Experimental study on flow boiling of R134a and R410A in a horizontal microfin tube at high saturation temperatures

Applied Thermal Engineering 31 (2011) 3814e3826

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Experimental study on flow boiling of R134a and R410A in a horizontal microfin tube at high saturation temperatures Andrea Padovan, Davide Del Col*, Luisa Rossetto Dipartimento di Fisica Tecnica, Università di Padova, Via Venezia 1, 35131 Padova, Italy

a r t i c l e i n f o

a b s t r a c t

Article history: Received 21 September 2010 Accepted 16 July 2011 Available online 27 July 2011

This paper presents an experimental study on vaporization of R134a and R410A inside a horizontal microfin tube at 30
C and 40
C saturation temperatures. A wide range of operating conditions is investigated: mass flux from 80 to 600 kg m2 s1, heat flux from 14 to 83.5 kW m2 and vapour quality from 0.1 to 0.99. The experimental database includes measurements of heat transfer coefficient and dryout inception vapour quality. The heat transfer data are compared with those for a plain tube at similar operating conditions. The experimental data are then used to assess the accuracy of models for the prediction of the heat transfer coefficient and dry-out vapour quality. The present data are measured at high evaporating temperature, covering values of reduced pressure between 0.19 and 0.49, which are higher than usual for air-conditioning applications with halogenated refrigerants. The main applications for the present study are related to heat pumps reaching high condensing temperature (above 70
C) and high evaporating temperature (above 30
C), although other applications of evaporators operating at high reduced pressure may also benefit from the present data. Ó 2011 Elsevier Ltd. All rights reserved.

Keywords: Flow boiling Microfin R134a R410A High reduced pressure Heat pumps

1. Introduction Most studies available in the open literature on flow boiling inside horizontal tubes report heat transfer data sets of halogenated fluids at low values of saturation temperature (usually from 15
C to 20
C), which are typical of refrigeration and airconditioning applications. Nevertheless, there is need for a validation of existing heat transfer and dry-out models at higher values of saturation temperature and reduced pressure, especially in the case of enhanced tubes, where the number of available data is much lower. A primary application for high temperature flow boiling data is represented by some heat pumps. In the heat pumping equipment the evaporator is a key component since it strongly affects the overall performance (Kim et al. [1]). An example of high evaporating temperature application with HFC fluids is represented by the heat pump clothes dryers, where the refrigerant in the evaporator extracts heat from the hot and humid air stream, reaching much higher values of saturation temperature as compared to common applications. From the analysis of the R134a drying cycle reported by Mancini et al. [2], one can see that the evaporating temperature can be higher than 30
C. * Corresponding author. E-mail address: [email protected] (D. Del Col). 1359-4311/$ e see front matter Ó 2011 Elsevier Ltd. All rights reserved. doi:10.1016/j.applthermaleng.2011.07.026

Abou-Ziyan et al. [3] investigated solar assisted R134a heat pumps in a wide range of evaporating temperatures (0e45
C) and condensing temperatures (50e70
C). There are also examples of heat pumps used to produce hot water for industrial applications that recover and utilize the waste heat as source for the evaporator. For instance, Wang et al. [4] have recently studied the performance of a heat pump using parallel cycles with serial heating on the water side: in the high temperature stage of the heat pump, the inlet water temperature at the evaporator is 45
C, whereas the condensing temperature is between 70
C and 90
C. However, the present study addresses the effect of high saturation temperature and high reduced pressure on the flow boiling heat transfer and thus it may be useful for a bunch of other applications. In reboilers for chemical and petrochemical processes, for example, new data regarding the effect of temperature on the accuracy of predicting models should be highly useful. Another example is represented by the use of halogenated fluids in ORC cycles, where the optimization of evaporators working at high values of reduced pressure is searched. Coming back to the heat pump applications, Kew and Reay [5] have pointed out a growing interest on the use of CO2 as working fluid for heat pump water heaters and domestic air-conditioning systems, due to the global warming issue. The use of carbon

A. Padovan et al. / Applied Thermal Engineering 31 (2011) 3814e3826

dioxide asks for the study of vaporization at high values of reduced pressure. Table 1 reports the thermodynamic and thermophysical properties of R134a, R410A and CO2 as a function of saturation temperature. At Ts ¼ 5
C, CO2 displays 0.39 reduced pressure, which is far higher than the values of halogenated refrigerants at the same temperature. For R410A, the value of 0.39 reduced pressure is associated with 30
C saturation temperature. When comparing the three refrigerants reported in Table 1 at the same saturation temperature, a higher reduced pressure results in a higher vapour density, lower surface tension, higher vapour viscosity and this yields flow boiling characteristics quite different depending on the fluid. But, when comparing the three fluids at the same reduced pressure, similar values of liquid to vapour density ratio, liquid to vapour viscosity ratio and surface tension are observed. So, the reduced pressure appears as a key parameter to establish the flow pattern and the heat transfer features in vaporization. Del Col [6] has recently published experimental heat transfer coefficients of halogenated refrigerants during flow boiling inside a horizontal smooth tube at saturation temperature between 25
C and 45
C (reduced pressure from 0.19 to 0.53). At Ts ¼ 40
C the heat transfer coefficient of R410A decreases when vapour quality increases, exhibiting a trend similar to the one of carbon dioxide at the same values of reduced pressure. As compared to smooth channels, microfin tubes ensure a large heat transfer enhancement with a relatively low pressure drop increase and reduce the range of operating conditions leading to dry-out and partial dry-out. The presence of micro-fins may change the two-phase flow pattern and the relative importance of nucleate boiling and convective evaporation heat transfer mechanisms. Moreover, flow pattern investigations by Yoshida et al. [7], during vaporization inside a horizontal spirally grooved tube, revealed the presence of a mechanism of liquid transport from the bottom to the top of the tube at low mass velocity. This mechanism was attributed to the action of surface tension, promoted by the presence of microgrooves. At G ¼ 50 kg m2 s1, the liquid in the grooves near the top of the tube appeared to be almost stagnant, but when increasing both mass velocity and vapour quality the liquid velocity in the spiral grooves increased too. At 200 kg m2 s1 mass velocity and vapour quality lower than 0.8 they visualized an annular flow pattern characterized by a liquid film thicker than the groove height. Del Col et al. [8] suggested that during two-phase annular flow inside a microfin tube the surface tension effect would exist only at high vapour quality, when the liquid does not flood the grooves. Bandarra Filho and Saiz Jabardo [9] compared flow boiling heat transfer coefficients measured in a plain and a microfin tube at Ts ¼ 5
C, G ¼ 100 kg m2 s1 and q ¼ 5 kW m2. In the case of the Table 1 Thermodynamic and thermophysical properties of R134a, R410A and CO2 at different saturation temperatures.

Ts [
C] ps [kPa] pred rL[kg m3] rV [kg m3] rL/rV mL [mPa s] mV [mPa s] mL/mV lL [mW m1 K1] lV [mW m1 K1] s [N m1]

R134a

R134a

R410A

R410A

CO2

5 350 0.086 1278 17 74.6 250.1 10.9 22.9 89.81 11.95 0.01084

30 770 0.190 1187 38 31.6 183.1 11.9 15.4 78.99 14.34 0.00742

5 935 0.191 1150 36 32.1 158.9 12.9 12.3 106.72 12.85 0.00829

30 1886 0.385 1033 77 13.5 113.2 15.0 7.6 91.55 16.70 0.00460

5 3046 0.392 956 83 11.5 108.4 14.3 7.6 116.50 18.17 0.00550

3815

plain tube, the heat transfer coefficient remains essentially constant over the whole range of vapour quality and this was associated with the stratified flow pattern in the visualizations. For the microfin tube, instead, a flow pattern similar to the annular one was observed by the authors and the corresponding heat transfer coefficient increased with vapour quality. This paper presents an experimental study on vaporization of R134a and R410A inside a horizontal microfin tube at 30
C and 40
C saturation temperatures (reduced pressure from 0.19 to 0.49) in a wide range of operating conditions: mass flux from 80 to 600 kg m2 s1, heat flux between 14 and 83.5 kW m2 and vapour quality from 0.1 to 0.99. The experimental database includes measurements of heat transfer coefficient and dry-out inception vapour quality. The present data of heat transfer coefficients are discussed and then used to assess the prediction accuracy of the models by Cavallini et al. [10], Chamra and Mago [11], Hamilton et al. [12], Koyama et al. [13] and Thome et al. [14] at high values of saturation temperature and reduced pressure. The data of dry-out inception vapour quality are compared against the predictions given by the Mori et al. [15] correlation. A modification of their correlation is proposed to predict the present data. 2. Experimental set up and measuring procedure 2.1. Experimental apparatus The measurements have been taken on the experimental set up available at the two-phase heat transfer laboratory of the University of Padova, Italy. A schematic view of the experimental apparatus is reported in Fig. 1. The test facility is composed of three loops: the refrigerant loop, the heating water loop and the cooling water loop. In the primary loop, the refrigerant is pumped as subcooled liquid in the pre-heater, where it is heated and in some cases partially vaporized to achieve the desired vapour quality at the inlet of the test section. The test section is a microfin copper tube equipped for the measurement of the heat transfer coefficient. Both the test section and the pre-heater are counter flow tube in tube heat exchangers in which the refrigerant, flowing inside the tube, is heated and vaporized by hot water flowing in the annulus. After the two-phase mixture has left the test section it goes to a braised plate type condenser, where it is fully condensed and subcooled by the cooling ground water. A bladder accumulator, connected to a nitrogen bottle and a pressure regulator, is installed in the refrigerant loop: thus, the refrigerant pressure can be controlled by varying the charge of nitrogen in the accumulator. The pump is a variable speed gear pump, magnetically driven and oil free, which allows to control independently the refrigerant mass flow rate. 2.2. Test section and instrumentation The test section is obtained using a single microfin copper tube, 1400 mm long, and three separate hot water jackets. In this way, the microfin tube is divided in a pre-section, 600 mm long, the measuring heat transfer section, 300 mm long, and a post-section, 200 mm long. In the pre-section the refrigerant achieves a fully developed flow regime and the vapour quality can be adjusted before entering the measuring section. The measurement of the heat transfer coefficient is done in the central sector. Finally, the post-section is used to avoid disturbance to the measurements due to heat conduction along the wall. The tested tube is a commercial microfin tube provided by Hitachi Cable. It has 60 fins, smoothed at the tip, with 0.23 mm fins height, 13
helix angle and 43
apex angle. The inside diameter,

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A. Padovan et al. / Applied Thermal Engineering 31 (2011) 3814e3826

Fig. 1. Schematic view of the experimental apparatus.

measured at the fin tip, is equal to 7.69 mm. A metal helix is wounded around the outside surface of the tube, inside the annulus, to avoid stratification in the water flow and to get high heat transfer coefficient on the water side with reasonable values of water temperature decrease. The measuring heat transfer section is instrumented with four copper-constantan (type T) thermocouples embedded in its wall to measure the surface temperature of the microfin tube. The thermocouples are located in the middle point of the section; they are inserted and soldered into four equidistant axial grooves arranged circumferentially as reported in the draw of Fig. 2.

The water temperature change across each heat transfer sector (pre-heater, pre-section, measuring section and post-section) is measured by means of four-junctions copper-constantan thermopiles inserted into appropriate adiabatic mixing chambers, where copper-constantan thermocouples are also inserted to measure the water temperature. Magnetic type flow meters are used to measure the water flow rate. A digital Rosemount pressure transducer is connected to a manometric tap to measure the refrigerant pressure immediately upstream the microfin tube. The pressure drop along the entire length of the microfin tube is measured by a digital

Fig. 2. Details of the experimental test section.

A. Padovan et al. / Applied Thermal Engineering 31 (2011) 3814e3826

EndresseHauser differential pressure transducer connected to manometric taps upstream and downstream the test tube. The refrigerant temperatures at inlet and outlet of the test section and the pre-heater are measured in adiabatic sectors by means of thermocouples inserted into the refrigerant flow and the tube wall. The refrigerant mass flow rate is measured by a Coriolis effect Micromotion mass flow meter placed downstream the pump.

2.3. Measuring technique and data reduction The measured heat transfer coefficients are quasi-local values, which means that the vapour quality change across the measuring section is small, lower than 0.20. The heat transfer coefficient is obtained as the ratio of the heat flux to the temperature difference between wall and refrigerant:

HTC ¼

Q
AðTwall  T s

(1)

where Q is the heat flow rate transferred to the refrigerant, A is the heat transfer surface area, Twall and Ts are the wall and saturation temperatures, respectively. The heat transfer surface area is that of a smooth tube with the same diameter as the fin tip diameter of the microfin tube. The wall temperature, Twall, is obtained as the average value between the readings of the four thermocouples located circumferentially in the middle of the measuring section. The local saturation temperature, Ts, is determined from the measurements of the saturation pressure at inlet and outlet of the microfin tube: the local pressure in the middle of the measuring section is obtained by assuming a linear profile of the pressure between inlet and outlet. The heat flow rate, Q, comes from a water side energy balance on the measuring section and by the measurement of the heat flow rate dissipated to the surrounding air ambient:

Q ¼ mw cp;w DTw  Qd

(2)

The heat flow rate dissipated to the ambient, Qd, has been determined from specific test runs. To perform these tests, the microfin tube has been vacuumed and water has been sent at various temperatures. In this way, the heat dissipation has been measured from a water side thermal balance. The ambient air temperature has been also measured in order to correlate the heat dissipation with the temperature difference between the water in the measuring section and the ambient air. The heat flow rate dissipated to the surroundings goes from 2% of Q for the data taken at Ts ¼ 30
C, G ¼ 600 kg m2 s1 and q ¼ 88 kW m2 up to 14% when Ts ¼ 40
C, G ¼ 100 kg m2 s1 and q ¼ 14 kW m2. By using the same technique, the heat dissipation has been also measured for the other heat transfer sectors (pre-heater, presection and post-section) and a proper correlation has been established for each section and used in the data reduction. The local thermodynamic vapour quality at the middle point of the measuring section is obtained as:

uHTC

x ¼ xin þ

Dx ¼

3817

Dx

(3)

2

Q mr hLV

(4)

where xin is the vapour quality at the inlet of the heat transfer measuring section, Dx is the vapour quality change across the measuring section, Q is the heat flow rate transferred from the heating water to the refrigerant, mr is the refrigerant mass flow rate and hLV is the latent heat of evaporation at the local pressure. The vapour quality at the inlet of the heat transfer measuring section, xin, is given by an energy balance performed on the preheater and the first sector of the test section:






Dhpreh þ Dhpres ¼ hL;s  hL;sub þ hLV xin

(5)

where Dhpreh and Dhpres are the specific enthalpy variations of the refrigerant in the pre-heater and pre-section, respectively, hL,s is the specific enthalpy of the saturated liquid refrigerant at the local pressure, hL,sub is the specific enthalpy of the subcooled liquid refrigerant and hLV is the latent heat of evaporation at the local pressure. The specific enthalpy of the subcooled refrigerant at the inlet of the pre-heater is known from the measurements of pressure and temperature. The specific enthalpy variations, Dhpreh and Dhpres, derive from the heat flow rate transferred from the heating water to the refrigerant. Thermophysical and thermodynamic properties needed for the reduction of experimental data have been obtained from Refprop 7.0 [16]. The quasi-azeotropic mixture R410A has a low temperature glide, around 0.1
C; thus, the fluid properties used in the data reduction are based on the average temperature between bubble and dew point at the local pressure. 2.4. Calibration procedure and experimental uncertainty analysis The experimental uncertainty analysis has been done in agreement with the guidelines provided by ISO [17]. The standard uncertainty of a measurement is composed of type A and type B evaluations of uncertainty. Type A evaluation derives from the statistical analysis of the repeated measurements taken under steady state conditions during the recording time interval. Type B uncertainty is due to instrument and data logger used and it is provided by the manufacturers or specific calibration procedures. The standard uncertainty is obtained by combining both the uncertainty terms:

u ¼

qffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi u2A þ u2B

(6)

where uA and uB are the type A and type B uncertainties, respectively. When a parameter is indirectly determined from the measurements of other quantities, the experimental uncertainty is obtained by means of the law of uncertainty propagation. For example, starting from Eq. (1) and Eq. (2), the standard uncertainty of the heat transfer coefficient is given as:

sffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 2  2  2  2  cp;w DTw mw cp;w HTC 1 ¼ uTwall Ts þ umw þ uDTw þ uQd Twall  Ts AðTwall  Ts Þ AðTwall  Ts Þ AðTwall  Ts Þ

(7)

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A. Padovan et al. / Applied Thermal Engineering 31 (2011) 3814e3826 Table 2 Type B expanded uncertainty (95% level of confidence) of sensors. 0.05 K 0.05 K 1% 0.4% 1 kPa 200 Pa

Thermocouples Thermopile Water Flow meter Refrigerant Mass Flow meter Pressure transducer Differential pressure transducer

where the four terms in the brackets are the uncertainty contributions due to water mass flow rate, water temperature difference, wall to saturation temperature difference and heat dissipation. The uncertainty contributions due to the heat transfer surface area and the fluid properties can be neglected in comparison with the other terms. The standard uncertainty is then multiplied by an appropriate coverage factor to get a confidence level of 95%. Table 2 reports a list of the type B uncertainty of the sensors used here. The reported uncertainties refer to the expanded values with 95% of confidence level. In the case of thermocouples and thermopiles, the uncertainty comes from an off-site calibration procedure; their accuracy has been then checked on-site by comparison with calibrated thermistor probes Pt100, getting the uncertainty values reported in Table 2. Some additional checks have been performed to ensure high accuracy of the experimental tests. The energy balance on the heat transfer sections has been checked, by comparing the water side heat flow rate to that measured on the refrigerant side when subcooled liquid enters and superheated vapour exits. When considering the heat flow rate dissipated to the ambient, the deviation between water side and refrigerant side heat flow rate is lower than 2.5%. Wall and local saturation temperatures have been compared to each others during adiabatic two-phase flow. For instance, at 400 kg m2 s1 mass velocity the deviation between the two measurements was within the uncertainty range of sensors. The refrigerant temperature and pressure at the inlet of the presection have been regularly checked under saturated conditions: when the pure fluid R134a flows at 400 kg m2 s1 mass velocity, the temperature directly measured by the thermocouples typically differs by less than 0.15 K from the value reduced from the pressure, even at vapour quality around 0.2. Larger deviations were found at low values of mass velocity and vapour quality, probably due to the

non equilibrium in the flow and local subcooling at the inlet of the test section. The sensitivity of heat transfer data to the refrigerant conditions at the inlet of the microfin tube has also been investigated. The present analysis has shown no difference when the refrigerant enters the tube as subcooled or saturated liquid. Table 3 reports the operating conditions of the present experimental database. Each data set has been obtained varying the vapour quality at fixed saturation temperature, mass velocity and heat flux. The expanded uncertainty of the heat transfer coefficient, with 95% of confidence level, goes from 5% to 10%, before the inception of dry-out. Higher uncertainties, up to 15%, can be observed for the data at G ¼ 200 kg m2 s1 under dry-out condition. In the measurement of heat flux, the water temperature difference has been maintained higher than 1 K in order to reduce the uncertainty contribution associated with the thermopile. At the same time, no measurements of the heat transfer coefficient have been taken when DTw was higher than 2 K to keep low the wall temperature change between inlet and outlet. The wall superheat goes from 1 K to 6 K; values below 1.3 K have been found only for a limited number of data measured at G ¼ 200 kg m2 s1, q ¼ 15 kW m2 and x > 0.8. For most data, the expanded uncertainty of the vapour quality goes from 0.001 up to 0.020. Uncertainties between 0.02 and 0.03 have been found only for the data taken at G ¼ 80 kg m2 s1 and vapour quality greater than 0.55. A more detailed uncertainty analysis is reported in [18].

3. Heat transfer experimental results In this section the experimental measurements of heat transfer coefficient and dry-out inception vapour quality are presented. The heat transfer coefficients reported in the graphs represent quasilocal values. The dry-out inception vapour quality is determined as the last point before a drop of the heat transfer coefficient larger than the experimental uncertainty. In Fig. 3 the measured heat transfer coefficients are plotted as a function of vapour quality for R134a at around 30
C saturation temperature, 15 kW m2 heat flux and three different values of mass velocity: 80, 100 and 200 kg m2 s1. At G ¼ 80 kg m2 s1 and G ¼ 100 kg m2 s1 and vapour quality lower than around 0.35, the

Table 3 Experimental conditions of the present database (the ranges of saturation temperature and heat flux are reported for each data set; the mass flux varies by less than 1%). Fluid

pred [/]

Ts [
C]

G [kg m2 s1]

q [kW m2]

x [/]

R134a

0.19e0.27

R410A

0.39e0.49

31.0 (0.4, þ0.5) 30.3 (0.5, þ0.6) 30.3 (0.6, þ0.5) 30.5 (1.3, þ0.5) 30.5 (0.4, þ0.5) 30.0 (0.1, þ0.2) 30.2 (0.6, þ0.5) 30.3 (0.8, þ0.4) 29.7 (1.0, þ0.5) 29.7 (0.1, þ0.1) 29.8 (0.1, þ0.2) 29.8 (0.4, þ0.3) 42.3 (0.7, þ0.6) 30.3 (0.2, þ0.1) 30.2 (0.2, þ0.2) 30.2 (0.5, þ0.2) 30.1 (0.1, þ0.1) 39.9 (0.1, þ0.1)

80 100 100 200 200 200 400 400 400 600 600 600 200 200 400 400 600 400

14.7 (5%, þ8%) 14.8 (6%, þ8%) 21.5 (5%, þ3%) 14.9 (9%, þ6%) 22.3 (6%, þ4%) 42.4 (1%, þ1%) 22.3 (7%, þ7%) 28.5 (4%, þ4%) 43.3 (5%, þ8%) 42.4 (2%, þ2%) 58.9 (2%, þ2%) 83.5 (1%, þ1%) 14.0 (10%, þ10%) 44.1 (2%, þ3%) 21.4 (3%, þ3%) 44.2 (5%, þ3%) 44.2 (2%, þ3%) 21.1 (2%, þ3%)

0.20e0.88 0.16e0.97 0.16e0.98 0.13e0.99 0.17e0.97 0.22e0.91 0.19e0.97 0.10e0.98 0.24e0.94 0.16e0.75 0.19e0.57 0.10e0.64 0.17e0.99 0.18e0.95 0.21e0.99 0.17e0.97 0.20e0.80 0.17e0.99

A. Padovan et al. / Applied Thermal Engineering 31 (2011) 3814e3826

20 Ts 31.0, G 80, q 14.7 Ts 30.3, G 100, q 14.8

18

Ts 30.5, G 200, q 14.9

HTC [ kW m K ]

16

R134a

14 12 10 8 6 4 2 0

0

0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9

1

Vapour quality Fig. 3. Experimental heat transfer coefficient vs. vapour quality for R134a at 80, 100 and 200 kg m2 s1 mass velocity. The average values of saturation temperature, Ts, mass velocity, G, and heat flux, q, are reported for each data set in the legend of the graph.

heat transfer coefficient does not increase either with mass velocity or vapour quality. At higher vapour quality (above 0.35), the heat transfer coefficient increases with both vapour quality and mass velocity until the inception of dry-out. When mass velocity is increased to 200 kg m2 s1, the heat transfer coefficient is higher than that measured at G ¼ 100 kg m2 s1 over the whole vapour quality range. Fig. 4 reports the heat transfer coefficients measured at values of mass velocity between 100 kg m2 s1 and 400 kg m2 s1; the

20 Ts 30.3, G 100, q 21.5 Ts 30.5, G 200, q 22.3 Ts 30.2, G 400, q 22.3

18

HTC [ kW m K ]

16

R134a

14 12 10 8 6 4 2 0

0

0.1

0.2

0.3

0.4

0.5

0.6

0.7

0.8

0.9

1

Vapour quality Fig. 4. Experimental heat transfer coefficient vs. vapour quality for R134a at 100, 200 and 400 kg m2 s1 mass velocity. The average values of saturation temperature, Ts, mass velocity, G, and heat flux, q, are reported for each data set in the legend of the graph.

3819

heat flux is around 22 kW m2 and the saturation temperature is 30
C. At G ¼ 100 kg m2 s1 the trend of heat transfer coefficient versus vapour quality is similar to that observed in Fig. 3. At G ¼ 200 kg m2 s1 the heat transfer coefficient remains constant until x ¼ 0.3 and then increases with vapour quality. When comparing the heat transfer coefficients at G ¼ 200 kg m2 s1 and G ¼ 400 kg m2 s1, no effect of mass velocity is observed, except for the high vapour quality region, where the dry-out occurs first for the lower mass velocity data set. From the analysis of low mass velocity data (G ¼ 80 kg m2 s1 and G ¼ 100 kg m2 s1) reported in Figs. 3 and 4, one can deduce that heat transfer is nucleate boiling dominated at low vapour quality. In fact, the heat transfer coefficient is almost independent of vapour quality and mass velocity, but it depends on heat flux. After a certain vapour quality, the heat transfer coefficient begins to increase and this can be explained with the transport of the liquid in the micro-grooves towards the top of the tube, due to the surface tension effect. Moreover, as visualized by Yoshida et al. [7], the liquid velocity in the micro-grooves should increase with vapour quality, contributing to get higher heat transfer coefficients. When G  200 kg m2 s1, the two-phase flow is noticeably increased and so the convection mechanism: the heat transfer coefficient is higher than that at G ¼ 100 kg m2 s1 even in the low vapour quality region. In Fig. 5 the experimental heat transfer coefficients for R134a and R410A are plotted as a function of vapour quality for values of mass velocity between 200 and 600 kg m2 s1, saturation temperature around 30
C and similar values of heat flux. For both fluids, mass velocity does not seem to have a great influence on the heat transfer performance, although the heat transfer coefficient decreases a little as mass velocity increases: the nucleate boiling suppression, associated with the two-phase flow, increases with mass velocity and it seems to prevail over the enhancement of convection. However, the difference of heat transfer coefficients is within the experimental uncertainty of data. This same behaviour has been also observed in Fig. 4, when comparing data sets at G ¼ 200 kg m2 s1 and G ¼ 400 kg m2 s1. In Fig. 6 the heat transfer coefficients measured in the microfin tube are compared to the values taken in a horizontal plain tube (8 mm inside diameter) by Del Col [6] at similar operating conditions. The ratio of the effective heat transfer surface area of the microfin tube to the one in the smooth tube is equal to 1.8. The heat transfer coefficients are significantly higher in the microfin tube, exhibiting a peak just before the onset of dry-out. Moreover, in the microfin tube the dry-out inception is shifted to higher values of vapour quality. In fact, when the annular flow pattern takes place, the thickness of the liquid film diminishes continuously with the increase of vapour quality, but it does not break, because the microfins push the liquid in the upper wall of the tube, preventing the inception of dry-out and promoting high heat transfer coefficients. This result is particularly interesting in the case of R410A, since the plain tube data exhibit a behaviour like that of many CO2 data sets reported in the literature. As it can be seen in Fig. 6 (right), at low vapour quality the heat transfer coefficient decreases for both tubes, probably due to the nucleate boiling suppression associated with the increase of vapour quality. At x ¼ 0.6, the heat transfer coefficient of the microfin tube exhibits a local minimum and then increases, while the heat transfer coefficient of the plain tube continues to decrease because of the partial dry-out. The increase of the heat transfer coefficient, observed in the microfin tube, can be explained with the centrifugal effect resulting from the combination of the high fluid velocity and the helical shape of micro-fins. Fig. 7 reports the influence of heat flux: in this graph the experimental heat transfer coefficient is plotted as a function of vapour quality for R134a at around 30
C saturation temperature,

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A. Padovan et al. / Applied Thermal Engineering 31 (2011) 3814e3826

24

24 Ts 30.0, G 200, q 42.4 Ts 29.7, G 400, q 43.2 Ts 29.7, G 600, q 42.4

22 20

18

HTC [ kW m K ]

HTC [ kW m K ]

20

R134a

18 16 14 12 10 8

16 14 12 10 8

6

6

4

4

2

2

0

0

R410A

22

0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9

0

1

Ts 30.3, G 200, q 44.1 Ts 30.2, G 400, q 44.2 Ts 30.1, G 600, q 44.2

0

0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9

Vapour quality

1

Vapour quality

Fig. 5. Experimental heat transfer coefficient vs. vapour quality for R134a (graph on the left) and R410A (graph on the right) at around 200, 400 and 600 kg m2 s1 mass velocity. The average values of saturation temperature, Ts, mass velocity, G, and heat flux, q, are reported for each data set in the legend of the graphs.

200 kg m2 s1 mass velocity and heat flux from 14.9 to 42.4 kW m2. At low vapour quality, the heat transfer coefficient increases with heat flux, revealing a significant effect of nucleate boiling. When vapour quality increases the three sets of heat transfer coefficient tend to merge, showing that the role of nucleate boiling decreases when increasing vapour quality. When x ¼ 0.8 the heat transfer coefficients at q ¼ 14.9 kW m2 and q ¼ 22.3 kW m2 intersect. At q ¼ 42.4 kW m2 the profile of heat transfer coefficient is more flat and the heat transfer coefficient increase starts at higher vapour quality. The dry-out appears first for the data sets taken at higher value of heat flux. From the graph of Fig. 8, the effect of heat flux can be discussed in the case of R410A, when comparing data sets at around 30
C saturation temperature and different values of heat flux (q ¼ 21.4 kW m2 and q ¼ 44.2 kW m2). The heat transfer coefficient increases with heat flux even at high vapour quality, displaying a considerably influence of nucleate boiling heat transfer in the whole vapour quality range. Moreover, at higher heat flux,

corresponding to greater nucleate boiling contribution, the heat transfer coefficient exhibits a minimum. The effect of saturation temperature can be also observed in Fig. 8 by comparing data measured at Ts ¼ 30.2
C and q ¼ 21.4 kW m2 with data taken at Ts ¼ 39.9
C and q ¼ 22.1 kW m2. At low vapour quality, higher heat transfer coefficients occur at Ts ¼ 39.9
C: since the two data sets differ only in the value of saturation temperature, the difference between heat transfer coefficients can be explained by the rise of nucleate boiling due to the higher reduced pressure (Gorenflo et al. [19]). As vapour quality increases, the difference between the heat transfer coefficients measured at Ts ¼ 30.2
C and Ts ¼ 39.9
C decreases and at around 0.8 vapour quality the two curves of heat transfer coefficient merge. Indeed the vapour density increases with saturation temperature, resulting in a lower vapour velocity and convective heat transfer at high vapour quality. The data set taken at Ts ¼ 39.9
C exhibits a local minimum of heat transfer coefficient, which is the result of the competition between nucleate and convective

18

20 Smooth: Ts 31.0, G 400, q 28.5

Microfin: Ts 39.9, G 400, q 21.1

16

R134a

HTC [ kW m K ]

14

HTC [ kW m K ]

Smooth: Ts 41.0, G 400, q 24.0

18

Microfin: Ts 30.3, G 400, q 28.5

16

12 10 8 6 4

14 12 10 8 6 4

2

R410A

2

0 0

0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9

Vapour quality

1

0 0

0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9

1

Vapour quality

Fig. 6. Comparison between experimental heat transfer coefficients in microfin and plain tubes for R134a (graph on the left) and R410A (graph on the right). The average values of saturation temperature, Ts, mass velocity, G, and heat flux, q, are reported for each data set in the legend of the graphs. Plain tube data are taken from Del Col [6].

A. Padovan et al. / Applied Thermal Engineering 31 (2011) 3814e3826

4. Prediction of heat transfer coefficient

20 Ts 30.5, G 200, q 14.9 Ts 30.5, G 200, q 22.3 Ts 30.0, G 200, q 42.4

18

R134a

HTC [ kW m K

]

16 14 12 10 8 6 4 2 0

3821

0

0.1

0.2

0.3

0.4

0.5

0.6

0.7

0.8

0.9

1

Vapour quality Fig. 7. Experimental heat transfer coefficient vs. vapour quality for R134a at 14.9, 22.3 and 42.4 kW m2 heat flux. The average values of saturation temperature, Ts, mass velocity, G, and heat flux, q, are reported for each data set in the legend of the graph.

boiling mechanisms. This minimum becomes evident when increasing saturation temperature, due to the higher nucleate boiling heat transfer at low vapour quality. A similar effect of the saturation temperature has been also observed in the case of R134a data. The decrease of heat transfer coefficient at low vapour quality was recently discussed by Silva Lima et al. [20] and Del Col [6] during vaporization of R134a inside a plain tube: data reported in those papers show that the decrease of heat transfer coefficient is more pronounced at high saturation temperature.

In this section the prediction accuracy of some heat transfer models for vaporization of halogenated refrigerants in microfin tubes is assessed against the present experimental database, without considering the data points referred to the dry-out region. They are the models by Cavallini et al. [10], Chamra and Mago [11], Hamilton et al. [12], Koyama et al. [13] and Thome et al. [14]. The Cavallini et al. [10] model is an extended version of the original model (Cavallini et al. [21]) to low values of mass velocity, where the capillary forces in the inter-fins regions play an important role. Thus, the new model covers three different heat transfer mechanisms: nucleate boiling (anb), convective evaporation (acv) and capillarity (acap). The equations used for the calculation of the heat transfer coefficient with the Cavallini et al. [10] model are reported in Table 4. The Chamra and Mago [11] model is a modification of the model by Cavallini et al. [21], where the constants have been obtained using a different database. In their work the set of empirical constants for pure fluids are used also with R410A, due to its low temperature glide. Similarly, R410A is treated as a pure fluid in the Hamilton et al. [12] correlation, being a quasi-azeotropic mixture. The graphs reported from Fig. 9 to Fig. 12 show the results of the comparison in terms of calculated versus experimental heat transfer coefficient for the models by Cavallini et al. [10], Chamra and Mago [11], Hamilton et al. [12] and Thome et al. [14] . Data points corresponding to R134a and R410A are reported in separate graphs. Different markers are used to identify the data as a function of mass velocity. No results are reported for the Koyama et al. [13] model, since it provides a strong overestimation of experimental data, beyond 30%, for both fluids. The Cavallini et al. [10] model (Fig. 9) underestimates the R134a heat transfer coefficients measured at the lowest value of mass velocity, G ¼ 80 kg m2 s1. As mass velocity increases the accuracy of the model improves. It underestimates some heat transfer coefficients at G ¼ 200 kg m2 s1 and G ¼ 400 kg m2 s1; these data points correspond to the peak of heat transfer coefficient experimentally observed at high vapour quality. In the case of R410A, the model maintains high accuracy, predicting all measured data within 30%.

24 Table 4 Cavallini et al. [10] heat transfer model for pure fluids.

R410A

22

HTC ¼ HTCnb þ HTCcv þ HTCcap HTCnb ¼ HTCCooper $S$F1 ðdÞ ½log10 ðpred Þ0:55 M 0:5 q0:67 HTCCooper ¼ 55p0:12 red B S ¼ A1 $Xtt A1 ¼ 1:36; B ¼ 0:36 for G> 100 kg m2 s1 A1 ¼ 1:36sinðbÞ; B ¼ 0:36ðG=100Þ4 for G  100 kg m2 s1

20

]

18

K

16

HTC [ kW m

14

Xtt ¼ ½ð1  xÞ=x0:9 ðrV =rL Þ0:5 ðmL =mV Þ0:1 ; Xtt ¼ 1 F1 ðdÞ ¼ ðd0 =dÞ0:38 d0 ¼ 0:01 m

12

for Xtt > 1

HTCcv ¼ ðlL =dÞNucv;smooth tube Rx2:14 ðBond$FrV Þ F2 ðdÞ$F3 ðGÞ t ¼ 0:15 for G < 500 Kg m2 s1 ; t ¼ 0:21 for  500 Kg m2 s1 Nucv;smooth tube ¼ NuLO $F t

10

1=3

NuLO ¼ 0:023ðReLO Þ0:8 PrL PrL ¼ mL $cp;L =lL ; ReLO ¼ G$d=mL

8

F ¼ ½ð1  xÞ þ 2:63xðrL =rV Þ0:5 0:8 Rx ¼ f½2hfin $ng ð1  sinðg=2ÞÞ=½p$d$cosðg=2Þ þ 1g=cosðbÞ Bond ¼ g$rL $hfin $p$d=ð8s$ng Þ FrV ¼ G2 =ðr2V $g$dÞ

6 4

Ts 30.1, G 400, q 44.2 Ts 30.2, G 400, q 21.4

2

Ts 39.9, G 400, q 22.1

0 0

0.1

0.2

0.3

0.4

0.5

0.6

0.7

0.8

0.9

1

Vapour quality Fig. 8. Experimental heat transfer coefficient vs. vapour quality for R410A. The average values of saturation temperature, Ts, mass velocity, G, and heat flux, q, are reported for each data set in the legend of the graph.

d0 ¼ 0:01 m F2 ðdÞ ¼ ðd0 =dÞ0:59 G0 ¼ 100 kg m2 s1 F3 ðGÞ ¼ ðG0 =GÞZ Z ¼ 0:36 for G > 100 kg m2 s1 ; Z ¼ 3

for G  100 kg m2 s1

HTCcap ¼ 0:332lL =hfin ½G$hLV sinðbÞ=q0:4326 FG FG ¼ 0 for G > 100 kg m2 s1 FG ¼ 1  ðG=G0 Þ3 for 50 kg m2 s1 < G  100 kg m2 s1 FG ¼ 1:75ðG=G0 Þ

for G  50 kg m2 s1

3822

A. Padovan et al. / Applied Thermal Engineering 31 (2011) 3814e3826

26

26 G 80

24

G 200

G 400

G 600

22

Cavallini et al.

20 18 16 14

G 200

24

HTC cal [ kW m K ]

HTC cal [ kW m K ]

22

G 100

+30 %

12 10 8

-30 %

G 400

G 600

Cavallini et al.

20 18 16 14

+30 %

12 10 8

-30 %

6

6 4

4

R134a

2

R410A

2 0

0 0

2

4

6

0

8 10 12 14 16 18 20 22 24 26

2

4

HTCexp [ kW m K ]

6

8 10 12 14 16 18 20 22 24 26

HTCexp [ kW m K ]

Fig. 9. Calculated vs. experimental heat transfer coefficient for R134a (graph on the left) and R410A (graph on the right). Calculated values are obtained by the Cavallini et al. [10] model. Average values of mass velocity for data points are reported in the legend of the graph.

The Chamra and Mago [11] model (Fig. 10) overestimates the R134a data taken at low mass velocity (G ¼ 80 kg m2 s1 and G ¼ 100 kg m2 s1). At G ¼ 200 kg m2 s1 and G ¼ 400 kg m2 s1 good agreement between calculated and measured heat transfer coefficients is shown. When mass velocity increases to 600 kg m2 s1, the model somewhat overestimates experimental heat transfer coefficients. The scattering of data points increases in the case of R410A: the model here tends to underestimate measured data at G ¼ 200 kg m2 s1, whereas it overestimates measured data as mass velocity increases. The Hamilton et al. [12] model (Fig. 11) severely overestimates the R134a experimental heat transfer coefficients at low mass velocity; the performance of the correlation improves when mass velocity increases. When predicting the R410A heat transfer coefficients, the correlation is less accurate, underestimating most experimental data beyond 30%. The Thome et al. [14] model (Fig. 12) predicts with high accuracy the R134a heat transfer coefficients at G ¼ 80 kg m2 s1 and G ¼ 100 kg m2 s1. At higher values of mass velocity, a high scattering of data points is observed. This is even more evident in

the case of the higher pressure fluid R410A, particularly at G ¼ 400 kg m2 s1 and G ¼ 600 kg m2 s1. Fig. 13 reports a summary on the overall performance of models. The values of the indicators MAD (mean absolute deviation) and MD (mean deviation) obtained in the prediction of the experimental database are graphically reported. The model by Cavallini et al. [10] exhibits a good prediction accuracy for both fluids. In comparison with Cavallini et al. [10], the Chamra and Mago [11] model provides a slightly better prediction of the R134a heat transfer coefficients, but a worse prediction in the case of R410A. The Hamilton et al. [12] correlation is strongly affected by the reduced pressure: the model is the most inaccurate among the methods reported here and the prediction accuracy degrades when passing from R134a (pred ¼ 0.19e0.27) to R410A (pred ¼ 0.39e0.49). The model by Thome et al. [14] provides good values of MAD and MD for R134a, although the distribution of data points reported in Fig. 12 suggests a systematic dependence of the deviations from the operating parameters, in particular vapour quality. Higher inaccuracy is found in the prediction of the R410A heat transfer coefficients.

26

26 G 80

24

G 200

G 400

G 600

20 18 16 +30 %

12 10 8

-30 %

6

G 600

20 18 16 14

+30 %

12 10 8

-30 %

6

4

R134a

2 0

G 400

Chamra and Mago

22

Chamra and Mago

14

G 200

24

HTC cal [ kW m K ]

HTC cal [ kW m K ]

22

G 100

0

2

4

6

8 10 12 14 16 18 20 22 24 26

HTCexp [ kW m K ]

4

R410A

2 0

0

2

4

6

8 10 12 14 16 18 20 22 24 26

HTCexp [ kW m K ]

Fig. 10. Calculated vs. experimental heat transfer coefficient for R134a (graph on the left) and R410A (graph on the right). Calculated values are obtained by the Chamra and Mago [11] model. Average values of mass velocity for data points are reported in the legend of the graph.

A. Padovan et al. / Applied Thermal Engineering 31 (2011) 3814e3826

26

26 G 80

24

G 100

G 200

G 400

G 600

HTC cal [ kW m K ]

18 16 14 12

-30 %

10 8

G 400

G 600

Hamilton et al.

22 +30 %

20

G 200

24

Hamilton et al.

22

HTC cal [ kW m K ]

3823

20

R410A

18 16 14 12

+30 %

10 8 6

6 4 2

-30 %

4

R134a

2 0

0 0

2

4

6

0

8 10 12 14 16 18 20 22 24 26

2

4

HTCexp [ kW m K ]

6

8 10 12 14 16 18 20 22 24 26

HTCexp [ kW m K ]

Fig. 11. Calculated vs. experimental heat transfer coefficient for R134a (graph on the left) and R410A (graph on the right). Calculated values are obtained by the Hamilton et al. [12] model. Average values of mass velocity for data points are reported in the legend of the graph.

5. Prediction of dry-out inception vapour quality

When Regime-G2 conditions take place, the dry-out inception vapour quality is calculated as:

In this section the experimental measurements of dry-out inception vapour quality are compared with the predictions given by the Mori et al. [15] correlation. This correlation was developed from an experimental database of R22 and R134a data taken at 5
C saturation temperature, mass velocity from 100 to 600 kg m2 s1 and heat flux between 5 and 50 kW m2. Mori et al. [15] classified the dry-out inception qualities into two characteristic regimes, named Regime-G1 and Regime-G2. For a given set of system parameters the dry-out inception vapour quality is determined as:


xdi;2 ¼ min xdi;2a ; xdi;2b

xdi


¼ min xdi;1 ; xdi;2

(8)

where xdi,1 and xdi,2 are the values of dry-out inception quality calculated for Regime-G1 and Regime-G2, respectively. Under Regime-G1 conditions, the dry-out inception quality is given by:

xdi;1 ¼ 0:92

(9)

xdi;2a ¼ 0:44Fr 0:04 Bo0:07 ;

(11)

Fr ¼ G2 =½g$d$rV ðrL  rV Þ

(12)

Bo ¼ q=ðG$hLV Þ

(13)

where G is the mass flow rate, q is the heat flux, d is the inside diameter, g is the gravitational acceleration, hLV is the latent heat of evaporation, rL and rV are the liquid and vapour density, respectively. In Fig. 14 the difference between calculated and measured dryout inception vapour quality is plotted as a function of mass

26 G 80

24

G 100

G 200

G 400

G 600

HTC cal [ kW m K ]

18 16 +30 %

12 10 8

-30 %

G 400

G 600

Thome et al.

22

20

14

G 200

24

Thome et al.

22

HTC cal [ kW m K ]

xdi;2b ¼ 0:63Fr0:02 Bo0:33

where Fr and Bo are the vapour Froude number and the boiling number, respectively. They are calculated as:

26

20

R410A

18 16 14 12

+30 %

10 8 6

6 4

R134a

2 0

(10)

0

2

4

6

8 10 12 14 16 18 20 22 24 26

HTCexp [ kW m K ]

-30 %

4 2 0

0

2

4

6

8 10 12 14 16 18 20 22 24 26

HTCexp [ kW m K ]

Fig. 12. Calculated vs. experimental heat transfer coefficient for R134a (graph on the left) and R410A (graph on the right). Calculated values are obtained by the Thome et al. [14] model. Average values of mass velocity for data points are reported in the legend of the graph.

3824

A. Padovan et al. / Applied Thermal Engineering 31 (2011) 3814e3826 50

45

R134a

40

R134a

40

R410A

30

R410A

35

20

30

10

MD [%]

MAD [%]

50

25 20

0 -10

15

-20

10

-30

5

-40

0

-50 Cavallini et al.

Chamra and Mago Hamilton et al.

Thome et al.

Cavallini et al.

Chamra and Mago Hamilton et al.

Thome et al.

Fig. 13. Values of MAD (mean absolute deviation) and MD (mean deviation) obtained by the comparison of heat transfer models against the present database.

0.5

0.5

Mori et al.

0.4

R134a

0.3

R410A 0.2

xdi,cal - xdi,exp

xdi,cal - xdi,exp

R134a

0.3

R410A

0.2 0.1 0.0 -0.1

0.1 0.0 -0.1

-0.2

-0.2

-0.3

-0.3

-0.4

-0.4

-0.5

New equation

0.4

0

100

200

300

400

500

Mass velocity [ kg m -2 s -1 ]

-0.5

0

100

200

300

400

500

Mass velocity [ kg m -2 s -1 ]

Fig. 14. Difference between calculated and experimental dry-out inception vapour quality vs. mass velocity. Calculated values are obtained with the Mori et al. [15] correlation (graph on the left) and Eq. (14) (graph on the right).

velocity. The graph on the left reports the values calculated by means of the Mori et al. [15] correlation. All experimental data are underestimated by Mori et al. [15] and the prediction accuracy is significantly affected by mass velocity. No particular dependence on heat flux has been observed, instead. Therefore, a modification of the Mori et al. [15] correlation, based on the present database, is proposed. The new correlation comes from Eq. (11) for Regime-G2a, where the exponent of Fr has been changed in order to describe the dependence on mass velocity:

xdi ¼ 0:57Fr 0:02 Bo0:07

(14)

This equation has been validated with data taken in the present microfin tube within the following operating ranges: 3 < Fr < 50, 3$104 < Bo < 12$104, 0.19 < pred < 0.49. The graph on the right in Fig. 14 reports the results of the comparison with Eq. (14): the agreement between calculated and experimental data is within 0.03 for both fluids. 6. Conclusions 1. New heat transfer data, measured during vaporization of R134a and R410A at 30 and 40
C saturation temperature and mass flux from 80 to 600 kg m2 s1 inside a horizontal microfin tube, are reported. The reduced pressure of the present database varies from 0.19 to 0.49.

2. At Ts ¼ 30
C nucleate boiling heat transfer is relevant, but convective evaporation also plays a role, as shown by the strong increase of the heat transfer coefficient occurring just before the inception of dry-out. When the saturation temperature is increased from 30 to 40
C, the heat transfer coefficient rises only at low vapour quality, due to the effect on the nucleate boiling mechanism. 3. The experimental heat transfer coefficient increases with mass flux up to G ¼ 200 kg m2 s1; when the mass flux is further increased, the heat transfer coefficient slightly decreases. A similar effect of mass velocity on the heat transfer coefficient was also reported by Schael and Kind [22] and Dang et al. [23] in the case of vaporization of CO2 inside microfin tubes. These results suggest that, during flow boiling at high reduced pressure, an increase of the mass flux may be harmful not only with respect to the pressure drop but also for the heat transfer coefficient. 4. In the microfin tube the heat transfer coefficient is noticeably higher than that in the plain tube and the dry-out inception is shifted towards higher values of vapour quality. This is particularly clear in the case of R410A, where the action of micro-fins prevents the entrainment of liquid in the vapour core and the partial dry-out. The partial dry-out, which is responsible for the decreasing of the heat transfer coefficient in the plain tube, often occurs during vaporization of CO2 inside smooth channels (Yun et al. [24]); therefore the use of microfin tubes can be useful in this case to improve the performance and reduce the size of the evaporator.

A. Padovan et al. / Applied Thermal Engineering 31 (2011) 3814e3826

5. The present experimental heat transfer coefficients have been compared with models available in the literature. The Cavallini et al. [10] model is resulted as the most accurate method for the prediction of heat transfer coefficient. In fact, good agreement has been obtained for both R134a (MAD ¼ 14.5% and MD ¼ 9.9%) and R410A (MAD ¼ 10.3% and MD ¼ þ4.6%), showing that its prediction accuracy is not affected by reduced pressure. Satisfactory agreement is also obtained with the model by Chamra and Mago [11]. 6. The dry-out vapour quality has exhibited a dependence on the heat flux, displaying earlier dry-out at higher heat flux. In most cases, at constant values of saturation temperature and heat flux, the onset of dry-out is shifted to higher vapour quality when the mass flux is increased. When predicting the dry-out inception vapour quality, the correlation by Mori et al. [15] underestimates the experimental values, particularly at low mass velocity. A modified equation, which is able to predict the dry-out quality in the present data within 0.03, is here proposed. Nomenclature

A A1, B Bo Bond cp d FG F1(d), Fr FrV G

g HTC m h hfin hLV M MAD MD Np ng Nu pred Pr q Q Qd Rx Re S T t u uA

heat transfer area of a smooth tube with the same diameter as the fin tip diameter of the microfin tube, m2 coefficients in Cavallini et al. [10] model boiling number Bond number specific heat capacity, J kg1 K1 inside diameter at the fin tip, m mass velocity function in Cavallini et al. [10] model F2(d), F3(G) functions in Cavallini et al. [10] model vapour Froude number in Mori et al. [15] correlation, Eq. (12) vapour Froude number in Cavallini et al. [10] model (Table 4) mass velocity based on the cross-sectional area of a smooth tube with the same diameter as the fin tip diameter of the microfin tube, kg m2 s1 gravitational acceleration, m s2 heat transfer coefficient, W m2 K1 (kW m2 K1 in Figures) mass flow rate, kg s1 specific enthalpy, J kg1 fin height, m latent heat of evaporation, J kg1 molar mass, kg kmol1 P mean absolute deviation, ¼ (1/Np)[ (jacalaexpj/aexp)] 100, [%] P mean deviation, ¼ (1/Np)[ (acalaexp)/aexp]100, [%] number of data points number of grooves Nusselt number reduced pressure Prandtl number heat flux, ¼ Q/A, W m2 (kW m2 in Figures) heat flow rate, W heat dissipation, W geometry enhancement factor Reynolds number suppression factor temperature, K (
C in Figures and Tables) coefficient in Cavallini et al. [10] model standard uncertainty type A uncertainty of the measured quantity

uB x Xtt Z

3825

type B uncertainty of the measured quantity vapour quality Martinelli parameter constant in Cavallini et al. [10] model

Greek symbols b spiral angle, rad g apex angle, rad D difference F two-phase flow multiplier l thermal conductivity, W m1 K1 m viscosity, Pa s r density, kg m3 s surface tension, N m1 Other subscripts cap capillary Cooper Cooper correlation [25], cv convective d dissipated di dry-out inception in inlet L liquid LO liquid phase with total flow nb nucleate boiling preh pre-heater pres pre-section r refrigerant red reduced s saturation sub subcooled V vapour w water wall wall

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