International Journal of Heat and Mass Transfer 76 (2014) 523–533

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International Journal of Heat and Mass Transfer

journal homepage: www.elsevier .com/locate / i jhmt

Evaporative heat transfer of R134a and R407C inside a smooth tubewith different inclinations

http://dx.doi.org/10.1016/j.ijheatmasstransfer.2014.04.0560017-9310/� 2014 Elsevier Ltd. All rights reserved.

⇑ Corresponding author. Tel.: +91 844 9498 566; fax: +91 1332 285665.E-mail addresses: akundu261@gmail.com, ak261dme@iitr.ac.in (A. Kundu).

Arijit Kundu ⇑, Ravi Kumar, Akhilesh GuptaMechanical and Industrial Engineering, Indian Institute of Technology Roorkee, 247667, India

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

Article history:Received 10 February 2014Received in revised form 23 April 2014Accepted 23 April 2014Available online 28 May 2014

Keywords:RefrigerantsEvaporative heat transferR134aR407CSmooth tubeInclined

An experimental investigation on two phase flow evaporative heat transfer of refrigerants R134a andR407C in a smooth copper tube inclined at five different angles between 0� and 90� was conducted.The experimental data were obtained over a mass velocity range of 100–300 kg/m2 s, heat flux rangeof 3–10 kW/m2, inlet temperature range of 5–9 �C and vapor quality varied from 0.1 to 0.9. The test sec-tion was 1.2 m long, smooth copper tube with inner diameter of 7.0 mm and outside diameter of9.52 mm. The effects of mass velocity, heat flux, and vapor quality and tube inclinations on evaporativeheat transfer coefficient of both refrigerants are thoroughly compared. The experimental heat transfercoefficients were also compared with some existing correlations.

� 2014 Elsevier Ltd. All rights reserved.

1. Introduction alternative refrigerants evaluation program (AREP) committee

HCFC (hydrochlorofluorocarbon) refrigerant R22 is still one ofthe most used working fluid in refrigeration and air conditioningsystems world-wide, though the use and production of HCFCrefrigerants have been prohibited by the Montreal protocol [1]and UNFCCC [2] regulation due to the high ozone depletion poten-tial (ODP) and relatively high total equivalent warming impact(TEWI). And for this, researches for a suitable replacement havebeen escalating in recent times, but there is no such single compo-nent refrigerant which has a thermodynamic efficiency close toR22 and fulfills international amendment criteria in climatic secu-rity aspects as well. The HFCs (hydrofluorocarbons) are a new fam-ily of substances that might substitute HCFCs. Indeed, they areharmless towards the stratospheric ozone since they do not con-tain chlorine. R134a, a pure HFC, has come forth as a comparablesubstitute to R22 with its excellent thermal performances. Binaryor ternary mixtures are often used in place of pure fluids sincethe required overall properties could be obtained more easily bymixing two or three components. The fact that zeotropic refriger-ants do not boil and condense at constant temperature, the dispar-ity is known as temperature glide, is used to match the pressuredrop in heat exchangers; thereby increasing their efficiency. Thisin turn results in an improved COP of the refrigeration cycle. The

has published an updated list of alternative refrigerants in whichsome refrigerant mixtures came out; R410A, R410B, R407C andR507. R407C is one of the suitable candidates among them toreplace R22 in the appliances working on the near pressureranges [3].

Two-phase flow heat transfers of refrigerant mixtures in heatexchangers with small diameter tubes has become an importantand popular aim of research in recent times because of the demandin the compact heat exchanger design especially in residential andportable refrigeration and air conditioning systems. In addition toreduced refrigeration inventory, improved air-side heat transferperformances can be achieved with smaller tubes, where as a largereduction in tube diameter encountered larger pressure drop. Onthe other hand, flow boiling of mixtures involves convective andnucleate boiling phenomena simultaneously and that makes thediscussion more complex than that for a pure or single componentfluid. It is important to comprehend the boiling heat transfer andflow characteristics of pure fluid and refrigerant mixture in thesmall diameter tubes (lowering the diameter increases the pres-sure drop [4,5]) consisted in compact heat exchangers as evapora-tion is a significant component which inflects the performance of arefrigeration system. There are several researches [6–12] havebeen published comparing the thermodynamic performances ofR22 with R134a, R410A, R410B or R507 and some [13–15] forR407C. However, comparisons in two-phase flow characteristicsin the context of compact evaporator design and ecological aspectsfor R407C with R134a are really scarce.

Nomenclature

AN active nucleation sites (K N m/J/kg–kg/m3)CP specific heat (J/kg K)D diameter (m)G mass velocity (kg/m2 s)hev heat transfer coefficient (W/m2 K)h specific enthalpy (J/kg)I electric current (A)k thermal conductivity (W/m K)L length of tube in flow direction (m)_m total mass flow rate (kg/s)

M molecular weight (dimensionless)P pressure (Pa)Pr Prandtl number (CP l/k, dimensionless)q heat flux (W/m2)Q heat transfer rate (W)T temperature (�C)t time (s)V electric potential (V)x vapor quality (dimensionless)z length (m)

Greek lettersa tube inclination (�)l dynamic viscosity (Pa-s)q density (kg/m3)r surface tension (N/m)D difference

SubscriptsC criticalf saturated liquidg saturated gasIA intermittent to annular flow transitioni insideid ideal valuein inletout outletSat saturationt top sideb bottom sidesl left sidesr right side

524 A. Kundu et al. / International Journal of Heat and Mass Transfer 76 (2014) 523–533

The heat transfer characteristics keep changing as the flow pat-tern changes along an evaporator tube and the flow regime is influ-enced by interfacial shear stress, surface tension and gravitationalforce [12]. In the present study flow boiling heat transfer coeffi-cients of R134a and R407C are measured in a smooth tube inclinedat five different angles between 0� and 90� in the direction ofrefrigerant flow varying heat flux and mass velocity as the flowboiling in inclined tube test data for the aforesaid refrigerants arenot available in the literature. The objective of the present experi-mental study is to develop an accurate flow boiling heat transferdatabase for the new refrigerants, afford data to the refrigerationindustry for the design of high efficient evaporators, compare thethermal performances of R134a with R407Cand investigate theinfluence of tube inclination, heat flux and mass velocity of theflow boiling characteristics of these fluids.

2. Experimental facility and procedure

2.1. Test facility

The experimental plant was designed to investigate two phaseheat transfer phenomena during flow boiling under different flowconditions. The schematic diagram of the experimental facility isshown in Fig. 1. The test arrangement (shown in Fig. 2) consistsof a semi-hermetic compressor, water-cooled condenser, a thermo-static expansion device, pre-evaporator, post-evaporator and test-evaporator. The compressor was connected to a counter flow watercooled shell and tube type condenser. The sub-cooler after the con-denser was included for ensuring liquid refrigerant to enter to theCoriolis-effect mass flow meter. The mass flow rate has been con-trolled through the hand shut off valve fitted on the by-pass lineincorporated after the sub-cooler. A quality filter-dryer was accom-modated after the flow meter to entrap lubricating oil, foreign par-ticles and moisture in the refrigerant. A suitable pre-heater isdesigned and installed to control the vapor quality at the inlet ofthe test evaporator. By regulating power supply with variable volt-age AC, heat source, heat input to the pre-heater and test evapora-tor has been controlled. A suitable accumulator also was installedupstream of the compressor. It was ensured that the refrigerantshould be superheated when it enters the compressor.

The test section was prepared of the smooth copper tube withinner diameter of 7 mm and wall thickness of 1.26 mm. The out-side tube wall temperatures were measured by T-type copper-con-stantan thermocouples at five axial positions, including inlet andexit to the test evaporator tube. At each section, the temperatureswere measured at the top, two sides of the middle and bottom ofthe tube. The average of these temperatures indicates the local walltemperatures. The local saturation pressures at the inlet and outletof test evaporator were measured by piezoelectric pressure trans-ducers. The average of these two pressures indicates the saturationpressure at test section. The test evaporator tube was heated byflexible Nichrome (Nickel–Chromium 80–20% by weight) heaterwire (3.2 kW capacity at 8 m length, with calibrated accuracy of2 W) wrapped over the full test length of 1200 mm. Heat flow tothe heater was conducted using a variety AC voltage controllerand current flow was measured by standard clamp meter to deter-mine applied heat flux. The refrigerant flow patterns wereobserved through the sight glasses installed before and after thetest section. To calculate enthalpy at entry to pre-evaporator,another pressure transducer and T-type thermocouples were fittedat a section just prior to pre-heater. The evaporator tube and pre-evaporator were well insulated with glass wool from outside toensure adiabatic condition. Table 1 summates up the features ofthe plant instrumentation. The temperature and pressure measure-ment were recorded through a data acquisition system to storedata into a personal computer.

2.2. Data reduction

The heat transfer coefficient was calculated by using followingEq. (1):

heV ¼pDiL Two � TSatð Þ

Q� Di

2ln Do=Di½ �

k

� �� ��1

ð1Þ

This is simply based on heat generation by electrical resistanceheating of heater wire outside evaporator tube wall and conductionheat flow through the tube wall to the refrigerant flowing insideat steady state condition. The outside tube wall temperature, Two ,

Fig. 1. Schematic of the experimental layout.

Fig. 2. Photographic view of test arrangement.

A. Kundu et al. / International Journal of Heat and Mass Transfer 76 (2014) 523–533 525

Table 1Measurement equipments and accessories.

Quantity Apparatus Range Accuracy

Temperature T-type thermocouple �50–250 �C

±0.2 �C

Pressure Piezoelectric transducers 0–20 bar ±0.5% F.S.Mass flow rate Coriolis effect mass flow meter 0–200 kg/h ±0.2%Voltage AC variable voltage controller 0–220 V ±0.5%Current Clamp meter 0–100 A ±1.0% F.S.

±5digits

Table 2Uncertainty of variables.

Primary measurements Derived quantities

Parameter Uncertainty Parameter Uncertainty (%)

Diameter �2 lm Reynolds number �0:6Pressure �0:5% Mass velocity �0:6� 2Temperature �0:05 �C Heat transfer coefficient �3:9� 11Mass flow rate �0:2% Vapor quality �2:0� 9:5Heat flux �2:1� 3:6% Electrical power �0:80

526 A. Kundu et al. / International Journal of Heat and Mass Transfer 76 (2014) 523–533

is the mean of outside wall temperature at the top, two sides ofthe middle and bottom of the tube at each section in the testevaporator as:

Two ¼TtðzÞ þ TslðzÞ þ TsrðzÞ þ TbðzÞ

4ð2Þ

Tsat is the saturation temperature corresponding to the pressurereadings, taking the mean of the dew and bubble point tempera-ture for refrigerant mixtures. Accounting the heat loss to the ambi-ent by the heating coefficient, g defined as the ratio between theoutput power and input power, the resistance heat flow Q, to therefrigerant from outside of the tube (pre-heater and test section)has been calculated by using following Eq. (3):

Q ¼ g � V � I ð3Þ

g (typically around 0.935) is experimentally determined usingthe method proposed by Wambsganss et al. [16]. Average vaporquality is calculated by following Eq. (4):

xavg ¼ ðxin þ xoutÞ=2 ð4Þ

where, xin is the inlet dryness fraction of refrigerant to the test sec-tion, which is calculated by following Eq. (5):

xin ¼hin � hf ;in

hfg;inð5Þ

hin is the specific enthalpy of the refrigerant at the entry to thetest section and outlet of the pre-heater, which is determined byEq. (6) applying an energy balance on the pre-heater:

hin ¼ h1 þ Q PH= _mð6Þ

Enthalpy h1 is found with respect to the measured temperature andpressure at the entry to the pre-heater and hfg is the latent heat ofvaporization of the refrigerant. xout is the exit dryness fraction ofrefrigerant from test section can be determined from Eq. (7):

xout ¼hout � hf ;out

hfg;outð7Þ

hout is the exit enthalpy of the refrigerant is determined byapplying an energy balance on the test section as:

hout ¼ hin þ Q TS= _mð8Þ

Table 3Maximum uncertainty in heat transfer coefficient.

G (kg/m2 s) Q (W/m2) Uncertainty (%)

100 10,000 ±11200 6000 ±9300 3000 ±8.5

2.3. Uncertainty analysis

The error in measurement depends on the operating conditionsand mostly on the accuracy of the wall temperature difference. Theexperimental uncertainty analysis was done by following Eq. (5)proposed by Schultz and Cole [17]:

UR ¼

ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiXN

i¼1

@R@Vi

UVi

� �2vuut ð9Þ

where UR is the estimated uncertainty in calculating the value ofdesired variable R, due to the independent uncertainty UVi

in theprimary measurement of N number of variables Vi, affecting the

result. The experimental uncertainties for the sensors are listed inTable 2. Maximum uncertainty in the measurement of heat transfercoefficient is shown in Table 3.

2.4. Refrigerant property

In the present study R134a and R47C have been tested. R134a isa pure fluid, where R407C is non-azeotropic ternary blend of R32,R125 and R134a (23%, 25% and 52% by weight, respectively).Table 4 lists some of the most important physical properties ofthe tested refrigerants at different test pressures. The thermody-namic and transport properties of refrigerants were obtained fromREFPROP 8.0 [18]. The ranges of experimental parameters are listedin Table 5.

3. Results and discussion

In the present paper the heat transfer coefficients are measuredfor two refrigerants, pure fluid R134a and zeotropic blend R407C,by varying the heat flux with different evaporating pressures andmass velocities while maintaining the refrigerant inlet tempera-ture to the test evaporator between 5 and 9 �C. The experimentalconditions are summarized in Table 6.

3.1. Validation

In order to establish the integrity of the experimental setup, theheat transfer results of the flow boiling of R134a through the testevaporator tube in horizontal condition are compared with thepredicted values by Liu and Winterton [19], Gungor and Winterton[20] and Kandlikar [21] correlations as shown in Fig. 3. This con-firms that the experimental results on two-phase flow and flowevaporation using the test facility and measurement system arereliable. The experimental result shows the comparison of ±30%errors with a mean deviation ranging from +2% to �29%. Amongthem, the mean deviation is the smallest by using Gungor andWinterton’s correlation [20].

3.2. Flow pattern map

Since, the heat transfer process depends upon the flow pattern;one must distinguish the flow pattern during flow boiling experi-ments as shown in Fig. 4 for R134a and Fig. 5 for R407C for hori-zontal flow boiling in the present study. A number of flowpattern maps have been proposed over the years for predictingtwo-phase flow regime transitions in horizontal tubes. In this

Table 4Properties of test fluids.

Fluid PSat (bar) TSat (�C) hfg qf kf lf CPf ANa PC (bar) M(kJ/kg) (kg/m3) (W/m K) (lPa-s) (kJ/kg K)

R134a 3.61 5.92 194.02 1274.9 0.0894 247.22 1.3579 1.744e�06 40.593 102.033.79 7.34 192.88 1270.1 0.0887 242.84 1.3621 1.650e�064.01 9.00 191.55 1264.4 0.0880 237.81 1.3672 1.546e�06

R407C 6.22 2.83/8.88 209.61 1225.8 0.0948 203.42 1.4266 1.265e�06 46.34 86.2046.47 4.07/10.10 208.37 1221.2 0.0942 200.29 1.4313 0.980e�066.86 5.94/11.92 206.4 1214.2 0.0933 195.68 1.4386 0.912e�06

a Unit: K N m/J/kg–kg/m3.

Table 5Range of experimental parameters.

Parameters Range

Refrigerant mass velocity 100–300 kg/m2 sHeat flux 3000–10,000 W/m2

Vapor quality 0.1–0.9Evaporative pressure 3.61–4.01 bar (R134a); 6.17–6.86 bar (R407C)Inlet temperature 5–9 �C

Table 6Operating conditions.

Refrigerant G (kg/m2 s) Pev (bar) q (kW/m2) Dx

R407C 100.7 6.18 2.96 0.81–0.13599.8 6.23 4.51 0.831–0.129

101.1 6.40 6.01 0.859–0.138100.2 6.61 8.23 0.839–0.123

99.7 6.79 9.87 0.85–0.141198.2 6.17 3.11 0.866–0.134200.4 6.24 4.56 0.874–0.128199.6 6.47 5.89 0.881–0.127201.1 6.61 8.11 0.893–0.148200.8 6.76 10.12 0.874–0.135299.6 6.17 3.01 0.879–0.131302.2 6.31 4.46 0.881–0.129300.8 6.59 6.20 0.888–0.130296.8 6.72 7.88 0.883–0.136301.9 6.86 10.20 0.892–0.129

R134a 101.1 3.61 3.06 0.818–0.133100.8 3.69 4.48 0.841–0.139

98.9 3.76 6.10 0.867–0.128100.2 3.87 7.78 0.859–0.143102.3 4.01 10.07 0.897–0.134199.3 3.67 2.97 0.867–0.129202.1 3.74 4.60 0.891–0.138198.5 3.77 5.99 0.89–0.131201.8 3.81 8.01 0.863–0.145197.8 3.98 9.72 0.888–0.128300.6 3.59 3.21 0.906–0.149301.0 3.68 4.50 0.876–0.136298.5 3.79 6.08 0.846–0.136301.8 3.92 8.13 0.900–0.150299.9 4.00 10.17 0.898–0.128

Fig. 3. Comparison of test results with existing correlations.

Fig. 4. Flow pattern map for horizontal evaporator tube with R134a in the presentstudy.

A. Kundu et al. / International Journal of Heat and Mass Transfer 76 (2014) 523–533 527

paper, the latest version of the Kattan–Thome–Favrat map [22] andthe Wojtan–Ursenbacher–Thome map [23] has been used. Theflow patterns observed in the experiment are stratified-wavy flow,intermittent flow, semi-annular flow and annular flow. The flowpatterns are plotted in the coordinates of G and x for the evaporatortube. The transitions of flow pattern would be stratified-wavy,intermittent, semi-annular, and annular in the direction of increas-ing G and x. The transition lines are shown with continuous lineover the calculated points (calculated from their underlying transi-tion equations, as described in Appendix A) shown; where dashedline denotes the corresponding boundary from the Wojtan–Ursenbacher–Thome map [23]. xIA line shows the calculated flowtransition from intermittent to annular flow. The flow regimes’

identification criteria are the same as used in Mastrullo et al.[24]. The dry-out region is recognized by the sharp drop in the localheat transfer coefficient that occurs at very high vapor qualitiesafter it reaches a maximum in the annular flow regime forR134a; the same occurs earlier for R407C as can be seen fromFig. 5.

3.3. Effect of vapor quality and heat flux on heat transfer coefficient

In Figs. 6 and 7, the local boiling heat transfer coefficients arereported for R134a and R407C as a function of vapor qualityobtained by varying the mass velocity and the heat flux at

Fig. 5. Flow pattern map for horizontal evaporator tube with R407C in the presentstudy.

528 A. Kundu et al. / International Journal of Heat and Mass Transfer 76 (2014) 523–533

horizontal condition of the evaporator tube. In the experimentaltests, the heat transfer coefficient increases with increasing heatflux and vapor quality with fixed mass velocity of 100 kg/m2 sand a different dependence of the heat transfer coefficients onvapor quality is apparent as seen from Fig. 6. Indeed, at low valuesof the vapor quality, the heat transfer coefficient increases withincreasing vapor quality presenting a local maximum in the vaporquality range between 30% and 40% for R407C and 65–75% forR134a. The same trend was observed by Wang et al. [13] and Wangand Chiang [25].

The heat transfer progression in flow evaporation results fromthe interaction between nucleate boiling and liquid convection.The relative consequence of the two different mechanisms varieswith vapor quality and strappingly depends on flow conditions.At high heat fluxes and evaporating pressures, heat transfer ismostly prejudiced by the nucleate boiling mechanism [13]. Theconvective contribution to heat transfer prevails at low heat fluxesand evaporating pressures [26]. In the evaporation process whereliquid convection is the main mechanism, convective evaporationbeing predominant does not implicate that nucleate boiling isentirely inhibited. Under these conditions, the heat transfer coeffi-cient increases with vapor quality as explained by Greco and Van-oli [8]. The present study also provides the same occurrence asshown in Fig. 6; the increment is only 15–23% for R407C; but forR134a, it is quite high as up to 65–110% more with respect tothe heat transfer coefficient at very low vapor qualities in the range

Fig. 6. Comparison of heat transfer coefficient with horizontal tub

of 0.1–0.2 depending on heat fluxes. In order to explain the dispo-sitions of the R407C and R134a heat transfer coefficients at lowvapor qualities, it is necessary to take into account of the equilib-rium bubble radius r⁄ which depends on the refrigerant thermo-physical properties [26,27]. Bubbles that are smaller in radius thanr⁄ will collapse in a spontaneous manner, and bubbles that are big-ger in radius than r⁄ will maturate. Hence, a lower r⁄ implies a lar-ger number of active nucleation (AN) sites on the heated surfacefor bubble formation and, therefore, a stiffer nucleate boiling con-tribution. The equilibrium bubble radius is directly proportional tothe fluid property combination [27] as shown in Eq. (6):

AN ¼ 2TSatrf

hfgqgð10Þ

Here TSat is the bubble point temperature of refrigerant on cor-responding evaporating pressure expressed in Kelvin. It can beobserved that for the same operating conditions, the R407C prop-erty combination AN, is lower than that of R134a. As a conse-quence, for the same wall superheat, the nucleate boilingcontribution of R134a is larger than that of R407C. Besides, dueto a prominent temperature glide and variation of the propertiesof the different constituents inside the zeotropic refrigerant mix-ture R407C (mixture of R32, R125 and R134a) make the propaga-tion of the boiling phenomena different from that of the singlecomponent refrigerant R134a. Wang and Chiang [25] describedthe phenomena as at the lower quality region (early stage of vapor-ization), more volatile components of R407C, R32 and R125 evap-orated faster than the least volatile ingredient R134a as qg for R32and R125 (ql of R125 is nearly same to R134) are too high withlower ql for R32 than those of R134a at the corresponding evapo-rative pressures and the higher mean vapor phase density causeslower mean vapor phase velocity for R407C. Therefore, the lowermean vapor and liquid phase velocity for R407C than R134a maycause a sharp delay in flow pattern transition (can be seen for xIA

fro Figs. 4 and 5) and thus, too lower heat transfer coefficient forR407C than that of R134a till the refrigerant leaves the evaporatortube.

As the flow proceeds downstream and vaporization takes place,the void fraction increases, thus decreasing the density of theliquid–vapor mixture because of lower density of the vapor. As aconsequence, the flow accelerates intensifying convective trans-port from the heated wall of the tube. The increment in heat trans-fer coefficient proceeds until the liquid film disappears, leaving thetube wall partially or totally dry. In this region, the heat-transfercoefficient drops sharply because of the low thermal conductivityof the vapor. The same can be observed from the figures that after

e at mass velocity of G = 100 kg/m2 s for different heat fluxes.

Fig. 7. Comparison of heat transfer coefficient with horizontal tube at mass velocity of G = 200 kg/m2 s and G = 300 kg/m2 s for different heat fluxes.

A. Kundu et al. / International Journal of Heat and Mass Transfer 76 (2014) 523–533 529

vapor quality of x = 0.75–0.80 for R134a, the heat transfer coeffi-cient drops about 22–27% than that of the maximum one. ForR407C, the same occurs earlier, near about 30–40% of the tube.The cause may be the departure from nucleate boiling because ofthe local void spreads as a vapor blanket on the heating surfacefor the lower vapor velocity attained during the proceeding ofthe evaporation; and the bubble crowding and vapor blanketingdeflowers the surface cooling by reducing the incoming liquid.

3.4. Effect of mass velocity on heat transfer coefficient

The effect of mass velocity at a fixed heat flux of 3, 4.5, 6, 8 and10 kW/m2 with the horizontal position of the test evaporator canbe observed comparing the test data from the Figs. 6 and 7. Dueto mass transfer resistance, nucleate boiling heat transfer coeffi-cient for zeotropic mixtures are considerably lower than purerefrigerants. At low mass velocities, major contribution to the heattransfer mechanism is the nucleate boiling. As the mass velocityincreases, the mass transfer resistance to the convective boilingis reduced by the contribution of the more volatile heat transfercoefficient and a higher mass velocity induces liquid entrainmentto agitate the mass transfer resistance. The same phenomenonwas described by Kattan et al. [22]. At the initial stage of the boil-ing, due to larger contribution of the nucleate boiling part, the heat

Fig. 8. Comparison of the effect of tube inclination a

transfer coefficient increases with increase in mass velocity forboth the refrigerants; but for R407C, as discussed previous, dueto larger void spreading and vapor blanketing over the heated sur-face, the premature departure from nucleate boiling diminishes theincrement in heat transfer and thus the effect of mass velocityincrement is not as severe as the case for R134a, which can be seenfrom Fig. 7.

3.5. Effect of tube inclination on heat transfer coefficient

Fig. 8 shows the effect of tube inclination angles on the heattransfer coefficient of both refrigerant R407C and R134a at fixedheat flux of 4.5 kW/m2. The mass velocity of refrigerant has main-tained constant at G = 100 kg/m2 s. For all tube inclinations, theheat transfer coefficient increases up to vapor quality x = 0.30–0.35 for R407C and for R134a, it increases up to x = 0.65–0.75and then decreases because of dry out. Fig. 9 shows the variationof the local heat transfer coefficient for different tube inclinationsat mass velocity of 200 kg/m2 s and fixed heat flux of 4.5 kW/m2.

In the present investigation, at low mass velocity and lowimposed heat flux inside the evaporator tube, heat transfer coeffi-cient increases with vapor quality up to 30–35% of the tube forR407C. This nature of increment in heat transfer coefficient pre-vails same for all tube inclinations. When mass velocity increases,

t G = 100 kg/m2 s with constant q = 4.5 kW/m2.

Fig. 9. Comparison of the effect of tube inclination at G = 300 kg/m2 s with constant q = 4.5 kW/m2.

530 A. Kundu et al. / International Journal of Heat and Mass Transfer 76 (2014) 523–533

the trend in increment continues to vapor quality of 56–65% of thetube for R134a. After that, heat transfer coefficient decreases aslong as the fluid leaves the evaporator tube. A comprehensiveexploration of convective boiling and nucleate boiling heat transfercoefficient component separately reveals that as evaporation pro-ceeds through the length of the tube and vapor quality increases,due to mass transfer resistance, only convective part increasesunhurriedly for zeotropic mixture R407C. The mass transfer resis-tance to convective evaporation in R407C at high mass velocitiesused to be reduced significantly due to the disturbances at themass transfer resistance boundary created by the highly turbulentinterface between liquid and vapor interface and that causes agreat decrease in average heat transfer coefficient for R407C thanthat of the pure single component fluid R134a. The effect of tubeinclination on the heat transfer coefficient for mass velocityG = 300 kg/m2 s and constant heat flux q = 10 kW/m2 has beencompared for the two test refrigerants in Fig. 9.

Mass velocity and tube inclination affect the vapor quality atwhich boiling crisis occurs [12]. The same can be observed fromthe present study as shown in Fig. 10. The quality, at which theheat transfer coefficient reached maximum, increases with theincrease in mass velocity and decreases with an increase in heatflux for R407C as can be seen from Fig. 11; but for R134a, maxi-mum heat transfer coefficient increases with both increase in massvelocity and heat flux for almost all tube inclinations. The vaporquality corresponding to the maximum heat transfer coefficientis worst for 90� tube inclination for both the fluid with low mass

Fig. 10. Comparison of the effect of tube inclination at const

velocity and high heat flux. For mass velocity G = 100 kg/m2 s withR407C, dry out occurs at vapor quality about 34–36% of the evap-orator tube with tube inclinations 0–30� for constant heat flux of6 kW/m2. As the tube inclination increases, dry-out occurs in only45–48% of vapor quality at a mass velocity of 200 kg/m2 s. Dry-outoccurs at x = 62–84% for the tube when mass velocity is 300 kg/m2 s for R134a for tube inclination is 90�. It increases to x = 0.72–0.86 for R134a with horizontal evaporator tube. From Fig. 12, itis obvious that the tube inclination influences the average heattransfer coefficient significantly. The figure shows the effect of tubeinclination angle on the average heat transfer coefficient for fixedheat flux q = 6 kW/m2 with mass velocity varying from 100 to300 kg/m2 s. It reveals that for all mass velocities, the local heattransfer coefficient is best at 90� of tube inclination for both therefrigerant; where after dry-out, 90� inclination is the worst as alsocan be revealed from Figs. 8–10.

The average heat transfer coefficient for R407C increases to 12–28% for tube inclination increases from of 30� to 90� with respect tothe horizontal position of the evaporator tube at a low mass veloc-ity of 100 kg/m2 s with heat flux of 6 kW/m2, which can beobserved from Fig. 12. But the average heat transfer coefficientincreases to 3–14% of tube inclination increases from 30� to 90�with respect to the horizontal position of the evaporator tube, mostof which lies between the experimental uncertainty limit, whenmass velocity increases to 200 kg/m2 s. Before dry out, the averageheat transfer coefficient increases about 10–34% of tube inclinationincreases from 30� to 90� with respect to horizontal tube. The

ant G = 300 kg/m2 s with q = 3 kW/m2 and q = 10 kW/m2.

Fig. 11. Comparison of the effect of tube inclination on vapor quality for the maximum heat transfer coefficient at different mass velocities and heat fluxes of R407C andR134a.

Fig. 12. Comparison of the effect of tube inclination on average heat transfer coefficient with fixed heat flux of 6 kW/m2 at different mass velocities for R407C and R134a.

A. Kundu et al. / International Journal of Heat and Mass Transfer 76 (2014) 523–533 531

effect of tube inclination goes up with increment in mass velocitywithin intermittent flow, but descends as soon as annular flowregime prevails. At G = 300 kg/m2 s, average heat transfer coeffi-cient increases with increase in tube inclination from 30� to 90�only between 3% and 15% for R407C as can be seen from Fig. 12.The cause may be the lower buoyancy of the vapor bubbles accu-mulated prematurely by the effect of early dry-out of refrigerantmixture R407C and causes less liquid to flow faster; thus createsa delay in flow pattern transition at near the beginning stage ofannular flow or semi-annular flow. In comparison to R407C, thehorizontal position of the evaporator tube for R134a, the averageheat transfer coefficient in inclined tube increases only 4% and9% for mass velocity of 300 kg/m2 s for varying tube inclinationfrom 30� to 90� which were also within the uncertainty for exper-imental results. Vertical flow of both the refrigerants has the worstheat transfer coefficient for all mass velocities irrespective of heatflux after dry out. For mass velocity G = 300 kg/m2 s with heat fluxapplied of 6 kW/m2, the average heat transfer coefficient for verti-cal up flow boiling of R134a is about four times more than that forR407C, where at 0� and 45�, it is about 325% and 300%, respec-tively. For mass velocity G = 100 kg/m2 s, the average heat transfercoefficient for horizontal flow evaporation of R134a with the heatflux applied of 6 kW/m2 is about 234% more than that for R407C;where at 60�, it is about three times. Average heat transfer coeffi-cient, in general, is defined as the mean of local heat transfer coef-

ficients at different vapor qualities for a given imposed heat fluxand constant mass velocity.

The comparison of the flow boiling performance inside theevaporator tube inclined at 45� with different mass velocities and8 kW/m2 heat flux has been drawn in Fig. 13. At this inclination,heat transfer coefficient increases with mass velocity, but the dif-ference in increment decreases with respect to those in the hori-zontal position of the tube, if we compare those results as shownin Figs. 7 and 8. The vapor phase velocity has been reduced forlow vapor quality resulting in a corresponding decrease in inertiaforce. In low vapor quality region, thus, the gravitational forceaffects the two-phase flow in a considerable manner and thiscauses the variation of heat transfer coefficient for different tubeinclinations. Also the variation of heat transfer coefficient in vaporquality for a particular tube inclination has uneven behavior due tothe transition of different flow pattern through the growth of theboiling along the tube length. From Figs. 10, 11 and 13, it is evidentthat for high heat flux, as mass velocity increases, inclination inevaporator tube demeans the boiling performance of refrigerants.Average heat transfer coefficient increases 7% with R407C and43% with R134a for an increase in mass velocity from 100 to300 kg/m2 s at an evaporator tube inclination of 45�; and 6% and41% increase for inclination of 60� with heat flux of 8 kW/m2 forR407C and R134a, respectively. For horizontal tube, the incrementis 4% for R407C and for R134a, it is 47% as the mass velocity

Fig. 13. Comparison of the effect of mass velocity on heat transfer coefficient with fixed heat flux of 8 kW/m2 at tube inclination of 45�.

532 A. Kundu et al. / International Journal of Heat and Mass Transfer 76 (2014) 523–533

increases the same. The cause may be the buoyancy effect of vapor,which accelerates the upstream flow at a higher inclination otherthan horizontal would augment the convective contribution ratherthan nucleate boiling part and the convective boiling coefficientarouses a gentle but nucleate boiling being suppressed more withincrease in vapor quality as vaporization proceeds. This phenome-non also has been demonstrated by Kattan et al. [22] by describingthe lower heat transfer coefficient for downward flow with respectto upward and horizontal flow of pure refrigerant.

However, the interfacial shear stress has been increased by theturbulence created by the very high mass velocity and thus, thebubbles formed on the heated surface of the tube could not attainat a large size as seen in nucleate boiling. By increasing the massvelocity, the influence of nucleate boiling, which is the dominantpart of the low vapor quality region became less effective to con-tribute in flow boiling of refrigerant upstream. And for this, inthe current study for R407C and R134a, it is revealed that the aver-age heat transfer coefficient increases 8% and 18%, respectively, foran increase in mass velocity from 200 to 300 kg/m2 s i.e. transitionfrom intermittent and semi-annular region to annular flow withq = 10 kW/m2 but also an eminent decrease in local heat transfercoefficient occurs from inlet to outlet of the evaporator tube by10%, 13% and 21% for R407C at mass velocities of 100, 200 and300 kg/m2 s, respectively. But as the nucleate boiling contributionto flow evaporation of pure fluid R134a is much more, as describedby Eq. (10) and shown in Table 4, it is seen that the heat transfercoefficient much more increases at the outlet of the evaporatortube with respect to that at the entry to the evaporator; but consid-erable amount decrease of 22–24% of that after a dry-out of the exitof the evaporator tube.

4. Conclusion

1. An experimental plant has been established in the refrigerationlaboratory of Indian Institute of Technology Roorkee, India forevaluating the heat transfer performance of pure and mixedrefrigerants during convective boiling. Local heat transfer coef-ficients of pure fluid R134a and zeotropic mixture R407C, inflow evaporation through an inclined tube with five inclinationangles from 0� to 90� in the direction of refrigerant flow (incli-nation increased upward) were measured at the evaporatingpressure range of 3.61–4.01 for R134a and 6.22–6.86 bar forR407C with the temperature at the entry to the test evaporator

was maintained the same for both the fluid between 5 and 9 �Cdepending on evaporative pressures, the heat flux from 3 to10 kW/m2 and the mass velocity from 100 to 300 kg/m2 s. Theeffect of mass velocity, heat flux and vapor quality; and tubeinclinations on evaporative heat transfer coefficient has beeninvestigated for different flow regimes.

2. The experimental results indicate that the heat transfer coeffi-cients increase with mass velocity and heat flux. Also the localheat transfer coefficient increases for both the refrigerantsbefore the dry-out occurs. The heat transfer coefficientincreases with increasing vapor quality presenting a local max-imum in the vapor quality range between 30% and 40% forR407C and 65–75% for R134a.

3. The nature of the increment in heat transfer coefficient withrespect to mass velocity, heat flux and vapor quality varies withthe flow patterns for different refrigerants considered for thetest. The increase in heat transfer coefficient was at much lowerquality for R407C than R134a mainly due to the nature of thecontribution of nucleate boiling to the evaporation progressionof the refrigerants.

4. The effect of tube inclination was also much more severe on theperformance of R134a than R407C. The heat transfer coeffi-cients of pure fluid R134a are higher than that of refrigerantblend R407C for all mass velocities, heat fluxes and tube inclina-tions. For mass velocity G = 300 kg/m2 s with heat flux appliedof 6 kW/m2, the average heat transfer coefficient for verticalup flow boiling of R134a is about four times more than thatfor R407C, where at 0� and 45�, it is about 325% and 300%,respectively. For mass velocity G = 100 kg/m2 s, the averageheat transfer coefficient for horizontal flow evaporation ofR134a with the heat flux applied of 6 kW/m2 is about 234%more than that for R407C; where at 60�, it is about three times.

Conflict of interest

None declared.

Appendix A

A.1. Fluid and geometry definition

Input: Di, xavg, Tsat, G, qPhysical parameters from REFPROP: qf, qg, lf, lg, r, kf, kg

A. Kundu et al. / International Journal of Heat and Mass Transfer 76 (2014) 523–533 533

A.2. Flow pattern boundaries calculation

Stratified to Stratified Wavy transition is calculated from:

GStratified ¼ð226:3Þ2AfdA2

gdqgðqf � qgÞlf gx2ð1� xÞp3

( )1=3

ðA1Þ

Stratified wavy to intermittent/annular boundary is calculatedfrom:

GWavy ¼16A3

gdgDiqf qg

x2p2ð1� ð2Hfd � 1Þ2Þ0:5

p2

25H2fd

ð1� xÞ�F1ðqÞ WeFr

� ��F2ðqÞ

fþ 1

" #8<:

9=;

1=2

þ 50

ðA2Þ

where F1ðqÞ¼646:0ðq=qCÞ2þ64:8ðq=qCÞ; F2ðqÞ¼18:8ðq=qCÞ þ1:023

and qC ¼ 0:16q1=2g hfg grðqf � qgÞ

h i1=4.

Intermittent to annular flow transition is calculated from:

xIA ¼ 0:2914 qg=qf

� ��1=1:75lf =lg

� ��1=7� �

þ 1 �1

ðA3Þ

Annular to dry-out boundary is calculated from:

GDry-out ¼1

0:235 ln 0:58x

� �þ 0:52

� Dirqg

� ��0:17�

1gDiqg qf�qgð Þ

� ��0:37

� qg

qf

� ��0:25� q

qC

� ��0:70

8>><>>:

9>>=>>;

0:926

ðA 4Þ

Dry-out to Mist is calculated from:

GMist ¼1

0:0058 ln 0:61x

� �þ 0:57

� Dirqg

� ��0:38�

1gDiqg qf�qgð Þ

� ��0:15

� qg

qf

� ��0:09� q

qC

� ��0:27

8>><>>:

9>>=>>;

0:943

ðA 5Þ

A.3. Flow pattern classification

Stratified flowpattern

G < Gstrat

Stratified-wavyflow pattern

G > Gwavy(xIA) gives the slug zone

Gstrat < G < Gwavy(xIA) and x < xIA give theslug/stratified-wavy zone

x P xIA gives the stratified-wavy zone

Intermittent flowpattern

Gwavy < G < Gdry-out and x < xIA

Annular flowpattern

Gwavy < G < Gdry-out and x > xIA

Dry-out flowpattern

If Gstrat P Gdry-out, then Gdry-out = Gstrat

If Gwavy P Gdry-out, then Gdry-out = Gwavy

Mist flow pattern

If Gstrat P Gmist, then Gmist = Gstrat

If Gwavy P Gmist, then Gmist = Gwavy

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