Flow Boiling Heat Transfer of Refrigerants R134a and

Experimental Heat Transfer A Journal of Thermal Energy Generation, Transport, Storage, and Conversion

ISSN: 0891-6152 (Print) 1521-0480 (Online) Journal homepage: http://www.tandfonline.com/loi/ueht20

Flow Boiling Heat Transfer of Refrigerants R134a and R245fa in a Horizontal Micro-Channel R. Ali , B. Palm & M. H. Maqbool To cite this article: R. Ali , B. Palm & M. H. Maqbool (2012) Flow Boiling Heat Transfer of Refrigerants R134a and R245fa in a Horizontal Micro-Channel, Experimental Heat Transfer, 25:3, 181-196, DOI: 10.1080/08916152.2011.609962 To link to this article: http://dx.doi.org/10.1080/08916152.2011.609962

Published online: 27 Jun 2012.

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Date: 25 January 2016, At: 06:39

Experimental Heat Transfer, 25:181–196, 2012 Copyright © Taylor & Francis Group, LLC ISSN: 0891-6152 print/1521-0480 online DOI: 10.1080/08916152.2011.609962

FLOW BOILING HEAT TRANSFER OF REFRIGERANTS R134A AND R245FA IN A HORIZONTAL MICRO-CHANNEL R. Ali,1 B. Palm,1 and M. H. Maqbool 1 1 Division

Downloaded by [Kungliga Tekniska Hogskola] at 06:39 25 January 2016

of Applied Thermodynamics and Refrigeration, Department of Energy Technology, Royal Institute of Technology, Stockholm, Sweden

Micro-channel-based evaporators are a promising option for high heat flux cooling applications. Micro-channels offer several advantages, including a smaller coolant inventory, superior heat transfer performance, compactness, lightness of weigh. Despite being attractive, the governing phenomena in micro-channels, especially during phase change, are less understood. This article reports the experimental flow boiling heat transfer results of refrigerants R134a and R245fa in a horizontal micro-channel. A series of experiments was conducted to measure the heat transfer coefficients in a circular micro-channel made of fused silica having an internal diameter of 781 m and a uniformly heated length of 191 mm. The outer surface of the test tube was coated with a thin, electrically conductive layer of indium-tin-oxide. The surface coating with the electrically conductive layer of indium-tin-oxide made it possible to visualize the flow boiling process simultaneously with uniform heating of the test section. R134a and R245fa were used as working fluids and experiments were performed at a system pressure of 7.7 bar for R134a and at 1.8 bar for R245fa, corresponding to saturation temperature of 30ı C. Mass flux was varied from 175 kg/m2 s to 500 kg/m2s, and heat flux ranged from 5 kW/m2 to 60 kW/m2. A highspeed camera was used to capture the images in the case of flow boiling of R134a. The experimental results indicated that the heat transfer coefficient increased with heat flux while the mass flux proved to have a negligible effect on heat transfer coefficient. Keywords

two-phase, visualization, heat transfer, micro-evaporator

INTRODUCTION The micro-channel heat sink with flow boiling has emerged as a competitive and potential option for high heat flux cooling applications. Micro-channels offer several advantages, such as compactness, lower inventory of fluid, lightness of weight, and reduced cost (due to reduced material and charge requirements). The ever-increasing environmental concerns can also be addressed by using micro-channel heat exchange devices, thereby reducing the refrigerant charge in domestic and industrial refrigeration and air-conditioning applications. Advances in fuel-cell technology also require the design of compact and enhanced micro heat exchange devices. Received 7 April 2011; accepted 14 December 2011. This paper was originally presented at the Second European Conference on Microfluidics held in Toulouse, France, December 2010. Address correspondence to Björn Palm, Division of Applied Thermodynamics and Refrigeration, Department of Energy Technology, Royal Institute of Technology, Brinellvägen 68, SE 100 44 Stockholm, Sweden. E-mail: [email protected] 181

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NOMENCLATURE A Ac Cp D f G h I ifg k L M m P P Q q 00 Rp T V xth z

heat transfer area (m2 ) cross-sectional area (m2 ) specific heat (J/kg K) diameter (m) friction factor (—) mass flux (kg/m2s) heat transfer coefficient (kW/m2 K) current (A) latent heat of vaporization (J/kg) thermal conductivity (W/mK) length (m) conduction number (—) mass flow of refrigerant (kg/s) pressure (bar) power (W) heat flux (W/m2 ) mean roughness (m) temperature (ı C) voltage (V) thermodynamic vapor quality (—) length, axial position (m)

Subscripts exp experimental e exit f fluid g gas hs heated section i inside in inlet l liquid loc local o outer pred predicted s surface sat saturation tp two-phase w wall z length, axial position (m)

Greek P

Letters pressure drop (mbar) dynamic viscosity (Ns/m2 ) density (kg/m3 )

Pr

Dimensionless Numbers Bo Nu

Re

q 00 Gilg .hD/ Nusselt number, k .Cp / Prandtl number, k Reynolds number GD l

boiling number

In recent years, tremendous efforts have been made to understand the basic phenomena involved during the boiling of fluids in micro-channels. Heat transfer, pressure drop, and critical heat flux (CHF) are essential parameters to be predicted as accurately as possible for the optimum design of micro heat exchange devices. Despite the great efforts made by researchers to understand the basic phenomena involved in flow boiling heat transfer and dryout mechanisms, discrepancies still exist. It is therefore essentially required to conduct more flow boiling experiments focusing on flow pattern based studies in order to provide optimum design tools for micro heat exchange devices. Ong and Thome [1] experimentally investigated the flow boiling heat transfer characteristics in 1.030-mm micro-channel employing R134a, R236fa, and R245fa as working fluids. Experiments were performed at a saturation temperature of 31ıC and at different inlet sub-cooling degrees ranging from 2 to 9 K. The mass flux was varied from 200 to 1600 kg/m2 s and heat flux from 2.3 to 250 kW/m2 . The heat transfer coefficient was observed to increase with heat flux at low vapor qualities, and the convective heat transfer mechanism dominated in the annular flow regime at higher vapor qualities. At low vapor qualities, the heat transfer coefficient of R134a was the highest, followed by R236fa and R245fa. A transition from a confined bubble to anannular flow regime was observed to occur at lower vapor qualities with increasing mass flux. Shiferaw et al. [2] experimentally investigated the heat transfer characteristics of R134a in a 1.1-mm stainless-steel micro-channel. The operating parameters were mass


FLOW BOILING OF REFRIGERANTS IN A MICRO-CHANNEL

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flux 100 to 600 kg/m2 s, heat flux 16 to 150 kW/m2 , and system pressure 6 to 12 bar. The experimental results indicated an increase in heat transfer coefficient with heat flux and system pressure up to vapor qualities of about 30% to 50%; beyond that range of vapor fractions, the heat transfer coefficient decreased and was not dependent on heat flux. The results were compared with the three-zone evaporation model presented by Thome et al. [3]. At relatively low pressures, the model predicted their results fairly well, while at higher vapor fractions and in the dryout region, the model significantly over-predicted the results. Lee and Lee [4] investigated flow boiling heat transfer in small horizontal rectangular channels using R113 as the working fluid. The rectangular channels had the dimensions of 0.4 to 2 mm height and a fixed width of 20 mm. Experimental parameters ranged with mass flux from 50 to 200 kg/m2 s, heat flux up to 15 kW/m2 , and vapor quality from 0.15 to 0.75. Their experimental results indicated that the heat transfer coefficient was only weakly dependent on heat flux but increased with mass flux and vapor quality, and hence, the authors concluded that the convective mechanism was the dominant heat transfer mode during the experiments. For the smaller size and lower flow rate conditions, the authors concluded that the thin film evaporation was the controlling mechanism in their experiments. Bortolin et al. [5] performed flow boiling experiments in a single circular channel of 0.96-mm diameter using R245fa as a working fluid at a saturation temperature of 31ıC. Heat flux ranged from 5 to 85 kW/m2 , mass flux was varied from 200 to 400 kg/m2 s, and vapor quality changed from 5% to 80% during the experiments. The heat transfer coefficient was observed to increase with heat flux, and the effect of mass flux on the heat transfer coefficient was less important, as it was observed to decrease with increasing vapor quality. Comparison of experimental data was made with Lazarek and Black’s [6] correlation and Thome et al.’s [3] showing under-prediction of data by both models. Madrid et al. [7] performed flow boiling experiments in rectangular vertical minichannels using HFE-7100 as the working fluid. Test section consisted of 40 rectangular multi-channels each of 0.840-mm hydraulic diameter and 220-mm heated length. The authors observed that heat transfer coefficient was not dependent on heat flux, but mass flux and vapor quality were important parameters influencing the heat transfer. Kuznetsov and Shamirzaev [8] studied the flow boiling heat transfer coefficients of R21 in a rectangular mini-channel of 1:6 6:3 mm and using R21 as working fluid. They performed experiments at two mass fluxes of 50 and 215 kg/m2 s. The authors concluded that thin liquid film evaporation contributed to heat transfer, and the convective boiling was the dominant mechanism, especially for highest mass flux. They used modified forms of Cooper, Liu and Winterton, and Kandlikar and Balasubramanian correlations to correlate their experimental data. The data was predicted well by the Kandlikar and Balasubramanian [9] and Liu and Winterton [10] correlations when the flow regime was convective-regime dominant. At vapor qualities of less than 0.5, the modified form of Cooper correlation was also in good agreement with their experimental data. More recently, Schilder et al. [11] experimentally investigated the heat transfer and pressure drop characteristics of ethanol and water in a 600-m glass tube coated with a thin, transparent, electrically conductive layer of indium-tin-oxide (ITO). Experimental results indicated that the two-phase Nusselt number was independent of heat flux and increased with mass flux. Flow visualization of ethanol was also performed. Based on

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Table 1. Comparison of thermo-physical properties of R245fa and R134a at Tsat 30ı C

Refrigerant


R134a R245fa Percent difference with R134a

Pressure (bar)

Liquid density (kg/m3)

Vapor density (kg/m3)

7.7 1.79 78

1,187 1,325 11.6

37.60 10.16 73

Liquid viscosity (Pa-s)

Surface tension (mN/m)

Liquid specific heat (J/kgK)

Latent heat of vaporization (kJ/kg)

Liquid Prandtl number (—)

Liquid thermal conductivity (W/mK)

183 380 107.7

7.40 13.4 81

1,447 1,350 6.7

173 188 8.7

3.35 5.80 73

0.0789 0.0883 11.9

the images captured, the authors identified the thin film evaporation, formed around an elongated bubble, as the dominant heat transfer mechanism. In the present study, experiments have been performed with a single, circular, horizontal micro-channel with an internal diameter of 0.781 mm and a uniformly heated length of 191 mm using refrigerants R-134a and R245fa as working fluids. R134a and R245fa have significantly different thermo-physical properties, as shown in Table 1, where the R134a is taken as the reference fluid and properties of R245fa are shown as a percentage difference with reference to R134a. When developing correlations for design purposes of compact heat exchangers, it is of utmost practical importance to use fluids having different thermo-physical properties. The heat transfer studies for R245fa in micro-channels are very limited in literature; therefore, it will be of wide interest to explore and compare the two-phase heat transfer characteristics of the two refrigerants in order to be able to develop design tools for their applications in compact heat exchangers. Heat transfer and pressure drop in the phase change process are dependent on flow regimes existing in the channel. As recommended by Celata [12] and Thome [13], the models for predicting heat transfer and pressure drop need to take into account the respective flow patterns, and flow-regime-based studies can greatly improve the understanding of the hydrodynamic and thermal transport phenomena. However, the visualization-based studies of the boiling process are very limited in the literature. Therefore, the data for R134a was recorded with simultaneous visualization of the boiling process using a high-speed camera in order to elucidate the underlying phenomena in the process.

EXPERIMENTAL SET-UP AND CALCULATION PROCEDURES Experimental Set-Up and Instrumentation A schematic diagram of the test rig is shown in Figure 1. The design of the facility was such that the system pressure, heat flux, mass flux, and inlet sub-cooling degree could be adjusted independently. The fluid was driven in the flow loop by a magnetic gear pump with microprocessor control, type ISMATEC MCP-Z standard, and was measured with a Coriolis-effect mass flow meter. The pump could be run in different modes and gave a wide variety of flow rates depending upon the type of head used. The evaporation temperature was defined by the system pressure, which was controlled by adjusting the flow rate of the cooling water through the condenser. Further fine control and maintenance



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Figure 1. Schematic diagram of experimental set-up and test section. (color figure available online)

of the system pressure was achieved by adjusting the liquid level in the condenser by controlling the electrical heat input to a refrigerant tank connected to the main loop. The sub-cooling degree at the inlet of the test section was adjusted by controlling the electrical heat input to the pre-heater located just before the test section. A 7-m filter was used in the loop to prevent any small particles from entering the test section. The test facility was instrumented with an absolute pressure transducer (Druck, USA, 20 bar) to measure the pressure at the inlet of the test section and a differential pressure transducer (Druck, USA, 350 mbar) to measure the pressure drop across the test section. Apart from the thermocouples attached to the test section, some stainless-steel sheathed thermocouples were inserted at different locations in the experimental set-up to measure the fluid bulk temperature. The entire test rig was properly insulated to minimize the heat losses. The test tube was made of fused silica with an internal diameter of 0.781 mm and a total length of 261 mm with 191 mm of heated length. The outer surface of the test tube was coated with a thin electrically conductive layer of ITO. The internal surface structure of the test tube was scanned to obtain the surface roughness information, which can be


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Figure 2. Inner surface characterization of glass test tube. (color figure available online)

seen in Figure 2. The details of inner surface roughness are shown in Table 2, where Ra is the arithmetic mean roughness, Rv is the maximum valley depth, and Rp is the maximum peak height. The inner surface characterization of the tube was performed using the method of conical stylus profilometry, and five independent profiles were obtained for surface characterization of the tube for a fixed sampling length. It is readily observed from the surface characterization results that the glass tube used in present study is quite smooth and will probably hold fewer active nucleation sites than a rougher surface, for example, a metallic tube, as nucleation sites are well known to form at defects at the surface. Seven thermocouples, which were calibrated using an ice bath (mixture of ice and water), were glued on the outer surface of the tube using an electrically insulating and thermally conductive epoxy. Swage lock brass connections were used at the two ends of the test section to connect it to the main loop. Two holes were drilled in the brass connections at the inlet and outlet of the test section as pressure taps to measure the differential pressure as shown in Figure 1. Heat Losses The test section was enclosed in another glass tube having an outer diameter of about 40 mm, and low pressure was provided in the annulus in an effort to reduce heat

Table 2. Inner surface characterization details of glass test tube Test tube Glass

Ra (m) 0.015

Rv (m) 0.036

Rp (m) 0.070



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losses, but it was found that this system was not able to eliminate the heat losses completely. Therefore, the convective heat losses to the ambient were estimated by correlating the applied power with the temperature difference between the tube wall and the ambient when there was no refrigerant flow. The power was increased in small steps, and the wall temperatures were noted once they became stable. The ambient temperature was controlled in the room and was the same in all cases. The heat losses so correlated were also compared with predictions by different correlations for natural convection taken from literature. The results obtained from natural convection correlations as recommended in Itoh et al. [14], Churchill and Chu [15], Kast and Klan [16], and Incropera and Dewitt [17] were used to calculate the heat losses to the ambient and were compared with experimentally obtained heat losses. Heat losses due to radiation heat exchange were also estimated, and results obtained for combined convection and radiation heat exchange were in close agreement (10–20% higher) with the experimental heat losses obtained by correlating the wall temperature and applied power without refrigerant flow. Moreover, the experimentally estimated heat losses are up to 10% for actual powers during the flow boiling test. Therefore, a 10% to 20% difference in heat losses in a single phase will give a 1% to 2% difference in actual power used for flow boiling experiments. In order to estimate the conduction heat losses in the axial direction, the conduction number as defined by Chiou [18] and Maranzana et al. [19] was used. This number gives the relative importance of wall conduction and convection in the fluid. Maranzana et al. [19] used a numerical model to develop a criterion, stating that if the conduction number M is less than 10 2 , then the conduction losses may be neglected. The criterion for a circular tube as defined by [18, 19] is given by Eq. (1): M D

ks kf

Do2 Di2 Di L

1 Re Pr

> 10 2 :

(1)

The conduction number based on the above equation for different Reynolds numbers was calculated for all the single-phase data points and was found to range from 0.000202 to 0.001226, which is very small compared to the above criteria. Therefore, the conduction losses were assumed to be negligible.

Experimental Method The system pressure, mass flow rate, and inlet temperature were first set to a desired level and the system was run without heating for a long enough time to obtain steady-state conditions. The electrical power was then applied and increased in small steps until boiling was initiated, which could be observed by an immediate drop in outer wall temperatures starting at the last thermocouple and also visually by looking at the glass test section. Each data point at a certain heat flux was recorded after steady-state conditions were reached. Each data point at a given condition was obtained by averaging more than 100 data points recorded using a data logger with a frequency of 0.3 Hz. The data logger was connected to a personal computer and a computer program written in HP VEE was used to display and obtain the real-time data in an Excel sheet for further processing. Thermodynamic properties for R-134a and R245fa, including density,

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enthalpy, viscosity, and thermal conductivity, were calculated with the computer code REFPROP 7.0 developed by National Institute of Standards and Technology (NIST). The use of a transparent tube together with an ITO coating allowed simultaneous visualization and electrical heating. Flow patterns were recorded using a high-speed camera at six positions along the tube. Images captured with the camera were transferred to a PC, and the required number of frames was saved on the computer.


Data Reduction The data reduction methods for both single and two phases are described in this section. The experimental single phase friction factor f was calculated using the Darcy-Weisbach equation from the pressure drop P measured during the experiments as f D P

2 D ; G2 L

(2)

where is the density, D is the inner diameter, G is the mass flux, and L is the total length. The test section was heated electrically using a DC power supply. The power Q in watts applied to the test section was calculated using the voltage and current across the test section: Q D I V;

(3)

where I and V are current and voltage in ampere and volts, respectively. For a given test point, the heat flux added to the test section q 00 was calculated as q 00 D

Q

Qloss ; A

(4)

where A is the heat transfer area, given as A D Dzhs . Qloss is heat loss to the ambient, which was calculated as described in the “Heat Losses” section. A Yokogawa WT130 power meter (Japan) was used to measure the voltage and current applied to the test section. Heat transfer coefficient h was calculated using heat flux and the local wall and fluid temperatures as hz D

q 00 Tw;z

Tf;z

;

(5)

where Tw;z is the inner wall temperature, and Tf;z is the fluid temperature. The inside wall temperature Tw;z was calculated from the measured outside surface temperature using the one-dimensional heat conduction equation for a cylinder [20]. For a single phase under sub-cooled conditions, the bulk temperature at any axial position was calculated from the inlet temperature and the heat added to the test section: Tf;z D Tf;in C z

q 00 D ; mC P p

where z is any axial location, Cp is the specific heat, and m P is the mass flow rate.

(6)


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In the case of boiling, Tf;z was the local saturation temperature. The local saturation temperature Tsat;z was determined by taking into account the drop in saturation temperature due to pressure drop, where pressure drop was assumed to vary linearly along the test section. Average heat transfer coefficients were calculated by arithmetically averaging the local heat transfer coefficients. The thermodynamic vapor quality at any axial position xth;z was calculated from the heat transferred to the fluid as


xth;z D

q 00 D.z zsat / ; Ac Gifg

(7)

where D is the internal diameter of the test tube, z is any axial location, zsat is the location where saturated conditions are reached, Ac is the cross-sectional area of the test tube, G is the mass flux, and ifg is the latent heat of vaporization. The axial position where saturated conditions are reached can be calculated as [21] zsat D

mC P p .Tsat Tin / ; q 00 D

(8)

where m P is the mass flow rate, Cp is the specific heat of the fluid, Tin is the inlet temperature at the test section, and D is the inner diameter of test section.

EXPERIMENTAL RESULTS AND DISCUSSION As part of validation of the experimental set-up and related instrumentation, singlephase pressure drop and heat transfer experiments were first performed prior to flow boiling tests. Besides the experimental set-up and instrumentation validation, the single-phase experiments also provide a fairly good estimate of the heat losses, thereby confirming the accuracy in two-phase heat transfer measurements.

Single-Phase Results The friction factor was calculated from the single-phase pressure drop obtained during the experiments. The friction factor so obtained was compared with that obtained from Hagen-Poiseuille (Eq. (9)) in the laminar regime and with Blasius (1913) Eq. (10)) in the turbulent regime: f D

64 ; Re

f D 0:3164 Re

(9) 0:25

:

(10)

The single-phase heat transfer coefficient expressed in the form of a dimensionless Nusselt number was compared to a constant value of 4.36 for fully developed laminar tube flow at constant heat flux and with the Gnielinski (1976) correlation and the Dittus-Boelter (1930) correlation for turbulent flow conditions. Experimental results for single-phase heat transfer of R245fa are in fairly good agreement with classical theory for single-phase flow,


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Figure 3. Single-phase heat transfer and pressure drop results for R245fa in 0.781-mm internal diameter micro-channel. (color figure available online)

as can be seen in Figure 3. Similar results for R134a were found, and the experiments were repeated with very good repeatability of results. Single-phase heat transfer and pressure drop results together with the rather good energy balance of the micro-evaporator was thought to confirm the accuracy of the measuring instruments. Visualization Observations for R134a Major flow patterns observed in the experiments were bubbly (including isolated and confined bubbly), elongated bubble, slug, semi-annular, and annular flow. The detailed results of visualization study can also be seen in Ali et al. [22]. Figure 4 shows images captured during boiling of R134a at a heat flux of 9.6 kW/m2 and a mass flux of

Figure 4. Visualization images at G D 185 kg/m2 s, q 00 D 8:9 kW/m2 , and Tsat D 30ı C. (color figure available online)



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185 kg/m2 s. The exit vapor quality for this condition is 0.239, and the local vapor fraction at the last thermocouple position is 0.221. The images were captured from the position of the tube where boiling initiated up to the exit of the section. The flow patterns changed from confined bubbly to elongated bubble, slug, and finally to semi annular flow. Soon after the departure of tiny bubbles from the nucleation site, the bubbles grew up to the size of the tube diameter. Further growth of these bubbles in radial direction was restricted by the tube diameter. Due to this restriction in growth, these bubbles expanded in the axial direction, and the flow pattern changed to elongated bubble at xth D 0:07. This type of flow is characterized by a thin liquid film between the vapor and the channel wall. Due to increase in vapor quality along the tube, the bubble velocity also increased, which ultimately helped the elongated bubbles to merge and form larger vapor slugs (slug flow regime). Farther downstream of the tube, the slug flow gave rise to wavy annular or annular flow. As the heat flux was increased further, the semi annular/annular flow regime transition shifted farther upstream of the tube. Based on the flow boiling visualization observations for all mass flux and heat flux conditions, the major findings are summarized in the following. Due to the small channel dimensions, the bubbly flow zone lasts for a very short length after the nucleation initiation point, as is also seen in Figure 4. Due to confinement of vapor bubbles, the flow pattern quickly changes to elongated bubble flow with a thin liquid film between the vapor and the channel wall. Further increase in heat flux shifts the nucleation initiation point toward upstream of the channel, and at higher heat fluxes, the nucleation was essentially restricted to the inlet of the channel. No nucleation was observed in the thin liquid film between the vapor and the channel wall in the elongated bubble regime, which shows that the nucleation was suppressed due to change in flow pattern and that thin film evaporation could be the main heat transfer mechanism in the elongated bubble regime. As a result of the confinement effects, the annular flow was reached at comparatively low vapor fractions of about 18% to 35%, depending on the mass flux. Increase in mass flux resulted in transition to annular flow at lower vapor quality. Average Heat Transfer Coefficient The average heat transfer coefficient as a function of exit vapor quality and heat flux for R134a at Tsat D 30ıC is shown in Figure 5. The average heat transfer coefficient increases with exit vapor quality xe and heat flux for all the mass flux conditions. For a given xe , a higher mass flux gives a higher average heat transfer coefficient. This is explained by the fact that for a given exit vapor quality condition; a higher mass flux will mean a higher heat flux. For a given mass flux, the higher heat flux gives a higher heat transfer coefficient, which can also be seen in this figure. A comparison of average heat transfer coefficient for R134a and R245fa at Tsat D 30ı C is shown in Figure 6. The same trend for R245fa is observed as in the case of R134a. The heat transfer coefficient increases with heat flux for a given mass flux condition. The difference in heat transfer coefficient between the two refrigerants is small. A comparison

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Figure 5. Average heat transfer coefficient for R134a, Tsat D 30ı C. (color figure available online)

shows that at low exit vapor fractions (up to about xe D 0:2), the heat transfer coefficient is the same for the two refrigerants for a given mass flux and vapor fraction. As the vapor fraction increases, a slightly larger increase in heat transfer coefficient is observed for R245fa as compared to R134a for a given condition of mass flux and exit vapor quality. In general, for current experimental results, it is obvious that the heat transfer coefficients are strongly dependent on heat flux, but are to a much lesser extent dependent on mass flux or vapor fraction. This type of behavior of heat transfer coefficient is generally observed where nucleate boiling is a dominant mechanism of heat transfer. However, from a previous visualization study [23], and also from a visualization study for the current tube [22], it is observed that this type of heat transfer behavior is present even in the cases where the nucleation of the bubbles is essentially restricted at close to the inlet of the tube (i.e., low vapor fractions), while no nucleation occurs in the rest of the heated length of the tube. Comparison with Correlations The plots in Figures 7 and 8 show the comparison with correlations for R134a and R245fa, respectively, where Figure 7a is the comparison with the Lazarek and Black [6] correlation, Figure 7b is the comparison with Tran et al.’s [24] correlation,

Figure 6. Comparison of average heat transfer coefficient for R134a and R245fa, Tsat D 30ı C, filled symbols for R134a and empty symbols for R245fa. (color figure available online)



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Figure 7. Comparison of R134a experimental data with correlations: (a) Lazarek and Black [6] (b) Tran et al. [24], (c) Cooper [25], and (d) Owhaib [26]. (color figure available online)

Figure 7c is the comparison with Cooper [25] pool boiling correlation, and Figure 7d is the comparison with the correlation by Owhaib [26]. The correlations by Cooper [25] and Lazarek and Black [6] are in closer agreement with R134a data, while the other correlations either over-predict or under-predict the experimental data. The R245fa data is under-predicted by all the correlations. The correlation by Lazarek and Black [6] is given as Nu D 30 Re0:857 Bo0:714: In order to see the effects of different thermodynamic and transport properties of the two refrigerants on the heat transfer coefficient, the above equation can be rearranged as follows: htp D

30G 0:143q 000:714kl : 0:714 Di0:147 0:857 ifg l

The liquid viscosity and latent heat of vaporization are 108% and 8% higher for R245fa than R134a at the same saturation temperature. The two quantities being higher for R245fa might be responsible for lower predicted heat transfer coefficients by the Lazarek and Black correlation. Similarly, the correlations by Tran et al. [24] and Owhaib [26]


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Figure 8. Comparison of R245fa experimental data with correlations: (a) Lazarek and Black [6], (b) Tran et al. [24], (c) Cooper [25], and (d) Owhaib [26]. (color figure available online)

use surface tension and liquid density, which are different; especially the surface tension is significantly higher in the case of R245fa, as seen in Table 1. The different thermophysical properties are thought to be the main reason for the discrepancy in prediction capability of different correlations for the two refrigerants at the same saturation temperature. CONCLUSIONS An experimental study focusing on flow boiling of R134a and R245fa in a horizontal micro-channel was presented. The test tube was made of fused silica with an internal diameter of 781 m and a heated length of 191 mm and total length of 261 mm. A high-speed camera with a close-up lens was used to visualize the boiling process of R134a. Average heat transfer coefficients were obtained, and the effects of refrigerant thermo-physical properties were elucidated. The following major conclusions from the current study can be drawn. 1. The average heat transfer coefficient increased with heat flux, while mass flux was observed to have a negligible effect on it. 2. The difference in heat transfer coefficient for the two refrigerants was smaller than expected, as the difference in thermo-physical properties is significant.


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Comparisons with correlations available in the literature were made. None of the correlations could predict all the experimental data satisfactorily for both refrigerants. However, reasonable accuracy was found in the case of R134a for the Cooper [25] and Lazarek and Black [6] correlations. In case of R245fa, all the correlations under-predicted the experimental data.


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