An experimental study of flow boiling characteristics of carbon dioxide in multiport mini channels

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1 Applied Thermal Engineering 24 (2004) An experimental study of flow boiling characteristics of carbon dioxide in multiport mini channels Xiulan Huai a, *, Shigeru Koyama b, T.S. Zhao c, Etsou Shinmura d, Kinoshita Hidehiko b, Morita Masaki b a Institute of Engineering Thermophysics, Chinese Academy of Sciences, Beijing , China b Institute for Materials Chemistry and Engineering, Kyushu University, Kasuga-koen 6-1, Kasuga-shi, Fukuoka , Japan c Department of Mechanical Engineering, The Hong Kong University of Science & Technology, Clear Water Bay, Kowloon, Hong Kong, China d Aluminum Technology Center, Showa Denko K.K. Oyama , Japan Received 30 July 2003; accepted 15 October 2003 Available online 5 January 2004 Abstract Due to its unique combination of ecological and personal safety, carbon dioxide (CO 2 ), a natural fluid, is becoming one of the most promising alternative refrigerants for air-conditioning and refrigeration systems. This paper presents an experimental study of boiling heat transfer and pressure drop of CO 2 flowing in a multi-port extruded aluminum test section, which had 10 circular channels, each with an inner diameter of 1.31 mm. CO 2 was heated by hot water flowing inside copper blocks that were attached at both sides of the test section. Temperatures at the outer surface of the test section were measured using 24 K-type thermocouples embedded in the upper and lower surfaces along the length. Local heat fluxes were measured using twelve heat flux sensors to obtain the local enthalpies, temperatures and heat transfer coefficients. Bulk mean temperatures of CO 2 at the inlet and outlet of the test section were measured using two K-type thermocouples. The measurements were performed for pressures ranging from 3.99 to 5.38 MPa, inlet temperatures of CO 2 from )3.08 to C, heat fluxes from 10.1 to 20.1 kw/m 2, mass velocities from to kg/m 2 s, and vapor quality from 0.0 to 1.0. The results indicate that pressure drop along the test section is very small, two-phase CO 2 flow exhibits a higher heat transfer coefficient than that of the single-phase liquid or vapor flow. It is also shown that the mass velocity and the applied heat flux have significant effects on flow boiling heat transfer characteristics. The measured heat transfer coefficients were compared with correlations reported in the literature and large discrepancies are observed. Ó 2003 Elsevier Ltd. All rights reserved. * Corresponding author. Tel.: ; fax: address: hxl@mail.etp.ac.cn (X. Huai) /$ - see front matter Ó 2003 Elsevier Ltd. All rights reserved. doi: /j.applthermaleng

2 1444 X. Huai et al. / Applied Thermal Engineering 24 (2004) Keywords: Carbon dioxide; Flow boiling; Heat transfer; Pressure drop; Mini channels 1. Introduction Although CFCs and HCFCs are banned owing to the potential global warming and ozone layer depleting characteristics, it is uncertain whether some alternative refrigerants such as HFCs and HFC blends (man-made substances) would cause other environmental problems. Therefore, it is preferable to use natural substances which are already present in nature and do not involve unknown hazards. Carbon dioxide (CO 2 ), a non-toxic and non-flammable natural fluid with a zero ODP (ozone depletion potential) and a negligible GWP (global warming potential) [1], will become one of the most promising refrigerants, and provide a safe, economical, and cost-effective natural solution. The promising applications of CO 2 refrigeration systems include automotive air conditioning, heat pumps, residential/commercial air conditioning and various refrigeration areas [2,3]. In traditional air conditioning systems, both the heat absorption and heat rejection processes undergo under subcritical conditions. In CO 2 air conditioning systems, however, due to its low critical temperature (31.1 C), the heat rejection process has to take place in a gas cooler that operates above the critical pressure (73.8 bar), although the heat absorption process remains undergoing under the subcritical condition. As such, the CO 2 air conditioning systems virtually operate in a so-called transcritical cycle (Fig. 1). In 1993, Lorentzen and Pettersen [4,5] theoretically studied the CO 2 transcritical cycle and experimentally investigated a CO 2 refrigeration system for automobile air conditioning. Since then, a number of related studies have been reported [1 3,6 12]. One of the major challenges for CO 2 refrigeration systems is to improve the performance of heat exchangers. For this reason, some researchers have studied flow boiling heat transfer of CO 2 in tubes. For example, Bredesen [6], Høgaard and Jensen [7] measured heat transfer coefficients of flow boiling of CO 2 in horizontal tubes with different inner diameters (10.06, 7.00 mm), and they found that measured heat transfer coefficients were about twice as high as the prediction by the previous correlations for CFCs and HCFCs. Heat transfer coefficients measured by Rieberer and Halozan [8] on flow boiling of CO 2 in a horizontal tube with 15.0 mm diameter were higher than those for CFCs and HCFCs. Pettersen and Aarlin [9] attributed the reason why the evaporation heat transfer coefficient of CO 2 is higher than that of HCFC22 and HFC134a to the fact that CO 2 has extremely small surface tension. It is worth mentioning that most of these early works on flow boiling of CO 2 were performed for large diameter tubes. Significant progresses have recently been made toward the study of heat transfer and flow characteristics of CO 2 in mini/micro tubes. For example, Pettersen et al. [10] measured overall heat transfer coefficients and pressure drops of CO 2 flow boiling in a multi-port extruded tube, which had 25 circular channels, each with an inner diameter of mm and with a length of mm. It was found that heat transfer coefficient dropped significantly from a certain vapor quality upwards, and the discrepancies between correlations and measured values were substantial. Hihara and Tanaka [11] tested the boiling heat transfer coefficient and pressure drop of CO 2 in a horizontal smooth tube with 1.0 mm diameter, experimental results were compared with correlations and large discrepancies were also observed both for heat transfer

3 X. Huai et al. / Applied Thermal Engineering 24 (2004) Nomenclature A TS contact area between a heat flux sensor and the test section [m 2 ] Bo boiling number C P isobaric specific heat [J/kg K] C p mean specific heat C p ¼ H r H w [J/kg K] T r T w D i inner diameter of the tube [m] D bubble departure diameter [m] D be equilibrium break-off diameter [m] F two-phase flow coefficient FA the area of a single heat flux sensor [m 2 ] FL the length of a single heat flux sensor [m] F r surface roughness factor in flow boiling f ðiþ friction coefficient G mass velocity [kg/m 2 s] Gr Grashof number Gr ¼ ðq w q r Þq r gd 3 i H r enthalpy of CO 2 [J/kg] m mass flow rate [kg/s] Nu Nusselt number P pressure [MPa] Pr Prandtl number DP total pressure drop [kpa] DP F friction pressure drop [kpa] DP a accelerating pressure drop [kpa] q heat flux [W/m 2 ] QðiÞ local heat transfer rate for upper and lower surface [W] Re Reynolds number R P surface roughness in flow boiling [lm] K heat-flux-fraction factor q nb =q S suppression factor T temperature [ C] x vapor quality X tt Lockhart MartinelliÕs parameter ZðiÞ distance between two measuring points [m] Greek symbols a heat transfer coefficient [W/m 2 K] d wall thickness of the test tube [m] k thermal conductivity [W/m K] lðiþ dynamic viscosity [kg/m s] qðiþ density [kg/m 3 ] l 2 r

4 1446 X. Huai et al. / Applied Thermal Engineering 24 (2004) U L U v r modified Lockhart-MartinelliÕs parameter two-phase multiplier factor surface tension Subscripts in inlet of the test section out outlet of the test section r CO 2 wi inner wall wo outer wall L liquid v vapor p T 4 =Const Fig. 1. Transcritical CO 2 -cycle. h coefficients and pressure drops. Koyama and Kuwahara [12] also conducted experiments of flow boiling of CO 2 in a small tube with 1.8 mm inner diameter and mm effective heating length. The above literature review indicates that although some studies of flow boiling of CO 2 in tubes have been reported, the problem is far from being well understood because very complicated phenomena are involved in the boiling process. Therefore, further experimental and theoretical investigations are essential. This paper reports on the studies of both local and the overall boiling and pressure drop characteristics of carbon dioxide flowing in mini channels. In addition, the effects of some important parameters on heat transfer and pressure drop behaviors are also examined. 2. Experimental apparatus and procedures The test loop used in this work is schematically shown in Fig. 2. It consisted of a CO 2 pump, a CO 2 accumulator, a relief valve, a flow meter, a pre-heater, an electrical heater, a test section, a cooler, a sub-cooler, a CO 2 liquid receiver, three mixing chambers, two differential pressure transducers, four absolute pressure transducers, and a couple of valves. Four absolute pressure transducers (PGR 200KA, the measuring error is less than ±0.05%) were used to measure the

5 X. Huai et al. / Applied Thermal Engineering 24 (2004) Fig. 2. Schematic diagram of the experimental loop. static pressure at both the inlet and outlet mixing chambers of the electrical heater, and at both the inlet and outlet mixing chambers of the test section. Two differential pressure transducers (the accuracy is ±0.25% for large gage DP15-4-N-3-S-4-A) were installed at both the ends and the middle of the test section to measure the pressure drops. Experiments were carried out using a test section with multi-port extruded aluminum channels as shown in Fig. 3. The effective length of the test section is mm, with the width and thickness of the test section 20.0 and 2.0 mm, respectively. The equivalent hydraulic channel diameter is 1.31 mm. Carbon dioxide was heated with hot water flowing inside the copper block that was attached on the outer side of the test section. Twelve heat flux sensors (75 mm 19 mm 0.4 mm with measuring error less than ±5%) were located between the copper block and the test section at the upper and lower surfaces to measure the local heat fluxes. The temperatures at the outer surface of the test section were measured using 24 K-type thermocouples embedded in the upper and lower surfaces along the length. Bulk mean temperatures of CO 2 at the inlet and outlet of the test section were measured using 2 K-type thermocouples, and the local bulk mean temperatures of CO 2 along the flow direction were determined based on the energy balance with the measured wall heat flux. The mass flow rate of CO 2 was measured using a mass flow rate meter (CX003H SS 200R).

6 1448 X. Huai et al. / Applied Thermal Engineering 24 (2004) Inlet port Outlet port A 20 A A A Fig. 3. Schematic diagram of the test section. 3. Data reduction The following procedures were performed to obtain the local enthalpy, the temperature, the pressure drop of CO 2, and the heat transfer coefficient. The calculation model is shown in Fig. 4 and the details are presented in Fig. 5. The flow boiling experiments were performed for two different cases. For the first case, liquid CO 2 was fed at the inlet of the test section and boiling began somewhere downstream. For the second case, liquid vapor two-phase mixture was fed at the inlet of the test section Enthalpy (H r ) The local enthalpy of CO 2 H r (i) is determined from the energy balance over a unit length of the test section: Heat flux sensor Inlet mixing chamber T P P T heating Outlet mixing chamber P(0) Tr(0) Hr(0) P(1) Tr(1) Hr(1) P(6) Tr(6) Hr(6) P(15) Tr(15) Hr(15) Pr(16) Tr(16) Hr(16) Fig. 4. Schematic diagram illustrating data reduction.

7 X. Huai et al. / Applied Thermal Engineering 24 (2004) Flow direction Z(i) z (i+1) Starting/Ending point Z i Z i+1. Z BSP. Z i+2 Measuring point H r ði þ 1Þ ¼H r ðiþþ QðiÞ m ZðiÞ FL where QðiÞ represents the measured local heat transfer rate at the wall, m is the mass flow rate of CO 2, FL is the length of a single heat flux sensor, ZðiÞ is the distance between the two measuring points Pressure and temperature (P, T r ) The procedures for determining the local pressures and temperatures are described below: In single-phase region (a) Assuming the temperature and pressure at the measuring point i þ 1, one can determine the corresponding dynamic viscosity lðiþ and density qðiþ of CO 2 using the PROPATH (a program package for thermophysical properties of fluids, version 10.2 by propath group) lðiþ; qðiþ ¼F (b) Reynolds number Q AVE (i) Q AVE (i+1) i i + 1 i + 2 P(i) Tr(i) Hr(i) P(i+1) Tr(i+1) Hr(i+1) known unknown unknown PðiÞþPðiþ1Þ ; T rðiþþt r ði þ 1Þ 2 2 P(i+2) Tr(i+2) Hr(i+2) Xr(i) Xr(i+1) Xr(i+2) Fig. 5. Control volumes for determining local parameters. ð1þ ð2þ ReðiÞ ¼ GD i lðiþ where G presents the mass velocity [kg/m 2 s], and D i is inner diameter of the tube [m]. (c) Friction coefficient Laminar ðre < 2300Þ f ðiþ ¼ 16 ReðiÞ ð3þ ð4aþ

8 1450 X. Huai et al. / Applied Thermal Engineering 24 (2004) Turbulent ðre > 2300Þ f ðiþ ¼ 0:079 ReðiÞ 0:25 (d) Pressure drop and the absolute pressure at point (i þ 1) DP ¼ 4f ðiþ ZðiÞ D i G 2 2qðiÞ Pði þ 1Þ ¼PðiÞ DP ð4bþ ð5þ ð6þ (e) Temperature at point i þ 1 can be obtained by PROPATH based on the pressure Pði þ 1Þ and enthalpy H r ði þ 1Þ T r ði þ 1Þ ¼F ðpði þ 1Þ; H r ði þ 1ÞÞ (f) Repeat the above steps until the following condition is satisfied: DDP ¼ DP new DP old DP < 10 8 new ð7þ ð8þ When the enthalpy of CO 2 is higher than that of saturated liquid CO 2, H L, but smaller than that of saturated vapor CO 2, H v, i.e.: H L < H r ði þ 1Þ < H v liquid-vapor two-phase flow of CO 2 occurs In two-phase region (a) Assuming the temperature and pressure at the measuring point i þ 1, one can determine the corresponding dynamic viscosity for saturated liquid and vapor, as well as vapor quality using the PROPATH. ðpðiþþpðiþ1þþ l L ðiþ; q L ðiþ ¼F ð9þ 2 PðiÞþPðiþ1Þ l v ðiþ; q v ðiþ ¼F ð10þ 2 Vapor quality T r ðiþþt r ði þ 1Þ x r ðiþ ¼F ; H rðiþþh r ði þ 1Þ 2 2 (b) Reynolds number ð11þ ReðiÞ ¼ Gð1 x rðiþþd i l L ðiþ ð12þ (c) Friction factor can also be obtained by (4a) and (4b) (d) Pressure drop and the absolute pressure at point (i þ 1)

9 X. Huai et al. / Applied Thermal Engineering 24 (2004) DP Frictional pressure drop gradient,, is obtained from the following correlation DZ F DP ¼ U 2 DP L ð13þ DZ F DZ L where U L is a modified Lockhart MartinelliÕs parameter based on the correlation proposed by Mishima Hibiki [13] for two-phase flow in a horizontal tube with small diameter, DP DZ L Gð1 xþ ¼ 2f ½ D i q L U 2 L ¼ 1 þ C X tt þ 1 X 2 tt Š2 C ¼ 21 ð1 e 333D i Þ ð16þ Lockhart MartinelliÕs parameter X tt ¼ 1 x 0:875 0:5 0:125 q v l L ð17þ x q L l v ð14þ ð15þ then the friction pressure drop can be obtained from DP F ¼ DP DZ DZ F Accelerating pressure drop can be determined by DP a ¼ G2 q ðx out x in Þ L 1 q L q v The total pressure drop DP for two-phase flow DP ¼ DP F þ DP a Pressure in point (i þ 1) Pði þ 1Þ ¼PðiÞ DP Temperature at point i þ 1 can be obtained by PROPATH based on the pressure Pði þ 1Þ T r ði þ 1Þ ¼F ðpði þ 1ÞÞ Repeat the above steps until the condition (8) is satisfied. ð18þ ð19þ ð20þ ð21þ ð22þ 3.3. Wall temperature The inner wall temperatures T wi can be determined based on the measured outer wall temperatures T wo and the heat flux q wo ðiþ: T wi ðiþ ¼T wo ðiþþ d k q woðiþfa 1 ð23þ A TS

10 1452 X. Huai et al. / Applied Thermal Engineering 24 (2004) where T wo is the outer wall temperature, d is the wall thickness of the test tube, k is the thermal conductivity, FA is the area of a heat flux sensor, and A TS is the contact area between a single heat flux sensor and the corresponding test section Local heat transfer coefficient The local heat flux at the inner wall, q wi ðiþ, can be obtained based on the corresponding measured value at the outer wall, q wo ðiþ. q wi ðiþ ¼ q woðiþfa 2 ð24þ pd i N FL where N is the channels number. Then the local heat transfer coefficient is defined as: q wi ðiþ aðiþ ¼ T wi ðiþ T r ðiþ where T r ðiþ is the bulk temperature of CO 2. ð25þ 4. Results and discussion Experiments were conducted by systematically varying the system pressure, the heat flux, the inlet temperature, and the mass velocity of CO 2. The parameters that were varied and their ranges are shown in Table Pressure, temperatures, heat flux, heat transfer coefficient and vapor quality Fig. 6 shows the variations in local pressure P, inner wall temperature T wi, bulk mean temperature of CO 2 T r heat flux at the inner wall q, heat transfer coefficient a and vapor quality x along the test section, as well as measured inlet and outlet temperature of CO 2 T rmes. The experimental data were collected under the condition of P ¼ 3:99 MPa, G ¼ 310:0 kg/m 2 s, T in ¼ 3:29 C and q ¼ 10:1 kw/m 2. It can be seen that the liquid CO 2 entered the test section and boiling began at about m away from the inlet of the test section. In the liquid phase region, the temperature of CO 2 increased with the axial distance, whereas in the two-phase region, it remained almost constant. As expected, the heat flux and the heat transfer coefficient in the twophase region were higher than those in the liquid phase region. It is also noted from Fig. 6 that vapor quality increased with the axial distance in the two-phase region. Pressure is seen to decrease very slowly with the axial distance, meaning that the pressure drop was rather small along Table 1 Tested parameters and their ranges Pressure P/(MPa) Temperature of CO 2 T r /( C) Heat flux q/(kw/m 2 ) Mass velocity G/(kg/m 2 s) Vapor quality x )3.08 to

11 X. Huai et al. / Applied Thermal Engineering 24 (2004) Temperature ( o C) Tr Trmes Twi P Pressure (MPa) (a) Distance (m) Heat flux / Heat transfer coefficient q [W /m 2 ] α [W /m 2 K] x Vapor quality (b) Distance (m) Fig. 6. (a) Variations in pressure and temperature along the test section for the case when liquid was fed at the inlet; (b) Variations in heat flux, heat transfer coefficient and vapor quality along the test section for the case when liquid was fed at the inlet, P ¼ 3:99 MPa, G ¼ 310:0 kg/m 2 s, T in ¼ 5:16 C, q ¼ 10:1 kw/m 2. the test section. Fig. 6 also indicates that the measured temperatures T rmes at both the inlet and outlet of the test section agree well with the values T r calculated using the local energy balance method, indicating that the calculated temperature of CO 2 is accurate. It is also noted that the inner wall temperature and the heat transfer coefficient exhibit certain fluctuations in the twophase region, which may be attributed to the fact that the heated copper blocks were not perfectly intimate contact with the test section during the experiments. Fig. 7 shows the variations in local pressure P, inner wall temperature T wi, the bulk mean temperature of CO 2 T r, heat flux at the inner wall q, heat transfer coefficient a and vapor quality x along the test section, as well as the measured inlet and outlet temperature of CO 2 T rmes. The corresponding operating condition is: P ¼ 4:58 MPa, G ¼ 283:0 kg/m 2 s, T in ¼ 11:49 C and q ¼ 11:1 kw/m 2. The main difference between Figs. 6 and 7 is that the CO 2 is two-phase at the inlet of test section (0 < x < 1), boiling ended at about m away from the inlet of test section, and it changed into vapor single-phase with increasing the temperature of CO 2. When vapor

12 1454 X. Huai et al. / Applied Thermal Engineering 24 (2004) Temperature ( o C) Tr Trmes Twi P Pressure (MPa) (a) Distance (m) Heat flux / heat transfer coefficient (b) q [W /m 2 ] α [W /m 2 K] x Distance (m) Vapor quality Fig. 7. (a) Variations in pressure and temperature along the test section for the case when two-phase mixture was fed at the inlet; (b) Variations in heat flux, heat transfer coefficient and vapor quality along the test section for the case when two-phase mixture was fed at the inlet, P ¼ 4:58 MPa, G ¼ 283 kg/m 2 s, T in ¼ 17:8 C, q ¼ 11:1 kw/m 2. quality x exceeded 0.923, the heat transfer coefficient decreased and the wall temperature increased sharply as the increase of vapor quality, which indicates that the dryout happens in the channels. It is also found that the heat flux and the heat transfer coefficient in the two-phase region are higher than that of in the vapor single-phase region, pressure drop is very small along the test section, the measured temperature T rmes agree well with the calculated temperature T r at the inlet and outlet of the test section The effect of mass velocity on heat transfer characteristics Fig. 8 illustrates local the inner wall temperature T wi, the bulk mean temperature of CO 2 T r, and heat transfer coefficient a along the test section for different mass velocities when pressure and heat flux were kept almost the same. As can be seen, the mass velocity has a significant effect on the heat transfer characteristics, the larger the mass velocity, the further the onset saturated

13 X. Huai et al. / Applied Thermal Engineering 24 (2004) T wi 6 Temperature ( o C) T r P=4.0M Pa, q=10.2k W/m 2-2 G =200.0 kg/ms 2 G =389.0 kg/ms 2 (a) Distance(m) 9000 Heat transfer coefficient ( W/m 2 K) P=4.0MPa, q=10.2 kw/m 2 G=200.0 kg/ms 2 G=389.0 kg/ms 2 (b) Distance (m) Fig. 8. (a) Effect of mass velocity on temperature for the case when liquid was fed at the inlet; (b) Effect of mass velocity on heat transfer coefficient for the case when liquid was fed at the inlet, P ¼ 4:0 MPa, q ¼ 10:15 kw/m 2, G ¼ 200:0, kg/m 2 s. boiling away from the inlet of the test section because more thermal energy is needed to heat up the liquid CO 2. As shown in the Fig. 8, CO 2 is in liquid phase at the inlet and the early part of the test section, subcooled boiling occurs when the inner wall temperature exceeds the corresponding saturated temperature of CO 2 while the bulk mean temperature of CO 2 is lower than the saturated temperature, and then saturated boiling occurs as both the inner wall and CO 2 temperatures exceed the saturated temperature. Obviously, the effects of mass velocity on heat transfer characteristics in liquid phase and subcooled boiling regions are larger than that in the saturated boiling region. With the increase of the mass velocity, the temperatures of wall and CO 2 decrease in the liquid single-phase and subcooled boiling regions, the heat transfer coefficient increases in the liquid single-phase and saturated boiling regions, while the heat transfer coefficient decreases in the subcooled boiling region. The reason why the heat transfer coefficient decreases in the

14 1456 X. Huai et al. / Applied Thermal Engineering 24 (2004) subcooled boiling region is attributed to the fact that the increase of the mass velocity suppresses the subcooled boiling The effect of heat flux on heat transfer characteristics Fig. 9 shows the local inner wall temperature T wi, the bulk mean temperature of CO 2 T r, and the heat transfer coefficient a along the test section for different heat fluxes when pressure and mass velocity were kept almost the same. It can be seen that the higher the heat flux, the higher the temperatures of inner wall and CO 2 in the liquid phase region, the higher the heat transfer coefficient both in the liquid and two-phase regions, and the nearer the onset boiling point away 10 8 T wi 6 Temperature ( o C) T r q= 15.1 kw/m 2 q= 20.1 kw/m 2-2 (a) Distance (m) Heat transfer coefficient (W/m 2 K) q=15.1 k W /m 2 q=20.1 k W /m 2 (b) Distance (m) Fig. 9. (a) Effect of heat flux on temperature for the case when liquid was fed at the inlet; (b) Effect of heat flux on heat transfer coefficient for the case when liquid was fed at the inlet, P ¼ 4:0 MPa, G ¼ 399:0, kg/m 2 s, q ¼ 15:1, 20.1 kw/m 2.

15 X. Huai et al. / Applied Thermal Engineering 24 (2004) from the inlet of the test section. Therefore, increasing the heat flux properly is helpful for improving the flow boiling heat transfer characteristics Flow pattern Since the flow pattern was not observed in the present experiment, the flow patterns in this work were identified using the modified BakerÕs map [14]. The results are shown in Fig. 10, where n and w are defined as follows: 1=2 q n ¼ v ql ð26þ q air q water w ¼ r water r " # 2 l 1=2 L qwater l water q ð27þ L with the subscripts air and water denoting the physical properties of air and water at the standard condition (temperature 20 C and pressure 1 atm), r water is the surface tension between water and air. It is found that within the experimental range of this work the slug flow is a dominant flow pattern in the channels and a few measuring point are in the range of the transition region of the slug-annular, wavy-annular and wavy flow G (kg/m 2 s) ANNULAR [kg/m 2 s] Gx/ζ 10 1 WAVY BUBBLE OR FROTH STRATIFIED SLUG (1-x)ζψ/x Fig. 10. Flow pattern of experiments.

16 1458 X. Huai et al. / Applied Thermal Engineering 24 (2004) Comparison between experimental data and correlations The measured local heat transfer coefficients in the two-phase region are compared with the predictions by four previous correlations, YoshidaÕ83, YoshidaÕ94, Yu and Tanaka et al. It is seen from Fig. 11 large discrepancies exist between the present experimental data and the previous predictions, and most of the previous correlations overestimate the experimental data as shown in Fig. 11. The main reason is probably because the flow characteristics of the present experiment differ from that of the correlations. As shown in the Fig. 10, the present experimental points are mainly in the range of slug flow, whereas all the correlations were mainly developed based on the experimental data for annular flow. Another reason leading to the discrepancies between the present experimental data and the previous correlations is attributed to the fact that the experimental conditions are different. For example the correlations by Yu and Yoshida are suitable for low pressure and large diameter tubes with nucleate boiling being predominant. However, in this Fig. 11. Comparison between the present experimental data and previous correlations for heat transfer coefficients.

17 X. Huai et al. / Applied Thermal Engineering 24 (2004) work the pressure is high and tube size is small. The two-phase flow coefficient F and nucleate boiling suppression factor S in the correlations maybe need to be modified for the present high pressure. Therefore, both the pressure and the diameter have significant effects on heat transfer characteristics of CO 2 flow boiling, and they should be considered carefully during practical applications. 5. Conclusion The pressure drop and heat transfer characteristics of flow boiling CO 2 in horizontal mini channels were investigated experimentally. A aluminum test section (500 mm long, 20 mm wide and 2 mm thick), which had 10 multi-port extruded circular channels and each with an inner diameter of 1.31 mm, was tested with pressures ranged from 3.99 to 5.38 MPa, inlet temperatures of CO 2 from )3.08 to C, heat fluxes from 10.1 to 20.1 kw/m 2, mass velocity from to kg/m 2 s, and vapor quality from 0 to 1. The experimental results indicate that heat transfer coefficient in the two-phase region is higher than that of in the single-phase region. It is also found that once dryout occurs, the wall temperature increases but the heat transfer coefficient decreases sharply. Both the mass velocity and the heat flux have significant effects on flow boiling heat transfer characteristics. Both a larger mass flux and a higher heat flux lead to a larger heat transfer coefficient. In addition, heat transfer correlations reported in the literature deviate significantly from the experimental data collected in this work. The results of this work are of significance for the design and optimization of carbon dioxide refrigeration systems. Appendix A. Experimental correlations in the literature Yoshida (1983) [15] " # 2:0 0:44 a ¼ 3:7 Bo 10 4 þ 0:23ðBo 10 4 Þ 0:69 1 ða:1þ a LO X tt a LO ¼ 0:023Re 0:8 k L L Pr0:4 L D i ða:2þ Re L ¼ Gð1 xþd i l L ða:3þ X tt ¼ 1 x 0:9 0:5 0:1 q v l L ða:4þ x Bo ¼ q Gh fg q L l v ða:5þ (P ¼ 0:29 0:59 MPa, D i ¼ 7:0, 10.2 mm, G ¼ kg/m 2 s, q ¼ 5 35 kw/m 2, refrigerant: R22)

18 1460 X. Huai et al. / Applied Thermal Engineering 24 (2004) Yoshida (1994) [16] a ¼ F a LO þ Sa b a LO ¼ 0:023Re 0:8 k L L Pr0:4 L D i Re L ¼ Gð1 xþd i l L F ¼ 1 þ 2X 0:88 tt ða:6þ ða:7þ ða:8þ ða:9þ X tt ¼ 1 x 0:9 0:5 0:1 q v l L ða:10þ x q L l v S ¼ Bo ¼ 1 þ 0:9 Re L F 1: q Gh fg 1 0:5ð104 BoÞ 0:5 Xtt 0:5 ða:11þ ða:12þ a b ¼ 207 k 0:745 0:581 L qd b q v Pr 0:533 L D be k L T sat q L sffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 2r D be ¼ 0:51 gðq L q v Þ ða:13þ ða:14þ (P ¼ 0:29 1:1 MPa, D i ¼ 11:0 mm, G ¼ kg/m 2 s, q ¼ 5:0 46:5 kw/m 2, Refrigerants: HCFC22, CFC114, HCFC123) Yu [17] Total heat transfer coefficient is the sum of the forced convective component and the nucleate boiling component. i.e. a ¼ a cv þ a nb ða:15þ forced convection a cv ¼ 0:023Re 0:8 k L tp Pr0:4 L ða:16þ D i Re tp ¼ F 1:25 Re LO Re LO ¼ Gð1 xþd i l L F ¼ 1 þ 2X 0:88 tt ða:17þ ða:18þ ða:19þ

19 X tt ¼ 1 x 0:9 0:5 0:1 q v l L ða:20þ x q L l v nucleate boiling a nb ¼ K 0:745 Sa pb K 0:745 ¼ X. Huai et al. / Applied Thermal Engineering 24 (2004) ða:21þ 1 1 þ 0:875g þ 0:518g 2 0:159g 3 þ 0:7907g 4 ða:22þ g ¼ a cv Sa pb ða:23þ S ¼ 1 n ð1 e n Þ ða:24þ n ¼ D ba cv k L ða:25þ D b ¼ C 1 1:25 0:5 q L C PL T sat 2r ða:26þ q v h fg gðq L q v Þ a pb ¼ C k L D be sffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 2r D be ¼ 0:51 gðq L q v Þ qd be 0:745 q v k L T sat q L 0:581 Pr 0:533 L F r ða:27þ ða:28þ F r ¼ð8R P Þ ð0:2 0:2PrÞ ða:29þ where C 1 is a constant related to the bubble departure diameter C 1 ¼ 0:00005; C 2 is a constant determined by the experimental data to adjust the pool boiling correlation, C 2 ¼ 1:25. R b is roughness factor R P ¼ 1:0 lm P r ¼ P P c sffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 2r D be ¼ 0:51 gðq L q v Þ ða:30þ ða:31þ (P ¼ 0:25 1:12 MPa, D i ¼ 8:4 mm, G ¼ kg/m 2 s, refrigerants: HFC134a, HCFC123, HCFC22) Tanaka et al. [18] Tanaka et al. modified the two-phase flow coefficient F and nucleate boiling suppression factor S of YoshidaÕ94.

20 1462 X. Huai et al. / Applied Thermal Engineering 24 (2004) F ¼ 1 þ X 1:2 tt ða:32þ 1 S ¼ h i ða:33þ ðre 1 þ K L 10 4 Þ 2 1 ðbo10 4 Þ K 2 when D i ¼ 1, 2 mm, K 1 ¼ 0:5, K 2 ¼ 0; when D i ¼ 0:7 mm, K 1 ¼ 15, K 2 ¼ 4(D i ¼ 0:7, 1, 2 mm, G ¼ 360, 720, 1440 kg/m 2 s, refrigerants: CO 2 ). References [1] H. Hirao, H. Mizukami, M. Takeuchi, M. Taniguchi, Development of Air conditioning System Using CO 2 for Automobile, in: Proceedings of 4th IIR-Gustav Lorentzen Conference on Natural Working Fluids, Purdue University, 2000, pp [2] J. Pettersen, R. Aapline, Progress in CO 2 vapor compression system, Oji International Seminar, Trondheim, Norway, 1997, pp [3] Hermann Halozan, Rene Rieberer, CO 2 as Refrigerant-Possible Applications, in: Proceeding of 4th IIR-Gustav Lorentzen Conference on Natural Working Fluids, Purdue University, 2000, pp [4] G. Lorentzen, J. Pettersen, A new, efficient and environmentally benign system for car air conditioning, International Journal of Refrigeration 16 (1) (1993) [5] J. Pettersen, Recent Advances in CO 2 Refrigeration, in: Proceedings of the 19th International Congress of Refrigeration, The Hague, The Netherlands, 1995, pp [6] A.M. Bredesen, K. Aflekt, J. Pettersen, A. Hafner, P. Neksa, G. Skaugen, Studies on CO 2 Heat Exchangers and Heat Transfer, in: Proceeding Meeting of International Institute of Refrigeration on CO 2 Technologies in Heat Pumps and Air Conditioning, Trondheim, Norway, 1997, pp [7] H.J. Høgaard Kundsen, P.H. Jensen, Heat Transfer Coefficient for Boiling Carbon Dioxide, in: Proceeding Meeting of International Institute of Refrigeration on CO 2 Technologies in Heat Pumps and Air Conditioning, Trondheim, Norway, 1997, pp [8] R. Rieberer, H. Halozan, CO 2 Heat Pump Water Heater: Simulation and Test Results, in: Proceedings of the 1998 International Refrigeration Conference at Purdue, 1998, pp [9] J. Pettersen, R. Aarlin, Progress in CO 2 vapor compression system, Thermal Science & Engineering 6 (1) (1998) [10] J. Pettersen, R. Rieberer, S.T. Munkejord, Heat Transfer and Pressure Drop Characteristics of Evaporating Carbon Dioxide in Microchannel Tubes, in: Proceedings of 4th IIR-Gustav Lorentzen Conference on Natural Working Fluids at Purdue, 2000, pp [11] E. Hihara, S. Tanaka, Boiling Heat Transfer of Carbon Dioxide in Horizontal Tubes, in: Proceedings of 4th IIR- Gustav Lorentzen Conference on Natural Working Fluids at Purdue, 2000, pp [12] S. Koyama, K. Kuwahara, E. Shinura, et al. Experimental study on flow boiling of carbon dioxide in a horizontal small diameter tube, in: Conference of the International Institute of refrigeration Commission B1, October 3 5, Paderborn, Germany, 2001, pp [13] K. Mishima, T. Hibiki, Effect of inner diameter on some characteristics of air water two-phase flows in capillary tubues, Trans. Of JSME(Series B) 61 (589) (1995) (in Japanese). [14] D.S. Scott, Properties of co-current gas liquid flow, in: Advances in Chemical Engineering, vol. 4, Academic Press, 1963, pp [15] S. Yoshida, K. Nishikawa, T. Matsunaga, H. Nakata, Heat transfer to refrigeratant in horizontal tubes of evaporator, Refrigeration 58 (666) (1983) 3 10 (in Japanese). [16] S. Yoshida, H. Mori, H. Hong, T. Matsunaga, Prediction of heat transfer coefficient for refrigerant flowing in horizontal evaporation tubes, Trans of the JAR 11 (1) (1994) (in Japanese).

21 X. Huai et al. / Applied Thermal Engineering 24 (2004) [17] J. Yu, Experimental study on flow Boiling of Pure and Mixed Refrigeration in Smooth and Microfin Tubes, Department of Thermal Energy system Interdisciplinary Graduate School of Engineering Science Kyushu University, Japan, 12, 1994, pp [18] S. Tanaka, H. Daiguji, F. Takemuka, E. Hihara, Boiling Heat Transfer of Carbon Dioxide in Horizontal Tubes, in: Proceeding of 38th National Heat Transfer Symposium of Japan, Saitama, Japan, 5, 2001, pp