Experimental Thermal and Fluid Science 36 (2012) 242–248

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Experimental Thermal and Fluid Science

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Experimental studies on a corrugated plate heat exchanger for smalltemperature difference applications

M. Faizal 1, M.R. Ahmed ⇑Division of Mechanical Engineering, Faculty of Science, Technology & Environment, The University of the South Pacific, Laucala Campus, PO Box 1168, Suva, Fiji

a r t i c l e i n f o a b s t r a c t

Article history:Received 10 March 2011Received in revised form 19 July 2011Accepted 28 September 2011Available online 4 October 2011

Keywords:Heat transferCorrugated plate heat exchangerChannel spacing

0894-1777/$ - see front matter � 2011 Elsevier Inc. Adoi:10.1016/j.expthermflusci.2011.09.019

⇑ Corresponding author. Tel.: +679 3232042; fax: +E-mail addresses: faizal_m@usp.ac.fj (M. Faiza

Ahmed).1 Tel.: +679 3232875; fax: +679 3231538.

Experimental studies were performed on a corrugated plate heat exchanger for small temperature differ-ence applications. Experiments were performed on a single corrugation pattern on 20 plates arrangedparallelly, with a total heat transfer area of 1.16298 m2. The spacing, DX, between the plates was varied(DX = 6 mm, 9 mm, and 12 mm) to experimentally determine the configuration that gives the optimumheat transfer. Water was used on both the hot and the cold channels with the flow being parallel andentering the heat exchanger from the bottom. The hot water flowrates were varied. The cold side flowrateand the hot and cold water inlet temperatures were kept constant. It is found that for a given DX, theaverage heat transfer between the two liquids increases with increasing hot water flowrates. The corru-gations on the plates enhance turbulence at higher velocities, which improves the heat transfer. The opti-mum heat transfer between the two streams is obtained for the minimum spacing of DX = 6 mm. Thepressure losses are found to increase with increasing flowrates. The overall heat transfer coefficients,U, the temperature difference between the two stream at outlet, and the thermal length are also pre-sented for varying hot water flowrates and DX. The findings from this work would enhance the currentknowledge in plate heat exchangers for small temperature difference applications and also help in thevalidation of CFD codes.

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1. Introduction

Heat exchangers are heat transfer devices that exchange ther-mal energy between two or more media. The heat transfer betweenthe media is purely based on temperature difference, without theuse of any external energy. Some of the applications of heatexchangers are in power production industries, chemical and foodindustries, electronics, waste heat recovery systems, manufactur-ing industries, and air-conditioning and refrigeration systems.There are basically two types of heat exchangers: a direct heat ex-changer and an indirect heat exchanger. In a direct heat exchanger,the two media between which heat is exchanged are in direct con-tact, e.g. cooling towers. In an indirect heat exchanger, the twomedia between which heat is exchanged are separated by a wall[1,2]. A plate heat exchanger is an indirect heat exchanger. Plateheat exchangers comprise of a stack of corrugated or embossedmetal plates with inlet and outlet ports and seals to direct the flowin alternate channels. The flow channels are formed by adjacentplates [3]. As shown in Fig. 1, the hot and cold fluids flow in

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679 3231538.l), ahmed_r@usp.ac.fj (M.R.

alternate channels and the heat transfer takes place between adja-cent channels [4]. The number and size of the plates are deter-mined by the flowrates, the physical properties of the fluids,pressure drops, and heat transfer requirements [3,5]. There are alsomany flow patterns that can be achieved for plate heat exchangers[3].

In the analysis of heat exchangers, all the thermal resistances inthe path of heat flow from one fluid to another are combined into asingle resistance [6], and an overall heat transfer coefficient, U, ofthe heat exchanger is determined. The overall heat transfer coeffi-cient is a measure of the resistance to heat flow from one mediumto another [2]. Phase change processes in heat exchangers havevery high U values due to high thermal conductivities. Because ofcomplex physical processes, it is not generally possible to predictaccurate values of U. Therefore, empirical formulas and U valuesare mostly derived from experimental data [7]. One of the require-ments is in ocean thermal energy conversion (OTEC) plants iseffective heat transfer with minimum pressure loss for small tem-perature difference between the hot and cold fluids (20–25 �C).Pressure losses in heat exchangers will affect the pumping powerof the pumps in OTEC plants. Studies reported in Refs. [8,9] showthat pressure drop increases significantly with flowrates.

Plate heat exchangers have many advantages compared tomany other heat exchangers. Plate heat exchangers can be used

Nomenclature

A total heat transfer area (m2)CpCW specific heat at average cold water temperature (kJ/

kg �C)CpHW specific heat at average hot water temperature (kJ/

kg �C)_VCW cold water flowrate (L/s)_VHW hot water flowrate (L/s)_QCW heat transferred by cold water (W)_QHW heat transferred by hot water (W)_QAverage average heat transfer between hot and cold water (W)TCWI cold water temperature at inlet (�C)TCWO cold water temperature at outlet (�C)THWI hot water temperature at inlet (�C)THWO hot water temperature at outlet (�C)

U overall heat transfer coefficient (W/m2 K)DPH pressure loss of hot water (kPa)DTCW temperature change of cold water (�C)DTHW temperature change of hot water (�C)DTm log mean temperature difference (LMTD)DToutlet temperature difference of hot and cold water measured

at outlet of heat exchanger (�C)DX plate spacing (mm)qCW water density at average cold water temperature (kg/

m3)qHW water density at average hot water temperature (kg/m3)hCW thermal length of the cold water channelshHW thermal length of the hot water channelshAverage average thermal length

M. Faizal, M.R. Ahmed / Experimental Thermal and Fluid Science 36 (2012) 242–248 243

for high-viscosity applications, because turbulence is induced atlow velocities which leads to effective heat transfer [3]. They alsohave high thermal effectiveness, large heat transfer per unit vol-ume, low weight, possibility of heat transfer between manystreams, ease of maintenance and a compact design [5,10].

2. Background

Corrugations in plate heat exchangers improve the heat transferrates by 20–30% by increasing the heat transfer area and byenhancing turbulence at low flowrates [7,11]. The corrugatedplates also improve the mechanical strength of the plates [4]. Manytypes of enhanced surface geometries are used on plate heatexchangers. The objective is to obtain high heat-transfer coeffi-cients without correspondingly increased pressure-loss penalties[12]. Special channel shapes, such as the wavy channels, providemixing due to secondary flows or boundary layer separation withinthe channel [3]. The corrugations or wavy fins induce secondaryflows (Görtler vortices) which assist in heat transfer augmentation[13]. The efficiency of plate heat exchangers can be improved bymodifying the boundary layer and by enlarging the surfaces [14].

Since wavy surfaces have noninterrupted walls in each flowchannel, the chances of fouling of and particulates being caughtin the channels are less. The waveform in the flow direction dis-rupts the flow and induces very complex flows. Görtler vorticesare formed as the fluid passes over the concave wavy surfaceswhich enhance heat transfer. In the low-turbulence regime (Re ofabout 6000–8000), the wall corrugations increase the heat transferby about nearly three times compared with the smooth wall chan-nel [3]. Therefore, wavy fins are often a better choice at the higher

Fig. 1. Hot and cold fluid flow in alternate passages in plate heat exchangers.

Reynolds numbers. A basic form of a corrugated or wavy geometryis shown by Fig. 2. As corrugation (or wave) height to wavelengthincreases, the separation zones in the troughs increase in relativesize, giving rise to disproportionately high pressure drop [13]. Avariety of corrugated or wavy patterns are proposed for plate heatexchangers [3].

Picon-Nuñez et al. [5] presented a methodology on the design ofcompact heat exchangers. A simple approach to surface selection ofthe heat exchangers is based on the volume performance index.Plain-fin (wavy configuration) and louvered fins were consideredin their study. They presented the volume performance index atdifferent Reynolds numbers. Several studies have been carriedout on heat transfer enhancement using corrugated plate heatexchangers. Tauscher and Mayinger [14] carried out numericaland experimental studies on heat transfer enhancement in plateheat exchangers with rib-roughened surfaces, which are also wavyconfigurations. They tested for various configurations of the ribs:shape, width, height, groove angle, spacing, angle, and arrange-ment patterns. They found out that the ribs show their best effectsin regions where they can induce turbulence. They generalized thatturbulence promoters (ribs in this case) show best performance inthe transition region from laminar to turbulent flow.

Ciofalo et al. [15] conducted studies of flow and heat transfer incorrugated – undulated plate heat exchanges for rotary regenera-tors. For a particular corrugation, they varied the angle betweenthe main flow direction and the axes of the furrows of the corruga-tions. They presented the Nusselt number distributions, the fric-tion coefficient, pressure drop and heat transfer characteristics,and numerically simulated results on the flow and thermal fieldsinduced by the wavy configurations. Kanaris et al. [16] performedCFD studies on a plate heat exchanger comprising of corrugatedwalls with herringbone design. They visualized the complex swirl-ing flow in the furrows of the corrugations, and the Nusseltnumber and the friction factor were compared with those ofsmooth plates. They reported that corrugations increase the heattransfer; however, the pressure losses also increase. Elshafei et al.

Fig. 2. Schematic geometry of corrugated surfaces (K is wavelength or pitch, b isplate spacing, and w is amplitude or channel height) [15].

244 M. Faizal, M.R. Ahmed / Experimental Thermal and Fluid Science 36 (2012) 242–248

[17] presented heat transfer and pressure drop results in corru-gated channels. They discussed the effect of channel spacing andphase shift of the corrugations on the heat transfer and the pres-sure drop. They showed that corrugations enhance heat transferbut with accompanying pressure drops. The results from theexperiments were compared with conventional parallel plate heatexchangers and they found that corrugations enhance the heattransfer significantly. They found that the friction factor is higherfor higher values of channel spacing. They also concluded thatthe area goodness factor decreases with increasing spacing ratio.Sparrow and Comb [18] also performed experimental studies oncorrugated plates and variable spacings.

Collins et al. [19] investigated the flow and heat transfer in cor-rugated passages. An experimental and numerical study of flowand heat transfer was conducted for a crossed-corrugated geome-try. The effects of corrugation angle, geometry, and Reynolds num-ber were investigated. Mitsumori et al. [20,21] compared theperformance of a closed cycle ocean thermal energy conversion(OTEC) plant using plate-type heat exchangers and tube-type heatexchangers. The results of their studies show that plate-type heatexchangers have more advantages and that they can be more com-pact. Test results on plate heat exchangers done at Saga University,Japan are presented in Ref. [7]. It was found that the overall heattransfer coefficients and the pressure losses generally increase asthe water velocity is increased. The best configurations tested atSaga University increase the overall heat transfer by a factor of 4in comparison with smooth plates. Lyytikäinen et al. [22] per-formed numerical studies for varying corrugation angles and cor-rugation lengths and found out that both heat transfer as well aspressure drop increase as the corrugation angle is increased. Theyhave stated that it is not easy to find a specific geometry that pro-vides both a low pressure drop and a high heat transfersimultaneously.

Fig. 3. Geometric details of the heat exchanger plates (dimensions in mm).

Table 1Geometric details of the heat exchanger plates.

Detail Dimension

3. Objectives

From the previous research carried on heat transfer enhance-ment, it is obvious that wavy corrugations for plate heat exchang-ers are an attractive option. On the basis of the above finding, thepresent work is aimed at experimentally studying the heat transfercharacteristics (with pressure drops) for corrugated plate type heatexchangers for use in small temperature difference applications.The results from this work will also be useful for the design of heatexchangers for OTEC applications where the objective is same –maximum heat transfer between two fluids having a temperaturedifference of 20–25 �C keeping the pressure loss at a minimum.The current design is chosen based on the enhancement of heattransfer characteristics due to the incorporation of wavy configura-tions in plate exchangers. The traditional geometry of the wavyconfigurations is retained to reduce the number of variables inthe present work and to study the effect of the flow rate and spac-ing. The hot water flowrates, _VHW; and the spacing between theplates, DX, are varied while the corrugations remain the same.The focus of the experiments is to measure the temperature ofthe two fluids at inlet and exit of the heat exchanger and then todetermine which DX value gives optimum heat transfer. A detailedphysical explanation of the flow and the enhanced turbulence bythe corrugations in the channels is also presented.

Plate height H 273 mmPlate width W 213 mmNo. of plates N 20No. of hot channels NH 9No. of cold channels NC 10Total area A 1.16298 m2

Spacing DX 6 mm, 9 mm, 12 mm

4. Experimental set-up and procedure

The experiments were carried out in the thermo-fluids labora-tory at the University of the South Pacific. Experiments were per-formed on a single corrugation pattern on 20 plates arrangedparallelly. The spacing between the plates, DX, was varied to

experimentally determine the spacing that gives the optimum heattransfer. Water was used on both the hot and the cold channelswith the flow being parallel. Both the hot and cold water enteredthe heat exchanger from the bottom. This allowed the water tofully fill the heat exchanger channels before exiting into the atmo-sphere, thus utilizing the full area of the plates for effective heattransfer and preventing the formation of hydraulic diameters.The flowrates, _VHW; for the hot side were varied from 0.18 L/s to0.63 L/s, while the cold side flowrate, _VCW, was kept constant at0.16 L/s. The inlet temperatures for both the hot and cold waterwere kept constant at 49 �C and 26 �C respectively. This gives atemperature difference of 23 �C at the inlet of the heat exchanger.The plates used are corrugated galvanized sheets, with a thicknessof 0.4 mm. The other geometric details of the plates and the heatexchanger are provided in Fig. 3 and Table 1.

A steam generator is used to maintain a constant temperatureof 49 �C in the hot water tank. The hot water is directed into theheat exchanger by a centrifugal pump with a rated capacity of81 L/min at a total head of 21 m and driven by a 0.5 HP variablespeed motor. The inlet temperature of the cold water is maintainedat a constant temperature of 26 �C. CABAC T6201 digital thermom-eters, with a resolution of 0.1 �C and a temperature range of �50 �Cto +250 �C, were mounted at the inlets and outlets of the heatexchangers. WIKA EN 837-1 pressure gauges, with an accuracy of1%, pressure range of 0–100 kPa, and a temperature range of�20 �C to 60 �C, mounted at the inlet of the hot and cold waterstreams measure the gauge pressure at which the fluids enterthe heat exchanger. Fig. 4 shows a schematic of the experimentalsetup. The repeatability of the temperature measurements waswithin 4% and that of pressure measurements was within 2.4%.The accuracies of measurement or estimation of q, Cp, _V and tem-peratures were taken into consideration for estimating the uncer-tainty of _Q , considering the fact that always the temperaturechange was used for estimating _Q (from which U was obtained di-rectly). The maximum error in the estimation of _Q was found to be3.3%.

The fluids exit into the atmosphere from the heat exchanger.The flowrate of a particular stream of water is equally divided in

Fig. 4. Schematic diagram of the experimental setup.

NH = 9

NC =10

Fig. 5. A schematic diagram of the heat exchanger showing exit ports and the flowdividers used at inlet and exit (blue for cold water and red for hot water). (Forinterpretation of the references to color in this figure legend, the reader is referredto the web version of this article.)

Fig. 6. Temperature change of the fluids (difference of inlet and outlet tempera-tures of the respective streams) (DX = 12 mm).

Fig. 7. Temperature difference of hot and cold water at the exit of the heatexchanger against _VHW.

M. Faizal, M.R. Ahmed / Experimental Thermal and Fluid Science 36 (2012) 242–248 245

all the channels, as shown in Fig. 5. The pipes that carried water toand away from the heat exchanger had its ends equally divided.This is done to achieve similar velocities and pressure of water intheir respective channels. There are a total of nine channels forhot water and ten channels for cold water.

Fig. 8. The average heat transfer between hot and cold water, _QAverage, at varying_VHW.

5. Results and discussion

The results are presented and discussed in this section. Fig. 6shows the change in the temperature of the hot and cold water(i.e. the difference of inlet and outlet temperatures of the respec-tive streams) with varying hot water flowrates, _VHW; forDX = 12 mm. The DTHW decreases with increasing flowrate, andis a minimum at the highest flowrate. The DTHW is a maximumat the lowest flowrate because the hot water gets more time to ex-change heat with the cold water. The DTCW is a maximum at themaximum _VHW because the hot water stream continuously sup-plies heat energy to the cold water stream at a higher rate withoutlosing much heat energy. At higher _VHW; the temperature change ofthe hot water from inlet to outlet is very small. Therefore, the hotwater acts as a continuous heat source to the cold water stream.Similar trends are observed for DX = 6 mm and 9 mm. Fig. 7 showsthe temperature difference between the hot and cold water,DToutlet, measured at the exit of the heat exchanger for all DXvalues. The temperature difference increases slightly and then de-creases as _VHW is increased. The minimum temperature differenceat the exit is obtained at the highest _VHW for all DX values. The inlettemperature difference is 23 �C for all _VH and DX values, and theminimum DToutlet is obtained for DX = 6 mm. Therefore, the opti-mum heat transfer between the two streams is obtained forDX = 6 mm.

The average heat transferred between the two streams is shownin Fig. 8. The heat transfer is calculated as:

_QHW ¼ qHWCPHW_VHWðDTHWÞ ð1Þ

_QCW ¼ qCWCPCW_VCWðDTCWÞ ð2Þ

_QAverage ¼_Q HW þ _QCW

2

!ð3Þ

where _QHW and _QCW are heat transferred by hot and cold waterstreams respectively, _QAverage is the average heat transfer betweenthe two streams. As seen from Fig. 8, _QAverage increases with increas-ing _VHW for all values of DX because of high turbulence at highvelocities, causing a much higher heat transfer. The optimum heattransfer is obtained for DX = 6 mm, because for a given _VHW, the

246 M. Faizal, M.R. Ahmed / Experimental Thermal and Fluid Science 36 (2012) 242–248

hot water velocity will always be higher in the DX = 6 mm channelsbecause of the reduced area. Similar trends for heat transfer (butpresented as Nusselt number) with increasing Reynolds numbersfor corrugated plates are reported in Ref. [14].

The inlet gauge pressures at which the two fluids flowed in theheat exchanger were recorded. Since the flowrate of the cold waterwas kept constant, the pressure variation of the cold water wasvery less for all values of DX, approximately 8–10 kPa. The pressureloss of the hot water varied a lot with _VH, as shown in Fig. 9. Thepressure loss increased with increasing _VHW. The highest pressureloss of 45 kPa is recorded for DX = 6 mm. The minimum pressurelosses are recorded for DX = 12 mm. Similar trends for pressurelosses are reported in Refs. [8,17]. The pressure losses are howeverpresented against the Reynolds numbers in their case. The pressurelosses are due to the promotion of unstable vortices due to the cor-rugations. The increase in pressure losses with increasing Reynoldsnumbers for corrugated plates are also reported in Ref. [14].

The variations of the overall heat transfer coefficient, U, for dif-ferent _VHW and DX are shown in Fig. 10. The U value is calculatedas:

U ¼_Q Average

ADTMð4Þ

DTM ¼ðTHWI � TCWIÞ � ðTHWO � TCWOÞ

ln ðTHWI�TCWIÞðTHWO�TCWOÞ

ð5Þ

Fig. 9. Pressure loss of hot water with varying _VHW.

Fig. 10. The variation of the overall heat transfer coefficient, U, with varying _VHW.

where _QAverage is the arithmetical mean of _QHW and _QCW, A is the to-tal heat transfer area and DTm is the log mean temperature differ-ence. The overall heat transfer coefficient takes into account allthe resistances that are present in the path of the heat transfer.As shown in Fig. 10, U increases with _VHW for all values of DX.The U value is higher for DX = 6 mm because the fluid velocitiesare higher in the 6 mm channels, thus higher turbulence which en-hances heat transfer. A similar trend for the overall heat transfercoefficients against water velocities has been reported in Refs.[7,9]. Sparrow and Comb [18] found that the heat transfer coeffi-cient for the larger plate spacing was slightly smaller than that ofthe lower plate spacing, but the pressure drop was also lower.

The variations of the average thermal length, hAverage, for vary-ing _VHW and DX are shown in Fig. 11. The thermal length repre-sents the performance and is the relationship between thetemperature difference in one stream and the LMTD. A higher ther-mal length means that the heat transfer and the pressure drop arelarge, whereas a lower thermal length means that heat transfer andpressure drops are low [23]. The thermal lengths are calculated as:

hHW ¼DTHW

DTmð6Þ

hCW ¼DTCW

DTmð7Þ

hAverage ¼hHW þ hCW

2ð8Þ

where hHW and hCW are the thermal lengths of the hot water andcold water channels respectively, and hAverage is the arithmeticmean of hHW and hCW. As seen from Fig. 11, hAverage increases with_VHW for all DX values, and is higher for DX = 6 mm compared toother DX values. Therefore, the heat exchanger with DX = 6 mmhas better performance.

The pumping costs of heat exchangers will be higher if the pres-sure losses are significant. The overall heat transfer coefficient, U, isalso considered when designing or choosing heat exchangers.Fig. 12 shows a relationship between the U value and the pressurelosses of the warm water for all values of DX. A similar criterion forthe selection of heat exchangers based on the heat transfer coeffi-cient and the pressure losses is reported in Ref. [10]. There is an in-crease in the U value with increasing pressure loss. For DX = 6 mm,there is a high pressure loss, therefore, the heat exchanger wouldhave higher operational costs. However, the heat exchanger withDX = 6 mm is appropriate because of significant heat transfer coef-ficients and effective heat transfer, even though the pressure lossesare higher. The operational cost could be higher due to high pres-sure losses, but the main objective is to obtain an effective heat

Fig. 11. The variation of the average thermal length, hAverage, with varying _VHW.

Fig. 12. The overall heat transfer coefficient, U, presented against the pressure lossof the hot water, DPH.

M. Faizal, M.R. Ahmed / Experimental Thermal and Fluid Science 36 (2012) 242–248 247

transfer rate between the two streams for such a low temperaturedifference.

Both the hot and cold water streams are single-phase flows thatundergo mainly forced convection and conduction in the heat ex-changer channels. A hydrodynamic and a thermal boundary layerbegin to develop as soon as the fluids enter the channels. The con-vex and concave surfaces in the closed channels cause instabilitiesin the flow, which enhance turbulence. The flow over the convexsurface is more stable because the velocity gradient maintains aconstant sign across the boundary layer [24], and any fluid elementthat gets displaced outward to a higher velocity region gets pushedback to a lower radius region due to higher radial pressure gradient[25]. The flow over the concave surface is unstable because thevelocity gradient changes sign in the boundary layer [24], andany fluid element that gets displaced to a greater radius moves intoa region of low velocities where the pressure gradient is too low topush it back to a lower radius (Görtler instability) [25]. The second-ary flows or the Görtler vortices induced by the corrugations causethe partial restarts of the boundary layer [13], and prevent it frombeing fully developed.

The boundary layer over the convex surface has a point ofinflection which slows down the flow near the surface and changesthe flow direction under a strong adverse pressure gradient. Whenthe incoming flow meets the reversed flow at some point, the fluidnear the surface is transported into the mainstream, or separatedfrom the surface. Since the flow is in a closed channel, and the plategeometries are same, most of the fluid elements are pushed back tothe surface. However, due to initial separation, there are vorticesformed in the wake region and their characteristics depend onthe Reynolds numbers. When on the concave surface, the flow getsunstable due to Görtler instability. As the flow moves forward, itencounters a rising wall (the next convex surface) and as a result,the flow close to the wall slows down and disturbs the incomingflow. As a result, turbulence is enhanced and this continues up tothe end of the channel. Metwally and Manglik [26] performed anumerical study on sinusoidal plate channels and concluded thatflow separation and attachment generates vortices that cause mix-ing which enhances the heat transfer. The corrugations on theplates always cause turbulence in the channels regardless of theflow being laminar or turbulent at the entrance of the channels.Turbulence in the channels leads to wall shear stresses which alsoreduces fouling on the plate surfaces. The heat transfer is a result ofthe disruptions of both the hydrodynamic and the thermal bound-ary layers.

Smooth plates are not so effective because once the hydraulicboundary layer is fully developed, the central region of the fluidsdo not receive much heat from the adjacent channel compared tothe fluid elements close to the wall. Also, as the wall spacing is

increased, the heat received by the central region decreases. Incontrast, corrugations on the plate surface lead to continuous dis-ruptions in the boundary layer across the length of a channel frominlet to exit. The secondary flow causes turbulent mixing of the flu-ids in the channels from one wall to another. This allows almost allthe fluid elements to have effective heat transfer from adjacentchannels. Therefore, it is advisable to always prefer corrugatedplates over smooth plates for plate type heat exchangers.

6. Conclusions

The heat transfer and pressure drops in a corrugated plate heatexchanger with variable spacing and variable warm water flow-rates were studied with the help of temperature measurementsat the inlet and exit of the plate heat exchanger. It is found thatfor a given plate spacing, DX, with increasing hot water flowrate,_VHW; the average heat transfer, QAverage, between the two streamsincreases due to high turbulence at higher velocities. The overallheat transfer coefficient, U, the pressure losses, and the averagethermal length are found to increase with increasing _VHW, andare higher for DX = 6 mm heat exchanger compared to other DXvalues. The plate heat exchanger with DX = 6 mm is found to beappropriate due to effective heat transfer and higher thermallength even though the pressure losses are higher. The corruga-tions on the plate surfaces induce secondary flows in the channelsand cause turbulent mixing which allows all the fluid elements in aparticular channel to have effective heat transfer with the adjacentchannels. The results from this work will be useful for the design ofheat exchangers for small temperature difference applications,irrespective of whether any phase change is involved or not.

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