in the literature. These charts were prepared from the results obtained theoretically by sol"ing the temperature distribution in multi1pass heat exchangers. igures /.0 and /./ show the two generally used heat exchangers and their corresponding plots for finding T. !t may be noted that the gi"en figures ha"e the representati"e plots and any standard boo- on heat transfer may be consulted for the accurate results.

!t should be noted that in case of condensation or e"aporation the correction factor becomes unity (T 90). 6hile designing a heat exchanger* the rule of thumb is that the T should not be less than %.4.

Fig. 2.1: FT $lot for 1)2 exchanger t: cold fl"id in the t"#e T: hot fl"id in the shell 1: inlet 2: o"tlet

Fig. 2.2: FT $lot for 2)/ exchanger t: cold fl"id in the t"#e T: hot fl"id in the shell 1: inlet 2: o"tlet

Individ"al heat transfer coefficient

!n section* we ha"e seen that the o"erall heat transfer coefficient can be calculated pro"ide the  parameters are -nown including indi"idual heat transfer coefficients. !n this* section we will discuss how to find out the indi"idual heat transfer coefficient* which is basically based on the well1established correlations and discussed earlier also. The heat transfer coefficient (hi) for the tube side fluid in a heat exchanger can be calculated either by Sieder1Tate e3uation or by =olburn e3uation discussed in earlier chapter. 'owe"er* the shell side heat transfer coefficient (ho) cannot be so easily calculated because of  the parallel* counter as well as cross flow patterns of the fluid. Moreo"er* the fluid mass "elocity as well as cross sectional area of the fluid streams "ary as the fluid crosses the tube bundle. The lea-ages between baffles and shell* baffle and tubes* short circuit some of the shell fluid thus reduces the effecti"eness of the exchanger. >enerally* modified 8onohue e3uation (e3.4.?) (suggested by 8.@. ern) is used to predict the ho*

where* h%9 shell side heat transfer coefficient 8h9 hydraulic diameter of the shell side - %9 thermal conducti"ity of the shell side fluid >s9 mass flow rate of the shell side The 8h and >s can be easily calculated if the geometry of the tube arrangement in the shell is -nown. The tubes may be generally arranged as a s3uare or triangular pitch* as shown in figure

Fig.2.0: T"#e arrangement in the shell (a triang"lar $itch (# s"are $itch

The hydraulic diameter ( Dh) for tubes on s3uare pitch9

 Dh or <%B triangular pitch9

where*

d 0 9 outer diameter of tube  p 9 tube pitch

where,

= fow rate o shell fuid as =

shell side fow area

Shell side flow area can be calculated using baffle information number of tubes in the shell and tube arrangement. !f /$C cut baffles are used* that means the shell side flow will be from this /$C area. 'owe"er we ha"e to reduce the area of the pipes which are accumulated in this opening. So depending upon the information we may determine the shell side fluid flow area. !t may also be found out by the following way*

where* = 9 tube clearance p 9 pitch of the tube

 9 baffle spacing

8s9 inside diameter of shell

Fo"ling factor or dirt factor D"er a time period of heat exchanger operation the surface of the heat exchanger may be coated  by the "arious deposits present in the flow system. Moreo"er* the surfaces may become corroded or eroded o"er the time. Therefore* the thic-ness of the surface may get changed due to these deposits. These deposits are -nown as scale. These scales pro"ide another resistance and usually decrease the performance of the heat exchangers. The o"erall effect is usually represented by dirt factor or fouling factor* or fouling resistance* f (Table 4.0) which must ha"e included all the resistances along with the resistances due to scales for the calculation of o"erall heat transfer  coefficient. The fouling factor must be determined experimentally using e3. 4.#*

Thus to determine the  f * it is "ery important to -now U clean for the new heat exchanger. The Uclean must be -ept securely to obtain the  f*  at any time of the exchangerFs life. Table-8.1 Fouling factor of a few of the industrial uids

 Tem$erat"re $rofiles in U)T"#e heat exchangers

ig. 4.? shows the temperature profile along the length of a 01/ exchangers and /1# exchangers. exchangers.

Fig.8. Tem$erat"re)length c"rve corres$onding to (a 1)2 exchanger  (# 2)/ exchanger

The nomenclature used in the fig.4.? is described below Tha, !nlet temperature of hot fluid Thb, Dutlet temperature of hot fluid Tca, !nlet temperature of cold fluid Tcb, Dutlet temperature of cold fluid Tci, !ntermediate temperature of cold fluid !n the abo"e arrangement it is assumed that the hot fluid is flowing in the shell side and cold fluid is flowing in the tube side of the exchangers. The fig.4.? (a) shows the 01/ exchangers in which the hot fluid enter into the exchanger from the left side and exits from the right side. The cold fluid enters concurrently that is from the left side to the tube of the exchangers and goes up

to right end of the exchangers and returns bac- to ma-e two tube pass* and exits from the left end of the exchangers. The temperature profile all along the length of the exchanger is shown in the corresponding temperature length profile. igure4.? (b) shows the flow direction and corresponding temperature length profile for /1# exchangers. The shell side fluid two passes and the tube side fluid has #1passes in the exchangers. !t can be easily understood that whene"er the number of passes is more than one* the flow cannot  be truly co1current or counter current. Thus it will be a mix of co1current and counter current flows in any multi pass heat exchangers. Though the temperature profile of the hot and cold streams can be easily predictable for single  pass heat exchangers but for the complex flow modes* the prediction of temperature distribution will be difficult as shown in fig.4.?. As can be seen when 01/ exchang ers was (fig.4.? (a)) used in co1current mode* the temperature profile was gi"en in the figure. 'owe"er* if the fluid streams enter in counter current mode a temperature cross may occur sometimes. Temperature cross is described as the positi"e temperature difference between the cold and the hot fluid* when these fluids lea"e the exchangers. !n that case the cold fluid will attain the maximum temperature inside the exchanger instead of at the exit (fig.4.0%).

Fig. 8.10: 1-2 ow pattern and temperature prole in echanger showing cross ow

At this temperature cross* the cold fluid temperature reaches the maximum at a point inside the exchanger and not at its exits. This temperature cross point also coincides with the point of  intersection of the temperature profile of the hot fluid and the co1current one of the cold fluid. The difference (Tc/ 1 Th/) is called the temperature cross of the exchanger. 'owe"er* if the

temperature cross does not appear then the (Tc/ 1 Th/) is called the approach. Moreo"er* on careful e"aluation it can be seen that for the multi shell side pass a significant length of the exchanger ha"e cross flow pattern in the tube flow when the shell side fluid is migrating from one shell pass to another shell pass. Thus calculating heat transfer co1efficient for shell side  becomes little challenging and will be explained in section 4./.<. Although the parallel flow or  counter flow are 3uite similar* the parallel flow and counter flow heat exchangers differ greatly in the manner in which the fluid temperatures "ary as the fluid pass through. The difference can  be understood in the figure 4.00. The fig.4.00 shows an important parameter* mcp* the product of mass flow rate (m) and the specific heat* cp* of the fluids. The product mcp is called the rate of heat capacity. The o"erall energy balance of the heat exchanger gi"es the total heat transfer between the fluids* 3* expressed by e3.4.$*

The fig.4.00 shows the relati"e "ariation of the two fluid temperatures through the heat exchanger* which is influenced by whether is greater or less than . !n particular* for counter flow* examination of the s-etches in fig.4.00 shows that limiting condition for  maximum heat transfer is determine by whether G

is greater or less than

. 6hen*

the maximum possible heat transfer is determined by the fact that the hot fluid can be cooled

to the temperature of the cold fluid inlet. Thus* for

G

Fig.8.11: Tem$erat"re $rofiles of (a $arallel flo%, and (# co"nter flo%, for different

ine"alities

FOU+IN3 F45TO6 

Material deposits on the surfaces of the heat exchanger tube may add further resistances to heat transfer in addition to those listed abo"e. Such deposits are termed fouling and may significantly affect heat exchanger performance. The heat exchanger coefficient* Uc* determined abo"e may  be modified to include the fouling factor f. !caling   is the most common form of fouling and is associated with in"erse solubility salts. Examples of such salts are =a=D&* =aSD#* =a&(7D#)/* =aSiD&* =a(D')/* Mg(D')/* MgSiD&*  +a/SD#* 2iSD#* and 2i/=D&. The characteristic which is termed in"erse solubility is that* unli-e most inorganic materials* the solubility decreases with temperature. The most important of  these compounds is calcium carbonate* =a=D&. =alcium carbonate exists in se"eral forms* but one of the more important is limestone. The material fre3uently crystallies in a form closely resembling marble* another form of calcium carbonate. Such materials are extremely difficult to remo"e mechanically and may re3uire acid cleaning. 5orrosion fo"ling  is classified as a chemical reaction which in"ol"es the heat exchanger tubes. Many metals* copper and aluminum being specific examples* form adherent oxide coatings which ser"e to passi"ate the surface and pre"ent further corrosion. Metal oxides are a type of  ceramic and typically exhibit 3uite low thermal conducti"ities. E"en relati"e thin coatings of  oxides may significantly affect heat exchanger performance and should be included in e"aluating o"erall heat transfer resistance. 5hemical reaction fo"ling  in"ol"es chemical reactions in the process stream which results in deposition of material on the heat exchanger tubes. 6hen food products are in"ol"ed this may be termed scorching but a wide range of organic materials are subHect to similar problems. This is commonly encountered when chemically sensiti"e process fluids are heated to temperatures near that for chemical decomposition. ecause of the no flow condition at the wall surface and the temperature gradient which exists across this laminar subIlayer* these regions will operate at somewhat higher temperatures than the bul- and are ideally suited to promote fa"orable conditions for such reactions. Free7ing fo"ling  is said to occur when a portion of the hot stream is cooled to near the freeing  point for one of its components. This is most notable in refineries where paraffin fre3uently solidifies from petroleum products at "arious stages in the refining  process* obstructing both flow and heat transfer. iological fo"ling  is common where untreated water is used as a coolant stream. 7roblems range from algae or other microbes to barnacles. 8uring the season where such microbes are said to  bloom* colonies se"eral millimeters deep may grow across a tube surface "irtually o"ernight* impeding circulation near the tube wall and retarding heat transport. iewed under a microscope* many of these organisms appear as loosely intertwined fibersImuch li-e the form of fiberglass insulation Traditionally these organisms ha"e been treated which chlorine* but present day concerns on possible contamination to open water bodies has se"erely restricted the use of  oxidiers in open discharge systems.

9artic"late fo"ling  results from the presence of rownian sied particles in solution. Under  certain conditions such materials display a phenomenon -nown as thermophoresis in which motion is induced as a result of a temperature gradient. Thermodynamically this is referred to as a cross1coupled phenomenon and may be "iewed as being analogous to the Seabec- effect. 6hen such particles accumulate on a heat exchanger surface they sometimes fuse* resulting in a  buildup ha"ing the texture of sandstone. 2i-e scale these deposits are difficult to remo"e mechanically. Most of the actual data on fouling factors is tightly held be a few specialty consulting companies. The data which is commonly a"ailable is sparse. An example is gi"en below,

9ress"re dro$ in the heat exchanger 7ressure drop calculation is an important tas- in heat exchanger design. The pressure drops in the tube side as well as shell side are "ery important and 3uite a few co1relations are a"ailable in the literature. Dne such co1relation is gi"en below in the subse3uent subsection.

4./.J.0 =orrelation for tube side pressure drop (e3. 4.0%)

where* ;7t*f 9 total pressure drop in the bundle of tube f 9 friction factor (can be found out from MoodyFs chart) >t 9 mass "elocity of the fluid in the tube 2 9 tube length n 9 no of tube passes g 9 gra"itational acceleration Kt 9 density of the tube fluid di 9 inside diameter of the tube m 9%.0# for e G /0%% %./$ for e L /0%%

The abo"e correlation is for the pressure drop in the tubes owing to the frictional losses. 'owe"er in case of multi pass flow direction of the flow in the tube changes when flow is from 01pass to another pass and the pressure losses due to the change in direction is called return1loss. The return1loss (;7t*r ) is gi"en by e3.4.00*

n 9 no of tube pass "t 9 "elocity of the tube fluid Kt 9 density of the tube fluid Therefore* the total tube side pressure drop will be* ;pt 9 ;7t*f  ;7t*r  8.2..2 5orrelation for shell side $ress"re dro$

The following correlation (e3.4.0/) may be used for an unbaffled shell*

The abo"e e3uation can be modified to the following form (e3.4.0&) for a baffled shell*

6here* 2 9shell length ns 9 no of shell pass n b 9 no of baffles Ks 9 shell side fluid density >s 9 shell side mass "elocity 8h 9 hydraulic diameter of the shell Dsi = inside diameter o shell

 s = shell side riction actor

The hydraulic diameter (8h) for the shell can be calculated by the following e3uation (e3. 4.0#)*

where,

nt = number o tubes in the shell do = outer diameter o the tube

 The riction actor ( s) can be obtained by the Moody’s chart or the corresponding Reynolds number

;eat transfer effectiveness and n"m#er of transfer "nits (NTU The 2MT8 is re3uired to be calculated for the e"aluation of heat exchanger performance. 'owe"er* the 2MT8 cannot be directly calculated unless all the four terminal temperatures (Tc*i* Tc*o* Th*i* Th*o) of both the fluids are -nown.

Sometimes the estimation of the exchanger performance (3) is re3uired to be calculated on the gi"en inlet conditions* and the outlet temperature are not -nown until 3 is determined. Thus the  problem depends on the iterati"e calculations. This type of problem may be ta-en care of using  performance e3ui"alent in terms of heating effecti"eness parameter (N)* which is defined as the ratio of the actual heat transfer to the maximum possible heat transfer. Thus*

7D2EM, 0. ind the D"erall heat transfer coefficient for a shell and tube counter flow heat exchanger where the heat exchanged is 00./J >O:hour with the heat transfer area of ?< m/. Assume the 2MT8 as J
!n a double pipe heat exchanger hot fluid is entering at //%B= and lea"ing at 00$B=. =old fluid enters at 0%B= and lea"es at J$B=. Mass flow rate of hot fluid 0%% -g:hr* =p of hot fluid 0.0 -cal:-gB=. =p of cold fluid %.?$ -cal:-gB=. =alculate 2MT8P a) !f the flow is parallel  b) !f the flow is counter current c) ind the mass flow rate of cold fluid if the heat loss during the exchange is $C.

&. !n a parallel flow double1pipe heat exchanger water flows through the inner pipe and is heated from /%B= to J%B=. Dil flowing through the annulus is cooled from /%%B= to 0%%B=. !t is desired

to cool the oil to a lower exit temperature by increasing the length of the heat exchanger. 8etermine the minimum temperature to which the oil may be cooledP #. The flow rates of hot and cold water streams running through a parallel flow heat exchanger are %./ -g:. and %.$ -g:s respecti"ely. The inlet temperatures on the hot and cold sides are J$B= and /%B= respecti"ely. The exit temperature of hot water is #$B=. !f the indi"idual heat transfer  coefficients on both sides are <$% w:m/B=* calculate the area of the heat exchanger. $. An oil cooler for a lubrication system has to cool 0%%% -g:h of oil (cp 9 /.%?-O:-gB=) from 4B= to #DB= by using a cooling water flow of 0%%% -g:h at &%B=. >i"e your choice for a parallel flow or counter1flow heat exchanger* with reasons. =alculate the surface area of the heat exchanger* if  the o"erall heat transfer coefficient is /# 6:m/B=. Ta-e =p of water 9 #.04 -O:-gB=. <. A counter1flow double pipe heat exchanger using superheated steam is used to heat water at the rate of 0%$%% -g:h. The steam enters the heat exchanger at 04%B= and lea"es at 0&DB=. The inlet and exit temperatures of water are &%B= and 4%B= respecti"ely. ! f o"erall heat transfer  coefficient from steam to water is 40# 6:m/B=* calculate the heat transfer area. 6hat would be the increase in area if the fluid flows were parallelP J. The amount of Q0/ used in compression refrigeration system is # tonnes:hour. The brine flowing at 4$% -g:min* with inlet temperature of 0/B= is cooled in the e"aporator. Assuming Q0/entering and lea"ing the e"aporator as saturated li3uid and saturated "apour respecti"ely* determine the area of e"aporator re3uired. Ta-e the following properties, or Q0/ , Saturation temperature , 1 /&B=5 cp 9 0.0 J -O:-gB=5 hfg 9 0
Im$ortance: 5alc"lation and designing of the heat exchanger

8ouble1pipe heat exchanger  The following steps may be used to design a double1pipe heat exchanger  0. =alculate 2MT8 from the -nown terminal temperatures. /. 8iameter of the inner and outer pipes may be selected from the standard pipes from the literature (generally a"ailable with the "endor and gi"en in the boo-s). The selection thumb rule is the consideration of higher fluid " elocity and low pressure drop in the pipe.

&. =alculate the eynolds number and e"aluate the heat transfer coefficient* hi* using the co1 relations gi"en in the chapter. #. Similarly* calculate the eynolds number of the fluid flowing through annulus. =alculation the e3ui"alent diameter of the annulus and find the outside heat transfer  coefficient* ho. $. Using hi and ho* calculate the o"erall heat transfer coefficients. +ote that it will be a clean o"erall heat transfer coefficient. !n order to find design outside heat transfer  coefficient using a suitable dirt factor or fouling factor. The tube fouling factor is suggested by TEMA (table 4.0). The calculations are based on trial and error. !f the heat transfer coefficient comes out to be "ery small or the pressure drop comes out to be "ery high* this procedure to be redone for different set of diameters in the step0. !hell and t"#e heat exchanger:

The shell and tube heat exchanger also in"ol"es trial and error but it is not as simple as in case of  double pipe heat exchanger. The design of shell and tube heat exchanger includes* a, heat transfer re3uired for the gi"en heat duty  b, tube diameter* length* and number* c, shell diameter* d, no of shell and tube passes* e, tube arrangement on the tube sheet and its layout* and f, baffle sie* number and spacing of the baffles. The calculation of 2MT8 can be done if the terminal temperatures are -nown. 'owe"er* the design heat transfer co1efficient (i.e.* heat transfer co1efficient including fouling factor) and the area are dependent on each other and thus challenges in"ol"e for the estimation. The also depends upon eynolds number* which depends upon the li3uid flow rate* sies and the number  of tubes. Therefore* is a function of diameter and the no of tubes and the parameter pro"ides the area. Moreo"er* can also be calculated is based on shell side co1efficient but then it re3uires tube number* diameters and pitch. Thus* the abo"e discussion shows that and A are not fully explicit and re3uires trial and error method of calculation. The guideline for shell1and1tube calculation is shown in b elow* energy balance and exchanger heat duty calculation* 0. find all the thermo1physical properties of the fluid* /. ta-e initial guess for shell1and1tube passes* &. calculate 2MT8 and T* #.

$.

<.

J.

4. ?.

0%.

00.

0/.

Assume (or select) Udirty* based on the outside tube area. =alculate corresponding heat transfer area* A. Select tube diameter* wall thic-ness and the tube length. ased on this "alues and heat transfer area* find out the no of the tubes re3uired. Assume the tube pitch and assume diameter of the shell* which can accumulate the no of  tube. +ow* select the tube1sheet layout. Select the baffle design. Estimate hi and ho* if the estimated shell1side heat transfer coefficient (ho) appears to be small5 the baffles at a close distance may be tried. !f the tube side co1efficient (hi) is low* the number of tube passes to be reconsidered such that the eynolds number increases (for a reasonable ;p) and henceforth hi. E"aluate Uclean on the outside tube area basis. Select a suitable fouling factor (d) and find Udirty =ompare Udirty and A "alues with the "alues assumed in step ($). !f Acalculate R Aassumed* it may be acceptable. Dtherwise a new configuration in terms of the sie and no of the tubes and tube passes* shell diameter is assumed and recalculation be done =alculate the tube1side and shell side ;p. !f ;p is more than the allowable limit* the re1 calculate after suitable adHustment has to be done.

0. 6hat are the heat transfer mechanisms in"ol"ed during heat transfer from the hot fluid to the cold fluidP /. !n heat exchange between air and water across a tube wall* it is proposed to use fins to enhance the o"erall heat transfer coefficient. 6ould you put the fins on the air side or on the water sideP &. 6hen is a heat exchanger classified as compactP #. 'ow does a cross flow heat exchanger differ from a counter flow oneP $. 6hat is the role of baffles in a shell1and1tube heat exchangerP 6hat is the implication about  pressure dropP <. Under what conditions is the effecti"eness +TU method preferred o"er 2MT8 method as a method of analysis of a heat exchangerP J. =an temperature of the hot fluid drop below the inlet temperature of the cold fluid at any location in a heat exchangerP 4. =an temperature of the cold fluid rise abo"e the inlet temperature of the hot fluid at any location in a heat exchangerP ?. =onsider two double pipe counterflow heat exchangers that are identical except that one is twice as long as the other one. 6hich of the exchangers is more li-ely to ha"e a higher  effecti"enessP 0%. =an effecti"eness be greater than oneP 00. Under what conditions can a counter flow heat exchanger ha"e an effecti"eness of oneP 6hat would be your answer for a parallel flow heat exchangerP