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

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

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Evaporative heat transfer of R134a and R407C inside a smooth tube with different inclinations Arijit Kundu ⇑, Ravi Kumar, Akhilesh Gupta Mechanical and Industrial Engineering, Indian Institute of Technology Roorkee, 247667, India

a r t i c l e

i n f o

Article history: Received 10 February 2014 Received in revised form 23 April 2014 Accepted 23 April 2014 Available online 28 May 2014 Keywords: Refrigerants Evaporative heat transfer R134a R407C Smooth tube Inclined

a b s t r a c t An experimental investigation on two phase flow evaporative heat transfer of refrigerants R134a and R407C 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 range of 3–10 kW/m2, inlet temperature range of 5–9 °C and vapor quality varied from 0.1 to 0.9. The test section was 1.2 m long, smooth copper tube with inner diameter of 7.0 mm and outside diameter of 9.52 mm. The effects of mass velocity, heat flux, and vapor quality and tube inclinations on evaporative heat transfer coefficient of both refrigerants are thoroughly compared. The experimental heat transfer coefficients were also compared with some existing correlations. Ó 2014 Elsevier Ltd. All rights reserved.

1. Introduction HCFC (hydrochlorofluorocarbon) refrigerant R22 is still one of the most used working fluid in refrigeration and air conditioning systems world-wide, though the use and production of HCFC refrigerants have been prohibited by the Montreal protocol [1] and UNFCCC [2] regulation due to the high ozone depletion potential (ODP) and relatively high total equivalent warming impact (TEWI). And for this, researches for a suitable replacement have been escalating in recent times, but there is no such single component refrigerant which has a thermodynamic efficiency close to R22 and fulfills international amendment criteria in climatic security aspects as well. The HFCs (hydrofluorocarbons) are a new family of substances that might substitute HCFCs. Indeed, they are harmless towards the stratospheric ozone since they do not contain chlorine. R134a, a pure HFC, has come forth as a comparable substitute to R22 with its excellent thermal performances. Binary or ternary mixtures are often used in place of pure fluids since the required overall properties could be obtained more easily by mixing two or three components. The fact that zeotropic refrigerants do not boil and condense at constant temperature, the disparity is known as temperature glide, is used to match the pressure drop in heat exchangers; thereby increasing their efficiency. This in turn results in an improved COP of the refrigeration cycle. The

⇑ Corresponding author. Tel.: +91 844 9498 566; fax: +91 1332 285665. E-mail addresses: [email protected], [email protected] (A. Kundu). http://dx.doi.org/10.1016/j.ijheatmasstransfer.2014.04.056 0017-9310/Ó 2014 Elsevier Ltd. All rights reserved.

alternative refrigerants evaluation program (AREP) committee has published an updated list of alternative refrigerants in which some refrigerant mixtures came out; R410A, R410B, R407C and R507. R407C is one of the suitable candidates among them to replace R22 in the appliances working on the near pressure ranges [3]. Two-phase flow heat transfers of refrigerant mixtures in heat exchangers with small diameter tubes has become an important and popular aim of research in recent times because of the demand in the compact heat exchanger design especially in residential and portable refrigeration and air conditioning systems. In addition to reduced refrigeration inventory, improved air-side heat transfer performances can be achieved with smaller tubes, where as a large reduction in tube diameter encountered larger pressure drop. On the other hand, flow boiling of mixtures involves convective and nucleate boiling phenomena simultaneously and that makes the discussion more complex than that for a pure or single component fluid. It is important to comprehend the boiling heat transfer and flow characteristics of pure fluid and refrigerant mixture in the small diameter tubes (lowering the diameter increases the pressure drop [4,5]) consisted in compact heat exchangers as evaporation is a significant component which inflects the performance of a refrigeration system. There are several researches [6–12] have been published comparing the thermodynamic performances of R22 with R134a, R410A, R410B or R507 and some [13–15] for R407C. However, comparisons in two-phase flow characteristics in the context of compact evaporator design and ecological aspects for R407C with R134a are really scarce.

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Nomenclature AN CP D G hev h I k L _ m M P Pr q Q T t V x z

active nucleation sites (K N m/J/kg–kg/m3) specific heat (J/kg K) diameter (m) mass velocity (kg/m2 s) heat transfer coefficient (W/m2 K) specific enthalpy (J/kg) electric current (A) thermal conductivity (W/m K) length of tube in flow direction (m) total mass flow rate (kg/s) molecular weight (dimensionless) pressure (Pa) Prandtl number (CP l/k, dimensionless) heat flux (W/m2) heat transfer rate (W) temperature (°C) time (s) electric potential (V) vapor quality (dimensionless) length (m)

The heat transfer characteristics keep changing as the flow pattern changes along an evaporator tube and the flow regime is influenced by interfacial shear stress, surface tension and gravitational force [12]. In the present study flow boiling heat transfer coefficients of R134a and R407C are measured in a smooth tube inclined at five different angles between 0° and 90° in the direction of refrigerant flow varying heat flux and mass velocity as the flow boiling in inclined tube test data for the aforesaid refrigerants are not available in the literature. The objective of the present experimental study is to develop an accurate flow boiling heat transfer database for the new refrigerants, afford data to the refrigeration industry for the design of high efficient evaporators, compare the thermal performances of R134a with R407Cand investigate the influence of tube inclination, heat flux and mass velocity of the flow boiling characteristics of these fluids. 2. Experimental facility and procedure 2.1. Test facility The experimental plant was designed to investigate two phase heat transfer phenomena during flow boiling under different flow conditions. The schematic diagram of the experimental facility is shown in Fig. 1. The test arrangement (shown in Fig. 2) consists of a semi-hermetic compressor, water-cooled condenser, a thermostatic expansion device, pre-evaporator, post-evaporator and testevaporator. The compressor was connected to a counter flow water cooled shell and tube type condenser. The sub-cooler after the condenser was included for ensuring liquid refrigerant to enter to the Coriolis-effect mass flow meter. The mass flow rate has been controlled through the hand shut off valve fitted on the by-pass line incorporated after the sub-cooler. A quality filter-dryer was accommodated after the flow meter to entrap lubricating oil, foreign particles and moisture in the refrigerant. A suitable pre-heater is designed and installed to control the vapor quality at the inlet of the test evaporator. By regulating power supply with variable voltage AC, heat source, heat input to the pre-heater and test evaporator has been controlled. A suitable accumulator also was installed upstream of the compressor. It was ensured that the refrigerant should be superheated when it enters the compressor.

Greek letters a tube inclination (°) l dynamic viscosity (Pa-s) q density (kg/m3) r surface tension (N/m) D difference Subscripts C critical f saturated liquid g saturated gas IA intermittent to annular flow transition i inside id ideal value in inlet out outlet Sat saturation t top side b bottom side sl left side sr right side

The test section was prepared of the smooth copper tube with inner diameter of 7 mm and wall thickness of 1.26 mm. The outside tube wall temperatures were measured by T-type copper-constantan thermocouples at five axial positions, including inlet and exit to the test evaporator tube. At each section, the temperatures were measured at the top, two sides of the middle and bottom of the tube. The average of these temperatures indicates the local wall temperatures. The local saturation pressures at the inlet and outlet of test evaporator were measured by piezoelectric pressure transducers. The average of these two pressures indicates the saturation pressure at test section. The test evaporator tube was heated by flexible Nichrome (Nickel–Chromium 80–20% by weight) heater wire (3.2 kW capacity at 8 m length, with calibrated accuracy of 2 W) wrapped over the full test length of 1200 mm. Heat flow to the heater was conducted using a variety AC voltage controller and current flow was measured by standard clamp meter to determine applied heat flux. The refrigerant flow patterns were observed through the sight glasses installed before and after the test section. To calculate enthalpy at entry to pre-evaporator, another pressure transducer and T-type thermocouples were fitted at a section just prior to pre-heater. The evaporator tube and preevaporator were well insulated with glass wool from outside to ensure adiabatic condition. Table 1 summates up the features of the plant instrumentation. The temperature and pressure measurement were recorded through a data acquisition system to store data into a personal computer.

2.2. Data reduction The heat transfer coefficient was calculated by using following Eq. (1):

heV ¼





pDi LðT wo  T Sat Þ Di ln ½Do =Di  Q



2

k

1 ð1Þ

This is simply based on heat generation by electrical resistance heating of heater wire outside evaporator tube wall and conduction heat flow through the tube wall to the refrigerant flowing inside at steady state condition. The outside tube wall temperature, T wo ,

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Fig. 1. Schematic of the experimental layout.

Fig. 2. Photographic view of test arrangement.

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Table 1 Measurement equipments and accessories. Quantity

Apparatus

Temperature

T-type thermocouple

Table 2 Uncertainty of variables. Range

50– 250 °C Pressure Piezoelectric transducers 0–20 bar Mass flow rate Coriolis effect mass flow meter 0–200 kg/h Voltage AC variable voltage controller 0–220 V Current Clamp meter 0–100 A

Accuracy

Primary measurements

±0.2 °C

Parameter

Uncertainty

Parameter

Uncertainty (%)

Diameter Pressure Temperature Mass flow rate Heat flux

2 lm 0:5% 0:05  C 0:2% 2:1  3:6%

Reynolds number Mass velocity Heat transfer coefficient Vapor quality Electrical power

0:6 0:6  2 3:9  11 2:0  9:5 0:80

±0.5% F.S. ±0.2% ±0.5% ±1.0% F.S. ±5digits

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

T wo ¼

T t ðzÞ þ T sl ðzÞ þ T sr ðzÞ þ T b ðzÞ 4

ð2Þ

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

Q ¼gV I

ð3Þ

g (typically around 0.935) is experimentally determined using the method proposed by Wambsganss et al. [16]. Average vapor quality is calculated by following Eq. (4): xav g ¼ ðxin þ xout Þ=2

ð4Þ

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

xin ¼

hin  hf ;in hfg;in

ð5Þ

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

hin ¼ h1 þ Q PH=m_

ð6Þ

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

xout ¼

hout  hf ;out hfg;out

ð7Þ

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

hout ¼ hin þ Q TS=m_

ð8Þ

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

vffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi u N  2 uX @R UR ¼ t UV i @V i i¼1

ð9Þ

where UR is the estimated uncertainty in calculating the value of desired variable R, due to the independent uncertainty U V i in the primary measurement of N number of variables Vi, affecting the

Derived quantities

result. The experimental uncertainties for the sensors are listed in Table 2. Maximum uncertainty in the measurement of heat transfer coefficient is shown in Table 3. 2.4. Refrigerant property In the present study R134a and R47C have been tested. R134a is a 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 of the tested refrigerants at different test pressures. The thermodynamic and transport properties of refrigerants were obtained from REFPROP 8.0 [18]. The ranges of experimental parameters are listed in Table 5. 3. Results and discussion In the present paper the heat transfer coefficients are measured for two refrigerants, pure fluid R134a and zeotropic blend R407C, by varying the heat flux with different evaporating pressures and mass velocities while maintaining the refrigerant inlet temperature to the test evaporator between 5 and 9 °C. The experimental conditions are summarized in Table 6. 3.1. Validation In order to establish the integrity of the experimental setup, the heat transfer results of the flow boiling of R134a through the test evaporator tube in horizontal condition are compared with the predicted values by Liu and Winterton [19], Gungor and Winterton [20] and Kandlikar [21] correlations as shown in Fig. 3. This confirms that the experimental results on two-phase flow and flow evaporation using the test facility and measurement system are reliable. The experimental result shows the comparison of ±30% errors with a mean deviation ranging from +2% to 29%. Among them, the mean deviation is the smallest by using Gungor and Winterton’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 experiments as shown in Fig. 4 for R134a and Fig. 5 for R407C for horizontal flow boiling in the present study. A number of flow pattern maps have been proposed over the years for predicting two-phase flow regime transitions in horizontal tubes. In this

Table 3 Maximum uncertainty in heat transfer coefficient. G (kg/m2 s)

Q (W/m2)

Uncertainty (%)

100 200 300

10,000 6000 3000

±11 ±9 ±8.5

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A. Kundu et al. / International Journal of Heat and Mass Transfer 76 (2014) 523–533 Table 4 Properties of test fluids. Fluid

a

PSat (bar)

TSat (°C)

hfg (kJ/kg)

qf

kf (W/m K)

PC (bar)

M

(lPa-s)

CPf (kJ/kg K)

ANa

(kg/m3)

lf

R134a

3.61 3.79 4.01

5.92 7.34 9.00

194.02 192.88 191.55

1274.9 1270.1 1264.4

0.0894 0.0887 0.0880

247.22 242.84 237.81

1.3579 1.3621 1.3672

1.744e06 1.650e06 1.546e06

40.593

102.03

R407C

6.22 6.47 6.86

2.83/8.88 4.07/10.10 5.94/11.92

209.61 208.37 206.4

1225.8 1221.2 1214.2

0.0948 0.0942 0.0933

203.42 200.29 195.68

1.4266 1.4313 1.4386

1.265e06 0.980e06 0.912e06

46.34

86.204

Unit: K N m/J/kg–kg/m3.

Table 5 Range of experimental parameters. Parameters

Range

Refrigerant mass velocity Heat flux Vapor quality Evaporative pressure Inlet temperature

100–300 kg/m2 s 3000–10,000 W/m2 0.1–0.9 3.61–4.01 bar (R134a); 6.17–6.86 bar (R407C) 5–9 °C

Table 6 Operating conditions. Refrigerant

G (kg/m2 s)

Pev (bar)

q (kW/m2)

Dx

R407C

100.7 99.8 101.1 100.2 99.7 198.2 200.4 199.6 201.1 200.8 299.6 302.2 300.8 296.8 301.9

6.18 6.23 6.40 6.61 6.79 6.17 6.24 6.47 6.61 6.76 6.17 6.31 6.59 6.72 6.86

2.96 4.51 6.01 8.23 9.87 3.11 4.56 5.89 8.11 10.12 3.01 4.46 6.20 7.88 10.20

0.81–0.135 0.831–0.129 0.859–0.138 0.839–0.123 0.85–0.141 0.866–0.134 0.874–0.128 0.881–0.127 0.893–0.148 0.874–0.135 0.879–0.131 0.881–0.129 0.888–0.130 0.883–0.136 0.892–0.129

101.1 100.8 98.9 100.2 102.3 199.3 202.1 198.5 201.8 197.8 300.6 301.0 298.5 301.8 299.9

3.61 3.69 3.76 3.87 4.01 3.67 3.74 3.77 3.81 3.98 3.59 3.68 3.79 3.92 4.00

3.06 4.48 6.10 7.78 10.07 2.97 4.60 5.99 8.01 9.72 3.21 4.50 6.08 8.13 10.17

0.818–0.133 0.841–0.139 0.867–0.128 0.859–0.143 0.897–0.134 0.867–0.129 0.891–0.138 0.89–0.131 0.863–0.145 0.888–0.128 0.906–0.149 0.876–0.136 0.846–0.136 0.900–0.150 0.898–0.128

R134a

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

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

paper, the latest version of the Kattan–Thome–Favrat map [22] and the Wojtan–Ursenbacher–Thome map [23] has been used. The flow patterns observed in the experiment are stratified-wavy flow, intermittent flow, semi-annular flow and annular flow. The flow patterns are plotted in the coordinates of G and x for the evaporator tube. The transitions of flow pattern would be stratified-wavy, intermittent, semi-annular, and annular in the direction of increasing G and x. The transition lines are shown with continuous line over the calculated points (calculated from their underlying transition equations, as described in Appendix A) shown; where dashed line denotes the corresponding boundary from the Wojtan– Ursenbacher–Thome map [23]. xIA line shows the calculated flow transition 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 local heat transfer coefficient that occurs at very high vapor qualities after it reaches a maximum in the annular flow regime for R134a; the same occurs earlier for R407C as can be seen from Fig. 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 are reported for R134a and R407C as a function of vapor quality obtained by varying the mass velocity and the heat flux at

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of 0.1–0.2 depending on heat fluxes. In order to explain the dispositions of the R407C and R134a heat transfer coefficients at low vapor qualities, it is necessary to take into account of the equilibrium bubble radius r⁄ which depends on the refrigerant thermophysical properties [26,27]. Bubbles that are smaller in radius than r⁄ will collapse in a spontaneous manner, and bubbles that are bigger in radius than r⁄ will maturate. Hence, a lower r⁄ implies a larger number of active nucleation (AN) sites on the heated surface for bubble formation and, therefore, a stiffer nucleate boiling contribution. The equilibrium bubble radius is directly proportional to the fluid property combination [27] as shown in Eq. (6):

AN ¼

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

horizontal condition of the evaporator tube. In the experimental tests, the heat transfer coefficient increases with increasing heat flux and vapor quality with fixed mass velocity of 100 kg/m2 s and a different dependence of the heat transfer coefficients on vapor quality is apparent as seen from Fig. 6. Indeed, at low values of the vapor quality, the heat transfer coefficient increases with increasing vapor quality presenting a local maximum in the vapor quality range between 30% and 40% for R407C and 65–75% for R134a. The same trend was observed by Wang et al. [13] and Wang and Chiang [25]. The heat transfer progression in flow evaporation results from the interaction between nucleate boiling and liquid convection. The relative consequence of the two different mechanisms varies with vapor quality and strappingly depends on flow conditions. At high heat fluxes and evaporating pressures, heat transfer is mostly prejudiced by the nucleate boiling mechanism [13]. The convective contribution to heat transfer prevails at low heat fluxes and evaporating pressures [26]. In the evaporation process where liquid convection is the main mechanism, convective evaporation being predominant does not implicate that nucleate boiling is entirely inhibited. Under these conditions, the heat transfer coefficient increases with vapor quality as explained by Greco and Vanoli [8]. The present study also provides the same occurrence as shown in Fig. 6; the increment is only 15–23% for R407C; but for R134a, it is quite high as up to 65–110% more with respect to the heat transfer coefficient at very low vapor qualities in the range

2T Sat rf hfg qg

ð10Þ

Here TSat is the bubble point temperature of refrigerant on corresponding evaporating pressure expressed in Kelvin. It can be observed that for the same operating conditions, the R407C property combination AN, is lower than that of R134a. As a consequence, for the same wall superheat, the nucleate boiling contribution of R134a is larger than that of R407C. Besides, due to a prominent temperature glide and variation of the properties of the different constituents inside the zeotropic refrigerant mixture R407C (mixture of R32, R125 and R134a) make the propagation of the boiling phenomena different from that of the single component refrigerant R134a. Wang and Chiang [25] described the phenomena as at the lower quality region (early stage of vaporization), more volatile components of R407C, R32 and R125 evaporated faster than the least volatile ingredient R134a as qg for R32 and R125 (ql of R125 is nearly same to R134) are too high with lower ql for R32 than those of R134a at the corresponding evaporative pressures and the higher mean vapor phase density causes lower mean vapor phase velocity for R407C. Therefore, the lower mean vapor and liquid phase velocity for R407C than R134a may cause a sharp delay in flow pattern transition (can be seen for xIA fro Figs. 4 and 5) and thus, too lower heat transfer coefficient for R407C than that of R134a till the refrigerant leaves the evaporator tube. As the flow proceeds downstream and vaporization takes place, the void fraction increases, thus decreasing the density of the liquid–vapor mixture because of lower density of the vapor. As a consequence, the flow accelerates intensifying convective transport from the heated wall of the tube. The increment in heat transfer coefficient proceeds until the liquid film disappears, leaving the tube wall partially or totally dry. In this region, the heat-transfer coefficient drops sharply because of the low thermal conductivity of the vapor. The same can be observed from the figures that after

Fig. 6. Comparison of heat transfer coefficient with horizontal tube at mass velocity of G = 100 kg/m2 s for different heat fluxes.

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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.

vapor quality of x = 0.75–0.80 for R134a, the heat transfer coefficient drops about 22–27% than that of the maximum one. For R407C, the same occurs earlier, near about 30–40% of the tube. The cause may be the departure from nucleate boiling because of the local void spreads as a vapor blanket on the heating surface for the lower vapor velocity attained during the proceeding of the evaporation; and the bubble crowding and vapor blanketing deflowers the surface cooling by reducing the incoming liquid.

transfer coefficient increases with increase in mass velocity for both the refrigerants; but for R407C, as discussed previous, due to larger void spreading and vapor blanketing over the heated surface, the premature departure from nucleate boiling diminishes the increment in heat transfer and thus the effect of mass velocity increment is not as severe as the case for R134a, which can be seen from Fig. 7.

3.5. Effect of tube inclination on heat transfer coefficient 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 and 10 kW/m2 with the horizontal position of the test evaporator can be observed comparing the test data from the Figs. 6 and 7. Due to mass transfer resistance, nucleate boiling heat transfer coefficient for zeotropic mixtures are considerably lower than pure refrigerants. At low mass velocities, major contribution to the heat transfer mechanism is the nucleate boiling. As the mass velocity increases, the mass transfer resistance to the convective boiling is reduced by the contribution of the more volatile heat transfer coefficient and a higher mass velocity induces liquid entrainment to agitate the mass transfer resistance. The same phenomenon was described by Kattan et al. [22]. At the initial stage of the boiling, due to larger contribution of the nucleate boiling part, the heat

Fig. 8 shows the effect of tube inclination angles on the heat transfer coefficient of both refrigerant R407C and R134a at fixed heat flux of 4.5 kW/m2. The mass velocity of refrigerant has maintained constant at G = 100 kg/m2 s. For all tube inclinations, the heat transfer coefficient increases up to vapor quality x = 0.30– 0.35 for R407C and for R134a, it increases up to x = 0.65–0.75 and then decreases because of dry out. Fig. 9 shows the variation of the local heat transfer coefficient for different tube inclinations at 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 low imposed heat flux inside the evaporator tube, heat transfer coefficient increases with vapor quality up to 30–35% of the tube for R407C. This nature of increment in heat transfer coefficient prevails same for all tube inclinations. When mass velocity increases,

Fig. 8. Comparison of the effect of tube inclination at G = 100 kg/m2 s with constant q = 4.5 kW/m2.

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

the trend in increment continues to vapor quality of 56–65% of the tube for R134a. After that, heat transfer coefficient decreases as long as the fluid leaves the evaporator tube. A comprehensive exploration of convective boiling and nucleate boiling heat transfer coefficient component separately reveals that as evaporation proceeds through the length of the tube and vapor quality increases, due to mass transfer resistance, only convective part increases unhurriedly for zeotropic mixture R407C. The mass transfer resistance to convective evaporation in R407C at high mass velocities used to be reduced significantly due to the disturbances at the mass transfer resistance boundary created by the highly turbulent interface between liquid and vapor interface and that causes a great decrease in average heat transfer coefficient for R407C than that of the pure single component fluid R134a. The effect of tube inclination on the heat transfer coefficient for mass velocity G = 300 kg/m2 s and constant heat flux q = 10 kW/m2 has been compared for the two test refrigerants in Fig. 9. Mass velocity and tube inclination affect the vapor quality at which boiling crisis occurs [12]. The same can be observed from the present study as shown in Fig. 10. The quality, at which the heat transfer coefficient reached maximum, increases with the increase in mass velocity and decreases with an increase in heat flux for R407C as can be seen from Fig. 11; but for R134a, maximum heat transfer coefficient increases with both increase in mass velocity and heat flux for almost all tube inclinations. The vapor quality corresponding to the maximum heat transfer coefficient is worst for 90° tube inclination for both the fluid with low mass

velocity and high heat flux. For mass velocity G = 100 kg/m2 s with R407C, dry out occurs at vapor quality about 34–36% of the evaporator tube with tube inclinations 0–30° for constant heat flux of 6 kW/m2. As the tube inclination increases, dry-out occurs in only 45–48% of vapor quality at a mass velocity of 200 kg/m2 s. Dry-out occurs 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, it is obvious that the tube inclination influences the average heat transfer coefficient significantly. The figure shows the effect of tube inclination angle on the average heat transfer coefficient for fixed heat flux q = 6 kW/m2 with mass velocity varying from 100 to 300 kg/m2 s. It reveals that for all mass velocities, the local heat transfer coefficient is best at 90° of tube inclination for both the refrigerant; where after dry-out, 90° inclination is the worst as also can 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 to the horizontal position of the evaporator tube at a low mass velocity of 100 kg/m2 s with heat flux of 6 kW/m2, which can be observed from Fig. 12. But the average heat transfer coefficient increases to 3–14% of tube inclination increases from 30° to 90° with respect to the horizontal position of the evaporator tube, most of which lies between the experimental uncertainty limit, when mass velocity increases to 200 kg/m2 s. Before dry out, the average heat transfer coefficient increases about 10–34% of tube inclination increases from 30° to 90° with respect to horizontal tube. The

Fig. 10. Comparison of the effect of tube inclination at constant G = 300 kg/m2 s with q = 3 kW/m2 and q = 10 kW/m2.

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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 and R134a.

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.

effect of tube inclination goes up with increment in mass velocity within intermittent flow, but descends as soon as annular flow regime prevails. At G = 300 kg/m2 s, average heat transfer coefficient 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 accumulated prematurely by the effect of early dry-out of refrigerant mixture R407C and causes less liquid to flow faster; thus creates a delay in flow pattern transition at near the beginning stage of annular flow or semi-annular flow. In comparison to R407C, the horizontal position of the evaporator tube for R134a, the average heat transfer coefficient in inclined tube increases only 4% and 9% for mass velocity of 300 kg/m2 s for varying tube inclination from 30° to 90° which were also within the uncertainty for experimental results. Vertical flow of both the refrigerants has the worst heat transfer coefficient for all mass velocities irrespective of heat flux after dry out. For mass velocity G = 300 kg/m2 s with heat flux applied of 6 kW/m2, the average heat transfer coefficient for vertical up flow boiling of R134a is about four times more than that for R407C, where at 0° and 45°, it is about 325% and 300%, respectively. For mass velocity G = 100 kg/m2 s, the average heat transfer coefficient for horizontal flow evaporation of R134a 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. Average heat transfer coefficient, in general, is defined as the mean of local heat transfer coef-

ficients at different vapor qualities for a given imposed heat flux and constant mass velocity. The comparison of the flow boiling performance inside the evaporator tube inclined at 45° with different mass velocities and 8 kW/m2 heat flux has been drawn in Fig. 13. At this inclination, heat transfer coefficient increases with mass velocity, but the difference in increment decreases with respect to those in the horizontal position of the tube, if we compare those results as shown in Figs. 7 and 8. The vapor phase velocity has been reduced for low vapor quality resulting in a corresponding decrease in inertia force. In low vapor quality region, thus, the gravitational force affects the two-phase flow in a considerable manner and this causes the variation of heat transfer coefficient for different tube inclinations. Also the variation of heat transfer coefficient in vapor quality for a particular tube inclination has uneven behavior due to the transition of different flow pattern through the growth of the boiling along the tube length. From Figs. 10, 11 and 13, it is evident that for high heat flux, as mass velocity increases, inclination in evaporator tube demeans the boiling performance of refrigerants. Average heat transfer coefficient increases 7% with R407C and 43% with R134a for an increase in mass velocity from 100 to 300 kg/m2 s at an evaporator tube inclination of 45°; and 6% and 41% increase for inclination of 60° with heat flux of 8 kW/m2 for R407C and R134a, respectively. For horizontal tube, the increment is 4% for R407C and for R134a, it is 47% as the mass velocity

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

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°.

increases the same. The cause may be the buoyancy effect of vapor, which accelerates the upstream flow at a higher inclination other than horizontal would augment the convective contribution rather than nucleate boiling part and the convective boiling coefficient arouses a gentle but nucleate boiling being suppressed more with increase in vapor quality as vaporization proceeds. This phenomenon also has been demonstrated by Kattan et al. [22] by describing the lower heat transfer coefficient for downward flow with respect to upward and horizontal flow of pure refrigerant. However, the interfacial shear stress has been increased by the turbulence created by the very high mass velocity and thus, the bubbles formed on the heated surface of the tube could not attain at a large size as seen in nucleate boiling. By increasing the mass velocity, the influence of nucleate boiling, which is the dominant part of the low vapor quality region became less effective to contribute in flow boiling of refrigerant upstream. And for this, in the current study for R407C and R134a, it is revealed that the average heat transfer coefficient increases 8% and 18%, respectively, for an increase in mass velocity from 200 to 300 kg/m2 s i.e. transition from intermittent and semi-annular region to annular flow with q = 10 kW/m2 but also an eminent decrease in local heat transfer coefficient occurs from inlet to outlet of the evaporator tube by 10%, 13% and 21% for R407C at mass velocities of 100, 200 and 300 kg/m2 s, respectively. But as the nucleate boiling contribution to flow evaporation of pure fluid R134a is much more, as described by Eq. (10) and shown in Table 4, it is seen that the heat transfer coefficient much more increases at the outlet of the evaporator tube with respect to that at the entry to the evaporator; but considerable amount decrease of 22–24% of that after a dry-out of the exit of the evaporator tube.

4. Conclusion 1. An experimental plant has been established in the refrigeration laboratory of Indian Institute of Technology Roorkee, India for evaluating the heat transfer performance of pure and mixed refrigerants during convective boiling. Local heat transfer coefficients of pure fluid R134a and zeotropic mixture R407C, in flow evaporation through an inclined tube with five inclination angles from 0° to 90° in the direction of refrigerant flow (inclination increased upward) were measured at the evaporating pressure range of 3.61–4.01 for R134a and 6.22–6.86 bar for R407C with the temperature at the entry to the test evaporator

was maintained the same for both the fluid between 5 and 9 °C depending on evaporative pressures, the heat flux from 3 to 10 kW/m2 and the mass velocity from 100 to 300 kg/m2 s. The effect of mass velocity, heat flux and vapor quality; and tube inclinations on evaporative heat transfer coefficient has been investigated for different flow regimes. 2. The experimental results indicate that the heat transfer coefficients increase with mass velocity and heat flux. Also the local heat transfer coefficient increases for both the refrigerants before the dry-out occurs. The heat transfer coefficient increases with increasing vapor quality presenting a local maximum in the vapor quality range between 30% and 40% for R407C and 65–75% for R134a. 3. The nature of the increment in heat transfer coefficient with respect to mass velocity, heat flux and vapor quality varies with the flow patterns for different refrigerants considered for the test. The increase in heat transfer coefficient was at much lower quality for R407C than R134a mainly due to the nature of the contribution of nucleate boiling to the evaporation progression of the refrigerants. 4. The effect of tube inclination was also much more severe on the performance of R134a than R407C. The heat transfer coefficients of pure fluid R134a are higher than that of refrigerant blend R407C for all mass velocities, heat fluxes and tube inclinations. For mass velocity G = 300 kg/m2 s with heat flux applied of 6 kW/m2, the average heat transfer coefficient for vertical up flow boiling of R134a is about four times more than that for R407C, where at 0° and 45°, it is about 325% and 300%, respectively. For mass velocity G = 100 kg/m2 s, the average heat transfer coefficient for horizontal flow evaporation of R134a 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, q Physical 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

A.2. Flow pattern boundaries calculation Stratified to Stratified Wavy transition is calculated from:

GStratified

( )1=3 ð226:3Þ2 Afd A2gd qg ðqf  qg Þlf g ¼ x2 ð1  xÞp3

ðA1Þ

Stratified wavy to intermittent/annular boundary is calculated from: GWav y ¼

8 <

"

16A3gd gDi qf qg

p2

:x2 p2 ð1  ð2H  1Þ2 Þ0:5 25H2fd fd

ð1  xÞF 1 ðqÞ



#91=2 F 2 ðqÞ = We þ1 þ 50 ; Fr f ðA2Þ

2

where F 1 ðqÞ ¼ 646:0ðq=qC Þ þ 64:8ðq=qC Þ; F 2 ðqÞ ¼ 18:8ðq=qC Þ þ 1:023 h i1=4 . and qC ¼ 0:16q1=2 g hfg g rðqf  qg Þ Intermittent to annular flow transition is calculated from:

xIA ¼



1  1=1:75  1=7  þ1 0:2914 qg =qf lf =lg

ðA3Þ

Annular to dry-out boundary is calculated from:

GDry-out

8 > > <

90:926

0:58  0:17 > > ln x þ 0:52 rDqi  = g 0:37   ¼    0:25 0:70 > > q > > 1 : ;  qg  qq gDi qg ðqf qg Þ C f 1 0:235

ðA 4Þ

Dry-out to Mist is calculated from:

GMist ¼

8 > > < > > :

1 0:0058



ln

90:943  0:38 > >  þ 0:57 rDqi = g 0:15     0:27 > q 0:09 > ;  qg  qq

0:61

1

gDi qg ðqf qg Þ

x

f

ðA 5Þ

C

A.3. Flow pattern classification Stratified flow pattern Stratified-wavy flow pattern

Intermittent flow pattern Annular flow pattern Dry-out flow pattern Mist flow pattern

G < Gstrat G > Gwavy(xIA) gives the slug zone Gstrat < G < Gwavy(xIA) and x < xIA give the slug/stratified-wavy zone x P xIA gives the stratified-wavy zone Gwavy < G < Gdry-out and x < xIA Gwavy < G < Gdry-out and x > xIA If Gstrat P Gdry-out, then Gdry-out = Gstrat If Gwavy P Gdry-out, then Gdry-out = Gwavy If Gstrat P Gmist, then Gmist = Gstrat If Gwavy P Gmist, then Gmist = Gwavy

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