ceutical industry, chemical industry, renery, etc. These tubetube heat exchangers are successfully demon-strated for superheat recovery water heating applications, condenser and evaporator in heat pumps, lube oil

cooler for shipboard gas turbines, milk chilling and pasteurizing application. This paper presents an exper-imental study on various layouts of TTHE for water-to-water heat transfer. The theoretical and experimen-tal results on this type of heat exchanger conguration could not be located in literature. Overall heattransfer coecient and pumping power were experimentally determined for a xed tube length and surfacearea of serpentine layouts with dierent number of bends and results are compared with straight tubeTTHE. In the case investigated, serpentine layout TTHE with seven bends has shown optimum perfor-mance, with overall heat transfer coecient 17% higher than straight tube TTHE. Two out of ve serpen-tine layout TTHE have shown poor heat transfer performance than straight tube TTHE. The experimentalresults also indicate that there is a denite optimum for a number of bends in serpentine layout TTHE. Ananalytical model for prediction of thermo-hydraulic performance of straight layout has been developed andvalidated experimentally.Water-to-water heat transfer in tubetube heat exchanger:Experimental and analytical study

Milind V. Rane *, Madhukar S. Tandale

Mechanical Engineering Department, Heat Pump Laboratory, Indian Institute of Technology Bombay, Powai,

Mumbai 400 076, India

Abstract

Tubetube heat exchanger (TTHE) is a low cost, vented double wall heat exchanger which increases reli-ability by avoiding mixing of uids exchanging heat. It can be potentially used for heat recovery fromengine cooling circuit, oil cooling, desuperheating in refrigeration and air conditioning, dairy, and pharma-

In this type of heat exchanger, there is a wide choice of congurations to select depending onapplication like liquidliquid, gasliquid, two-phase etc. Fig. 1 shows dierent congurations of

TTHE. Conguration 11 is preferred for high thermal conductivity tube material, 21 forgasliquid application and conguration nn for low thermal conductivity tube material. InKeywords: Double wall heat exchanger; Serpentine layout; Straight tube; Analytical model

1. Introduction

A tubetube heat exchanger (TTHE) is a double wall tubular heat exchanger wherein two ormore tubes are placed side-by-side and bonded thermally using thermal bonding material(TBM) for eective transfer of heat. Use of bends and straight lengths in tubetube heat exchan-ger results in signicant enhancement in heat transfer due to secondary ows induced in thebends. The secondary ows induced in bend leads to heat transfer enhancement in bend as wellas in straight length downstream of bend without signicant increase in pressure drop [1,2].Dean was rst to point out that the occurrence of a secondary ow at right angles to the main

ow is due to centrifugal force [4]. The distorted ow condition by the induced secondary owpersists at a downstream distance of more than 50dti for single-phase and 70dti for two-phase[5]. Chen et al. [6] proposed empirical correlation for U-type wavy tubes with small diameterand short separation between consecutive bends (l/dti = 1.937). Their correlation shows goodagreements with the experimental data. However, extrapolations of correlation with wider oper-ating range needs further examination to check their applicability. Recently, Chen et al. [5] haveproposed a new correlation for friction factor applicable for wider separation between the consec-utive bends (l/dti = 030) but valid for limited range of Re (5010,000).Ohadi et al. [7] have reported their study on eect of bend on pressure drop in a straight section

downstream of a 180 bend. They found about 9% higher pressure drop in downstream sectiondue to bend. Multi-stream Hampson heat exchanger with paired tubes reported by Kao [7] is heli-cal coil type for three uids. The paired tubes are soldered with tinlead solder having thermalconductivity about 10% of copper, which may not be eective in liquidliquid heat exchangedue to its low thermal conductivity.In many heat transfer augmentation techniques, the augmentation is usually accompanied by

signicant increase in the pumping power required to overcome increase in pressure drop forthe same heat transfer rate. But, in case of TTHE with serpentine layout the secondary ows in-duced due to bend continue their eect in the downstream portion of a bend [1,2]. Heat transferenhancement in downstream section is maximum at the leading edge and diminishes along itslength. Experimental and theoretical results on single-phase heat transfer on tubetube heat ex-changer conguration could not be located in literature. This paper experimentally investigatesthe thermo-hydraulic performance of serpentine layout TTHE with dierent number of bends(3, 6, 7, 8 and 9) in water-to-water heat transfer. Also, an experimental comparison of thermo-hydraulic performance of serpentine layout TTHE modules with straight tube TTHE is done.An analytical model for prediction of thermo-hydraulic performance of straight tube TTHE isdeveloped and validated.

2716some of the applications dierent diameter tubes are paired together allowing greater exibility

Nomenclature

A heat transfer surface area, m2

BV ball valveC Cold/heat capacity, J/Kd diameter of tube, mdp pressure drop, N/m2

dbend diameter of bend, mFS full scalef friction factor, dimensionlessH Hoth heat transfer coecient, W/m2 Kk thermal conductivity, W/m Kl length of tube, mln length of tube sector acting as a n, mlpm liters per minutelst.section length of straight section of serpentine tube, mLMTD log mean temperature dierence, degreem mass ow rate, kg/sNTU number of transfer unitsNu Nusselt number, dimensionlessPICCV pressure independent characterized control valvePr Prandtl number, dimensionlessQ heat transfer rate, WRe Reynolds number, dimensionlessR, Rth thermal resistance, K/Wr fouling resistance, m2 K/Wt temperature, CTBM thermal bonding materialU overall heat transfer coecient, W/m2 Kw thickness, mwg gap between adjacent tubes, mwpp.tot total pumping power, W

Subscriptsav averagebend tube bendc coldf foulingh hoti insideo outside

2717

min minimum

2718in optimising the heat exchanger with respect to heat transfer coecient and pressure drop. Theconguration 21 shown in Fig. 1(b) is used in TTHE for recovery of superheat from a 60 TRchiller for one of the hotels in Mumbai [3].

TBM

(b)

H C H

TBM

(a)

TBM

(c)

H H CC C

H C

H

Fig. 1. Sectional view of three TTHE congurations. (a) 11 TTHE, (b) 21 TTHE and (c) nn TTHE.

max maximumN number of variablespt per tubet tubetbm thermal bonding material

Greek symbolse eectiveness, dimensionlessgfs n surface eciency, dimensionless/ semi-ll angle subtended by thermal bonding material at tube centre, degree

Hot water temperature at inlet and outlet of the test section is measured by K type thermo-couples.

The second circuit is cooling water circuit as shown in the upper half of the Fig. 3. This is open

loop with the digital turbine meter for measuring cooling water ow rate and pressure indepen-dent characterized control valve (PICCV) for regulating the ow rate. The water ow rate is ad-justed to dierent values and maintained constant by pressure independent characterized controlvalve. Thermocouples are placed in cooling water stream to measure temperatures at inlet andoutlet of a test section.

2.1. Test section

The test section comprising tubetube heat exchanger module, is shown as a sectional view inFig. 2(a). Various modules of TTHE are fabricated for test. TTHE modules being tested consist oftwo copper tubes of 9.525 mm OD, which are brazed all along their lengths with copper llermaterial. The copper ller, which will be referred in this paper as thermal bonding material(TBM), subtends semi-ll angle /max of 36.7 at the tube centre. The gap between the tubes, wgis measured at dierent sections along the length to calculate the average gap between the tubesSome of the features of TTHE giving advantages over tube-in-tube and shell-and-tube heatexchangers are:

(a) Possibility of using dierent material for tubes carrying uids in order to reduce the cost.Reduction in material used by having two tubes used side-by-side, instead of one insidethe other as the case in tube-in-tube heat exchangers.

(b) For a particular design pressure, outer tube wall thickness is high which increases weight andhence cost of heat exchanger. In TTHE, tubes are placed side-by-side. Hence, diameter oftube replacing the outer tube is small. This reduces the thickness, weight and cost of thetubes. This along with reduced cost of headers leads to cheaper and reliable design in spiteof small additional weight and cost due to TBM.

(c) An ability of TTHE to handle uids with brous materials, partial or complete extraction orintroduction of uids at intermediate temperature, increase or decrease the capacity of heatexchanger by increasing or decreasing number of tube sets, good accessibility of all tubes forease of repair, in case of leakage.

2. Experimental setup

Schematic diagram of the experimental setup for conducting water-to-water heat transferexperiments on TTHE modules is shown in Fig. 3. It includes a test section (TTHE modulecovered with insulation), an electric water heater, valves, pump and instrumentation for measure-ment. The experimental setup consists of two uid circuits. The rst circuit, as shown in Fig. 3,is closed loop for hot water. Hot water is generated in an electric water heater, which has a capa-city of about 9 kW. Flow rate of hot water from water heater is controlled by adjusting valveand temperature of hot water at inlet to the test section is controlled by three-phase Variac.

2719of TTHE module. The simulation results show that for TTHE with copper tubes, a gap of up to

27201 mm has insignicant eect on heat transfer. However, with lower thermal conductivity TBM,the eect of gap will be considerable. The average gap between the tubes is less than 0.5 mmfor all modules tested. This small gap between the tubes allow thermal bonding material to owto other side of tubes during brazing.Hot water ows through one tube and cooling water through other. All modules during the

tests are oriented such that the hot water tube is below the cooling water tube. This eliminates

(a)

(b)Fig. 2. TTHE module. (a) Cross-sectional view of 11 TTHE module; (b) serpentine layout TTHE module with sixbends.

2721the eects, though insignicant, of variants other than number of bends on performance of mod-ules. All modules are insulated thermally from their surroundings during tests by the closed cellfoam insulation (thickness 25.4 mm and thermal conductivity 0.046 W/m K).

2.1.1. Straight layout11 TTHE module (one tube for hot water and another tube for cooling water) fabricated using

two copper tubes with 9.525 mm OD. These straight tubes are brazed all along their lengths withFig. 3. Schematic of experimental setup for water-to-water heat transfer.copper ller.

2.1.2. Serpentine layoutA pair of copper tubes with 9.525 mm OD is rst bent to a serpentine shape on tube bending

machine. The ratio of bend diameter to tube diameter (dbend/dti) is 5.29. These tubes are brazed allalong their lengths. Five modules of serpentine layout with 3, 6, 7, 8 and 9 bends are fabricatedgiving dierent ratio of length of the straight section to tube diameter (lst.section/dti). Fig. 2(b)shows serpentine layout TTHE with six bends.

2.2. Experimental method

The cooling water inlet temperature during the test is 29 0.5 C. The ow rate of water is ad-justed by pressure independent characterized control valve and held constant irrespective of tapwater pressure. The cooling water ow rate is changed from 3 to 10 lpm in step of about1 lpm. Flow rate on two sides is kept same in all sets of readings. The temperature of hot waterinlet to the test section is adjusted to 85 0.1 C by changing heater input through three-phaseVariac.

1.937.9 lpm.

(c) The pressure losses are measured by digital dierential pressure transducers. Make:HBM, Digibar, range: 01.2 bar, resolution: 0.5% of the measurement, and accuracy:

(d) The uid temperatures are measured using K type thermocouples with an accuracy of

Laboratory.The bends in serpentine TTHE modules tested are not perfectly circular in cross-section due tothe manufacturing process used. As a result of this, the tube cross-section became elliptical due toattening. The maximum attening of the bends of all modules tested is less than 8%. The atten-ing is calculated by following equation [12]

flattening d tmax d tmin0.5 d tmax d t.min 1

Since 10% attening results in very small 0.3% reduction in cross-sectional area; 8% attening inthe present case is expected to have insignicant eect on pressure drop.

3. Results and discussion

Pressure drop and heat transfer results for the layouts tested in water-to-water heat transfer areshown in Figs. 4 and 5.

3.1. Pressure drop

Fig. 4 represents variation in experimental pressure drop with mass ow rate on hot water sidefor all six modules of TTHE. In serpentine layouts, the pressure drop increases with number ofbends. The straight layout TTHE has shown lowest pressure drop amongst all layouts tested,

whic0.3 C. The range and resolution of instrument is 20 to 300 C and 0.1 C, respec-tively. Thermocouples are calibrated using thermocouple calibrator in Instrumentation2%. The instrument was calibrated using accurate dial gauge having accuracy of 0.1% ofFS.(b) Flow rate of cooling water below 3.8 lpm is measured by measuring weight of water usingweighing balance with an accuracy of 1% of reading. Make: PAG Oerlikon AG CH-Dieti-kon Precisa Balances, Model No. 87113 type 2300-9535/H6200D. The range and resolutionof instrument is 06.2 kg and 0.1 g respectively. This was done since the accuracy of turbinemeter is lower than 2% for ow rate below 3.8 lpm.2.2.1. InstrumentationInstruments used for measuring various parameters during the experimentation are:

(a) Water ow rate is measured using digital turbine meter. Make: GPI turbine meter, ModelNo. S050 N1/200, Mid-ow, accuracy: 2% of reading in the range 3.837.9 lpm, range:

2722h is as expected.

2 4 6 8 10 12Flow Rate of Water, mh(kg/min)

0

1

2

3

4

5

6

7Pr

essu

re D

rop

on H

ot W

ater

Sid

e, d

p h (ba

r) StraightSerpentine (3 bends)Serpentine (6 bends)Serpentine (7 bends)Serpentine (8 bends)Serpentine (9 bends)

Fig. 4. Variations in pressure drop with mass ow rate of water.

0 8 12 16 20 24

Total Pumping Power, Wpp.tot (W)

1.4

1.6

1.8

2.0

2.2

2.4

2.6

2.8

3.0

3.2

3.4

Ove

rall

Hea

t Tra

nsfe

r Coe

ffici

ent,

Ui.h

(kW

/m2

K) Straight

Serpentine (3 bends)Serpentine (6 bends)Serpentine (7 bends)Serpentine (8 bends)Serpentine (9 bends)

4

Fig. 5. Variations in experimental overall heat transfer coecient with total pumping power.

2723

diminishes long before the uid enters the subsequent bend.In serpentine layouts with eight and nine bends, with d /d of 20.9 and 15.8, respectively, thebend ti

secondary ow generated by the previous bend probably persists as the uid enters in the nextbend. Since the secondary ow is reversed in subsequent bends in a serpentine layout, pressuredrop increases without commensurate increase in heat transfer.Benets of the bends are not realized, since number of bends per unit length is large as in the

case of eight and nine bends, the secondary ow is rst neutralized as it passes the bend and re-versed as it comes out. The consecutive bend interactions are so severe that the performance of theserpentine layout is worse than that of a straight tube TTHE.The quantities measured directly include the volume ow rate of water, inlet and outlet temper-

ature of water. The uncertainty in measurement of volume ow rate of water is 2% of reading(maximum uncertainty: 0.2 lpm). The uncertainty of temperature measurement is 0.3 C.According to the uncertainty analysis based on Kline and McClintock method illustrated by Mof-fat [11], the maximum uncertainties of overall heat transfer coecient, heat transfer rate, and logmean temperature dierence is 3.3%, 3% and 0.76%, respectively.

4. Model development

An eectiveness-NTU method is used in the model developed for performance prediction oftubetube heat exchanger. The present model developed for straight tubetube heat exchangerincorporates following attributes:

(a) Layout: straight.(b) Flow conguration: counter ow or parallel ow.(c) Geometric parameters of tubes: number, diameter, thickness, material on two sides.(d) Thermal bonding of tubes: material and its size (/), gap between the adjacent tubes.

Appropriate parameters like length of n, ln (part of the tube transferring heat by n eect),number of tubes on two sides, semi-ll angle / are included in the model to analyse most of the3.2. Overall heat transfer coecient

Fig. 5 represents variation of experimentally determined overall heat transfer coecient withtotal pumping power for six modules tested. For a xed pumping power, serpentine layout TTHEwith seven bends has shown highest overall heat transfer coecient. Its maximum value is higherthan straight tube TTHE by 17%.When comparison is done amongst serpentine layouts, the overall heat transfer coecient is

highest in serpentine layout with seven bends and lowest with nine bends. For serpentine layoutwith three bends, the overall heat transfer coecient is lower than serpentine layout with sevenbends in this range of ow rate. Thus, the experimental results indicate that in serpentine layoutTTHE, there is a denite optimum for a number of bends for a particular application.In serpentine layout with three and six bends, with dbend/dti of 55.3 and 28.1, respectively, part

of the straight sections does not oer enhancement in heat transfer. Eect of secondary ow

2724congurations (various congurations with dierent number of tubes) of tubetube heat exchan-

the mfor th

Following assumptions were made in the model development.

(a)(b) No heat loss to the surroundings.

(d) Uniform uid distribution in all tubes in multiple tube tubetube heat exchanger.(e)(f)

Thmalcoetant equations used in the model are given below.

2725NTU is calculated by

NTU UACmin

2

Theoretical relations for eectiveness as a function of NTU and heat capacity ratio are availablefor dierent ow arrangements like counter ow or parallel ow. For counter ow, equation foreectiveness is

e 1 expNTU 1 Cratio1 Cratio expNTU 1 Cratio 3

where,

Cratio CminCmaxThe maximum possible heat transfer rate is calculated by

Qmax Cmin thi tci 4

The actual heat transfer rate in heat exchanger is calculated by equatione net heat transfer depends on the waterside heat transfer coecients, fouling factors, ther-resistance by tube walls, and thermal bonding material. The correlations for heat transfercients and friction factors reported in the literature have been used in the model. Few impor-4.1. Heat transfer equationsOne dimensional ow of heat through part of the tube and thermal bonding material.No phase change.(c) No heat transfer in the direction of ow.Steady state operation.ness, thermal bonding material and its size.odel are inlet temperatures and mass ow rates of water. Other input parameters necessarye analysis are ow arrangement, number of tubes, tube material, tube diameter, tube thick-ger. The performance parameters like overall heat transfer coecient, pressure drop, pumpingpower, cost, and size of heat exchanger are calculated for straight layout. Input parameters toQ e Cmin thi tci 5

as w

mental data to an accuracy of 5%.

FoKarm(4

PredHo

used

Th

fh thh thtbm thc f c

2726UA hh Ahgfsh hc Acgfsc

Rthtbm is thermal resistance of thermal bonding material. Thickness of thermal bonding material(measured in the direction of heat transfer) varies from minimum at / = 0 to maximum at thelimiting value /max. Rthh and Rthc are thermal resistance of portion of tubes in contact with ther-mal bonding material on hot water and cooling water side, respectively. Similarly, rfh and rfc arefouling resistances on hot water and cooling water side, respectively. Thermal resistance of ther-

male overall heat transfer coecient is obtained using following expression

1 1 r

1 R R R 1 r

1 91f

p 1.7372 ln Re1.964 lnRe 3.8215

8

This explicit form of PKN correlation agrees within 0.1% of PKN correlation for 104 6Re 6 2.5 108.

5. Overall heat transfer coecientictions of this correlation agree with the extensive experimental measurements within 2 [4].wever, the PKN correlation is not explicit form; Techo, Tickner, James correlation [10] isfor friction factor calculations.1f

p 1.7372 ln Ref

ph i 0.3946 7r friction factor in fully developed turbulent ow for a straight smooth circular tube, Prandtl,an, Nikuradse (PKN) correlation is classical correlation valid for wide range of Re

103 6 Re 6 107). The correlation is also used for comparison of recent correlations.4.3. Friction factorNu 1 12.7 f

2

0.5 Pr2=3 1 6For 2300 6 Re 6 5 106 and 0.5 6 Pr 6 2000.It is modied version of second Petukhov correlation, which agrees with most reliable experi-ell as fully developed turbulent ow for a straight smooth circular tube

f2Re 1000 Pr4.2. Heat transfer coecient

The net heat transfer depends on heat transfer coecients, fouling factors, thermal resistanceby tube walls, and thermal bonding material.Gnielinskis correlation [9] is used for heat transfer coecient calculation for transition regionbonding material is calculated considering dierential element as shown in Fig. 2(a).

Thermal resistance of dierential element is given by

R d to1 cos/ wgktbm dto2 d/ cos/ lpt

10where, R is thermal resistance of the dierential element in terms of dierential angle d/. The ther-mal resistances of all elements of TBM are in parallel. Therefore, the overall thermal resistance ofTBM is obtained by integrating 1/R term for the dierential element between the limits /max to/max. The derived equation for overall thermal resistance of TBM is

Rthtbm 2 Z /max0

ktbm dto2 cos/ lptd to1 cos/ wg d/

111

Another approach to obtain thermal resistance of thermal bonding material is by calculatingequivalent/average thickness considering the area equivalence. Following equation gives equi-valent/average thickness of thermal bonding material.

wtbmav wg sin/max 1 cos/max 12 /max p180 sin/max cos/max

sin/max

d to 12

The thermal resistance of thermal bonding material calculated using Eq. (11) and other calculatedby considering the approach of average thickness of thermal bonding material deviates within

0.5

aD

r

2727Predicted Pressure Drop, dph (bar)

0.0 0.1 0.2 0.3 0.4 0.5 0.6

0.00.1

0.2

Expe

rimen

t0.3

l Pre

ssur

e 0.4

rop,

dp h

(ba0.6

)

45 deg line10%.Fig. 6. Experimental vs. simulated pressure drop.

2.0

sfer

Co

27281.0 1.5 2.0 2.5 3.0

Predicted Overall Heat Transfer Coefficient, Ui.h (kW/m2 K)

1.0

1.5

Expe

rimen

tal O

vera

ll H

eat T

ran

Fig. 7. Experimental vs. predicted overall heat transfer coecient.2.5

3.0ef

ficie

nt, U

i.h (kW

/m2

K)

45 deg line6. Experimental validation of model

The present model developed for straight tube TTHE is validated experimentally for water-to-water heat transfer by testing a typical module of straight tube TTHE. The experimental setupused for conducting experiments is explained in earlier section.The experimental and simulated results on pressure drop and overall heat transfer coecient

for straight tube TTHE are represented in Figs. 6 and 7. The results show good agreement bet-ween the results prediction by model and the experimental results. The average dierence betweenthe predicted and experimental results of pressure drop on hot water side is 2.3% and the maxi-mum dierence is 4.2%. The average dierence between the predicted and experimental overallheat transfer coecient is 2% and maximum dierence is 6.2%.

7. Conclusions

Five dierent serpentine layouts of TTHE have been experimentally evaluated for the eect ofvarying straight lengths between bends on thermo-hydraulic performance. Tubes used in all veserpentine layout TTHE are 9.525 mm OD and 43 mm bend diameter. Maximum overall heattransfer coecient is obtained for the serpentine layout TTHE with seven bends when the perfor-mance is compared at same pumping power. Also, the overall heat transfer coecient in serpen-tine layout TTHE with seven bends is higher than TTHE with straight layout by 17%.

In waterwater heat transfer, not all serpentine TTHE layouts are better than straight layout.The serpentine layout TTHE with three and nine bends have shown lower heat transfer perfor-mance than straight layout TTHE. However, at very low pumping power (below 3 W), serpentinelayout with three bends has shown slightly better heat transfer performance than straight tubeTTHE. The maximum dierence in performance of optimum serpentine layout TTHE (seven

tubes, International Communication Heat Mass Transfer 31 (3) (2004) 303314.[6] I.Y. Chen, S.K. Lai, C.C. Wang, Frictional performance of small diameter U-type wavy tubes, ASME Journal of

2729Fluids Engineering 47 (2003) 22412249.[7] M.M. Ohadi, E.M. Sparrow, A. Walawalkar, A.I. Ansari, Pressure drop characteristics for a turbulent ow in a

straight circular tube situated downstream of a bend, International Journal of Heat Mass Transfer 33 (4) (1990)583591.

[9] V. Gnielinski, New equations for heat and mass transfer in turbulent pipe and channel ow, InternationalChemical Engineering 16 (2) (1976) 359368.

[10] R. Techo, R.R. Tickner, R.E. James, An accurate equation for the computation of the friction factor from smoothpipes from the Reynolds number, ASME Transaction Journal of Applied mechanics 32 (1965) 443.

[11] R.J. Moat, Describing the uncertainties in experimental results, Experimental Thermal and Fluid Science 1 (1988)317.

[12] D.F. Geary, Return bend pressure drop in refrigeration systems, ASHRAE Transactions 2342 (1980) 250265.bends) and nine bend serpentine TTHE, which is a non-optimum, is 30%. Thus, due care shouldbe taken while designing TTHE with serpentine layout to maximize the benets of this design.In case of serpentine layout TTHE, the experimental results indicate that there is a denite opti-

mum for a number of bends for a particular application. In the present case, the optimum numberof bends in serpentine layout is 7.An analytical model for simulation of straight tube TTHE in waterwater heat transfer is val-

idated experimentally. Eectiveness-NTU approach is used in the model for performance predic-tions. The average and maximum deviation between predicted and experimental values of overallheat transfer coecient is 2% and 6.2%, respectively. Similarly, average and maximum deviationis pressure drop 2.3% and 4.2%, respectively.

References

[1] M.V. Rane, M.S. Tandale, Tubetube heat exchangers, Filed PCT/IN03/00377, 2003.[2] M.V. Rane, M.S. Tandale, An experimental study on various layouts of tubetube heat exchanger in steam

condensation, Experimental Thermal and Fluid Science, submitted for publication.[3] M.V. Rane, M.S. Tandale, Benets of superheat recovery on chillers- case study for a hotel installation, Paper

presented at 21st IIR International Congress of Refrigeration, Washington, DC, August 2003, pp. 1722.[4] R.K. Shah, S.D. Joshi, Convective heat transfer in curved ducts, in: S. Kakak, R.K. Shah (Eds.), Handbook of

Single-phase Convective Heat Transfer, rst ed., John Wiley and Sons, 1987.[5] I.Y. Chen, J.C. Huang, C.C. Wang, Singe-phase and two-phase frictional characteristics of small U-tube wavy

Water-to-water heat transfer in tube -- tube heat exchanger: Experimental and analytical studyIntroductionExperimental setupTest sectionStraight layoutSerpentine layout

Experimental methodInstrumentation

Results and discussionPressure dropOverall heat transfer coefficient

Model developmentHeat transfer equationsHeat transfer coefficientFriction factor

Overall heat transfer coefficientExperimental validation of modelConclusionsReferences