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Evaporation heat transfer and pressure drop of R161 in a 7 mm micro-fin tube
International Journal of Heat and Mass Transfer 62 (2013) 638–646
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Evaporation heat transfer and pressure drop of R161 in a 7 mm micro-fin tube Xiaohong Han, Peng Li, Zheng Wang, Xuehui Wang, Xuejun Zhang ⇑, Guangming Chen Institute of Refrigeration and Cryogenics, State Key Laboratory of Clean Energy Utilization, Zhejiang University, Hangzhou 310027, China
a r t i c l e
i n f o
Article history: Received 14 November 2012 Received in revised form 28 February 2013 Accepted 2 March 2013 Available online 11 April 2013 Keywords: R161 Evaporation heat transfer Micro-fin tube
a b s t r a c t The evaporation characteristics of refrigerant fluoroethane (R161) in a horizontal micro-fin tube with an outside diameter of 7 mm were investigated by experiment. The heat transfer coefficient and the pressure drop were measured when the heat flux ranged from 28.15 to 49.29 kW/m2, the mass flux ranged from 100 to 250 kg/m2 s, and the saturation temperatures ranged from 5 to 8 °C. According to the results, the heat transfer coefficient went up with the mass flux and the heat flux, and dropped with the saturation temperatures within the experimental conditions. The pressure drop grew with the mass flux and decreased with the saturation temperatures. By comparison, the heat transfer coefficient of R161 was about 23% higher than that of R22. In addition, the ratio of pressure drop (R22/R161) ranged from 0.88 to 1.65. The presented research is helpful in designing more compact and effective heat transfer exchangers for the air-conditioning systems using R161. Ó 2013 Elsevier Ltd. All rights reserved.
1. Introduction As providing comfortable living conditions for us, the air conditioning and heat pump technology exert some negative influences on the environment. The research of alternative refrigerants is recognized as one of the urgent issues to ease the ozone depletion and global warming. For the excellent environmental performance (Ozone Depletion Potential (ODP) = 0, Global Warming Potential (GWP) = 12), high energy efficiency, and good commonality with the existing systems [1], fluoroethane (R161) is another potential substitute of R22 in the small-scale air conditioning. So far, a lot of researches about R161 and the related refrigerant mixtures have been conducted, for example, PVTx (Pressure–Volume–Temperature-Composition), the saturated vapor pressure, the vapor–liquid equilibrium, the solubility with POE (Polyol Ester), cycle performance, etc. [2–7]. On the other hand, micro-fin tubes are widely used in the air-conditioning applications for their better heat transfer ability together with larger heat transfer area. Though inner fins will inevitably cause larger refrigerant pressure drop in comparison with the smooth tubes, the enhancement of heat transfer is found to be more significant [8]. When a new refrigerant is proposed, the heat transfer characteristics are usually regarded as one of the research focuses in order to analyze the performance of the evaporators and condensers in the actual air conditioning systems. For example, Dang et al. [9] studied the boiling heat transfer of HFO1234yf in a small diameter horizontal smooth tube. The results ⇑ Corresponding author. Tel./fax: +86 571 87952446. E-mail address: [email protected] (X. Zhang). 0017-9310/$ - see front matter Ó 2013 Elsevier Ltd. All rights reserved. http://dx.doi.org/10.1016/j.ijheatmasstransfer.2013.03.017
indicated that heat transfer was influenced by the mass flux more significantly at high vapor qualities, and for the low quality region, heat flux showed large effect. Greco and Vanoli [10] measured the boiling heat transfer coefficient for R22, R507, R134a, R404A and R410A in a smooth tube, the results revealed that R134a had the better heat transfer performance than the other refrigerants. Based on research about the convective heat transfer inside an annular tube, Zeng et al. [11] reported that the tube wall temperature and the Nusselt number’s fluctuation profile were mostly influenced by the Stefan number and the dimensionless phase change temperature range in the phase-change heat transfer process. Kim et al. [12] found that the evaporating heat transfer coefficient of R410A in both the smooth tubes and the micro-fin tubes increased with the mass flux and the heat flux. They also reported that at the same test conditions, heat transfer performance of micro-fin tubes was significantly superior to the smooth tubes, and the improvement degree was affected by the tube diameters. Yu et al. [13] investigated the evaporating heat transfer performance of R134a in a 10.7 mm (OD) smooth tube and a 10.7 mm (OD) micro-fin tube by the experiment. They found that for the annular and intermittent flows, the experimental heat transfer coefficient kept an agreement with the Gungor and Winterton correlation with corrected constants. Lallemand et al. [14] studied the boiling heat transfer performance of R407C and R22 in the smooth and micro-fin tubes, respectively. The results showed that the heat transfer coefficient of R407C was 15% and 35% lower than that of R22 in the smooth and micro-fin tubes, respectively. Torrella et al. [15] concluded that boiling heat transfer coefficient of R407C was influenced by the refrigerant flow rate and the evaporation temperature. Hu et al.
X. Han et al. / International Journal of Heat and Mass Transfer 62 (2013) 638–646
639
Nomenclature A D Di,av Di,max e eav G GWP h hi hav H i ilg I L Lx _ m N ODP PVTx POE P 4P Dpa Dpf Dpi
area (m2) diameter (mm) average inside diameter (mm) bottom inside diameter (mm) bottom wall thickness (mm) average wall thickness (mm) mass flux (kg/(m2 s)) Global Warming Potential heat transfer coefficient (W/m2 K) local heat transfer coefficient (W/m2 K) average heat transfer coefficient (W/m2 K) fin height (mm) enthalpy (kJ/kg) latent heat of vaporization (kJ/kg) current (A) length (m) the length from the inlet to the certain measurement point mass flow rate (kg/s) number of fins Ozone Depletion Potential Pressure–Volume–Temperature-Composition Polyol Ester pressure total pressure drop (kPa) acceleration pressure drop (kPa) frictional pressure drop (kPa) inlet pressure drop (kPa)
[16] studied the boiling heat transfer characteristics of two-phase R410A–oil mixture in a micro-fin tube. It was revealed that at lower vapor qualities (x < 0.4), heat transfer was enhanced by the presence of oil. While at the higher qualities (x > 0.65), the heat transfer coefficient dropped significantly as the oil concentration increased. In order to enrich the database of the potential alternative refrigerant and supply reference for designing more compact and effective evaporators for air conditioning systems using R161, evaporation heat transfer of R161 inside a horizontal micro-fin tube was investigated by experimental methods in this work. According to the results, effect of the saturation temperature, the mass flux and the heat flux on the heat transfer coefficient and the pressure drop were analyzed in detail. Moreover, the measured results of R161were compared with those of R22. 2. Experiment 2.1. Experimental apparatus The experimental setup used to study the evaporation heat transfer characteristics of R161 is shown in Fig. 1. It mainly includes an electric heating system, a data collection system and three main loops: the refrigerant loop, the cooling loop, and the pre-heating loop. The electric heating system provides the heat needed to evaporate the refrigerant in the test section, and it mainly contains a DC power supply and an array of heating band. The DC power supply can continuously output the voltage and the current at the range of 0–100 V and 0–30 A, respectively, with high stability and display precision (uncertainty: ±0.2%), and its maximum output power is 3.0 kW. The material of the heating band is silicone rubber and Ni–Cr wire with a temperature resistance of 250 °C, a insulation resistance of 50 MX and a maximum heat capacity of 280 W/m. The data collection system mainly includes an Agilent 34970A and a PC installed with the data acquisition program, the sample
Dpo q Q T U x xi Greek k b
c
outlet pressure drop (kPa) heat flux (kW/m2) total heat input (kW) temperature (°C) voltage (V) vapor quality (–) local vapor quality
thermal conductivity (W/m K) helix angle (°) fin angle (°)
Subscripts bottom the bottom of the tube ev evaporation in inlet i inside l liquid phase out outlet o outside r refrigerant side the side of the tube sat saturation top the top of the tube wo outside wall wi inside wall
frequency of Agilent 34970A is set at 12 time/min in this study. The refrigerant loop consists of a test tube (7 mm O.D., 2 m length), two sight glasses, a condenser, a receiver, a refrigerant pump, an accumulator, a filter/dryer, a Coriolis mass flow meter (CMF) and a pre-evaporator. The refrigerant pump (lift: 0.6 MPa, rated flow: 0.316 mL/r) is a magnetically gear pump with high conveying stability and excellent sealing. Furthermore, it is a variable-speed pump working together with a frequency converter (0–600 Hz) to transfer the liquid refrigerant and adjust the mass flux for the test section. At the outlet of the pump, an accumulator is accessed in order to alleviate the fluctuation in pressure and flow rate of the refrigerant loop. The mass flow meter (range: 0–2180 kg/h, uncertainty: ±0.1%, sample frequency: 60 time/min) records the refrigerant mass flow rate in real time with high precision and low pressure loss. Considering the necessary overhaul and preventing the possible impurity and lubricant oil from accumulating in the mass flow meter when the refrigerant loop is cleaned, a by-pass loop is installed in parallel with the flow meter. The pre-evaporator, installed before the test section, is used to evaporate the sub-cooled refrigerant liquid to the two-phase state. Actually, in order to obtain the local two-phase heat transfer coefficient of test refrigerant at the limited tube length and heating power, the inlet refrigerant is expected to be in the saturated state with a certain vapor quality and a specified saturation temperature. A thermostatic water tank with the capacity of 8.0 kW supplies the hot water which is used to heat the refrigerant liquid in the pre-evaporator. An electromagnetic flow meter (range: 0–3.5 m3/h, uncertainty: ±0.2%, sample frequency: 60 time/min) is installed in the pre-heating loop to record the flow rate of hot water passing through the pre-evaporator and the inlet/outlet temperatures are measured by thermocouples. Indeed, after the pre-evaporator, there still exists a calming section before the test section. The calming section is to guarantee the inlet refrigerant to be in the fully developed flow regime, and it has a total length
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Fig. 1. Schematic diagram of the experimental system.
of 45 cm. For the turbulent flow in this study, the calming section with a length of >60Di (38 cm) can fully satisfy the need for fully developed flow [17] . The condenser installed after the test section is used to condense the evaporated refrigerant coming from the test section to be in the sub-cooled state. The water–alcohol mixture (40%:60% in mass fraction) works as the cooling medium in another side of the condenser. 2.2. Test section The detailed information about the test section and measurement positions for temperatures and pressures is shown in Fig. 2. Table 1 and Fig. 3 give the geometric parameters of the test micro-fin tube. The tube was heated by the DC power supply through silicone rubber heating band glued on the tube outside surface. A thin layer of heat conduction silicone was coated between the heating band and the tube surface to spread the heat transferred to the refrigerant. The test section was insulated to the ambient with three layers of rubber foam, and the total thickness of the insulation is 24 mm with the thermal conduction ability of
no more the 1 W/m2 K. At twenty locations along the axial direction of the test tube (seen in Fig. 2), T-type thermocouples with the uncertainties of ±0.1 °C were used to measure the local outside surface temperature. And for each location, there were four thermocouples fixed on the top, bottom, right, left sides of the tested tube, respectively. To avoid being disturbed by the current, each thermocouple was separated with the tube outside surface by a layer of 0.03 mm thick Teflon sheet. A high-precision resistance thermometer detector (RTD) with the uncertainty of ±0.02 °C was used to calibrate all the thermocouples before the measurement. The layout of Teflon sheet, thermocouples, heat conduction silicone, heating band and rubber foam are shown in Fig. 4. Applying Fourier’s heat conduction law for the one dimensional steady state, the inside surface temperature used in calculating of local heat transfer coefficient was obtained from the measured outside surface temperatures. At the inlet and the outlet of the test section, one four-wire Pt100 RTD with the uncertainty of ±0.03 °C was inserted to the flow stream to directly measure the refrigerant temperature (Tin and Tout), respectively. At the same time, the corresponding refrigerant pressures were recorded by the pressure
Fig. 2. Schematic diagram of the test section.
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X. Han et al. / International Journal of Heat and Mass Transfer 62 (2013) 638–646 Table 1 Dimension of the test tube.
Table 2 Experimental conditions.
Test tube parameter
Micro-fin
Outside diameter, Do (mm) Tube length, L (mm) Bottom wall thickness, e (mm) Average wall thickness, eav (mm) Average inside diameter, Di,av (mm) Bottom inside diameter, Di,max (mm) Helix angle, b (°) Fin angle , c (°) Fin-height, H (mm) Number of fins, N Inside surface area, Ai (cm2/m)
7.0 2000 0.28 0.33 6.34 6.41 15 34 0.1 65 286.7
Refrigerant
R161 R22
Conditions Saturation temperature (°C)
Mass flux (kg/m2 s)
Heat flux (kW/m2)
5, 0, 8
100, 175, 250
28.15, 38.62, 49.29
R22 and R161, respectively, shown in Table 2. Furthermore, in order to obtain the heat transfer characteristics in a wide quality range of 0.1–1, the length of micro-fin was chosen for 2.0 m. Thus, a relative high outlet quality will be acquired and heat transfer performance at the fully developed flow regimes is accessible. 3. Data reduction The heat transfer coefficient was given by:
q¼
Fig. 3. Detailed information of the microfin tube.
Fig. 4. Cross-section view of the test tube.
transducers with the uncertainties of ±0.2% F.S. (full scale: 0–2.5 MPa), and the whole uncertainties were within 5 kPa. By comparison, the measured inlet and outlet refrigerant temperatures were consistent with those calculated from the measured saturation pressures within ±0.3 °C. A differential pressure transducer with high-precision (±0.1% F.S., the whole uncertainty is within 0.037 kPa), was used to monitor the pressure drop. T-type thermocouples were used to record the inlet and outlet temperatures of the cold and hot liquids in the cooling loop and the preheating loop. The measurement signals of all the thermocouples, RTD, pressures transducers, differential pressure transducer are collected by the Agilent data acquisition unit. When all the parameters (temperatures, pressures, flow rate, output voltage/current, etc.) of different positions have kept stable for more than one hour, it is believed that the steady state was reached, and the data acquisition starts. 2.3. Test conditions The tests were conducted for three saturation temperature points, three mass flux points and three heat flux points with
Q ¼ hev ðT wi T ev Þ Ai
ð1Þ
where hev is the evaporation heat transfer coefficient, Twi is the inside wall temperature of the test tube, and Tev is the refrigerant evaporation temperature. The heat flux q over the whole test section was supposed to be uniform for the negligible heat conduction along the axial direction of the tube and the excellent heat transfer between the tube surface and the heating band. Due to the good insulation for the test tube, heat losses were neglected and the heat flux was estimated from the total heat input (Q(Q = U I), U and I are the output voltage and current of the DC power supply, respectively). The inside wall area Ai was calculated according to the average inside diameter Di,av of present tube (Di,av was a equivalent diameter, and it was calculated according to the wetted cross-section area of the microfin tube, and the wetted cross-section area could be calculated based on the Fig. 3). Applying the heat conduction equation (one-dimensional, radial, and steady-state), the local inside wall temperature in Eq. (1), Twi, was calculated by:
T wi ¼ T wo
Q ln DDi;aov
ð2Þ
2pkL
where k and L are the material thermal conductivity and the length of the test tube, respectively. Two is the average value of measured temperatures at the top, the two sides and the bottom of the tube outside surface:
T wo ¼
T wo;top þ 2 T wo;side þ T wo;bottom 4
ð3Þ
Due to the pressure drop, the local refrigerant temperature, Tev gradually reduced from the tube inlet to the tube outlet [12], and it was interpolated with a linear equation [12]:
T eV ¼ T in þ ðT out T in Þ
Lx L
ð4Þ
where Tin and Tout are the measured inlet and outlet refrigerant temperatures, respectively. Lx is the length from the inlet to the certain measurement point. The saturation temperature Tsat is defined as the average value of Tin and Tout. The refrigerant vapor quality at every measurement point of different location was obtained based on the local refrigerant temperature:
x¼
i il ilg
ð5Þ
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where il is the specific enthalpy of the saturated liquid phase at the certain temperature, ilg is the latent heat of vaporization, i is the specific enthalpy of two-phase refrigerant, determined by:
i ¼ iin þ
q Ai _r m
ð6Þ
where iin is obtained from the sub-cooled liquid through an energy _ r is the refrigerant mass flow rate. balance in the pre-evaporator, m The average heat transfer coefficient was defined by:
P20 h i xi hav ¼ Pi¼1 20 i¼1 xi
ð7Þ
where xi and hi are the local refrigerant vapor quality and heat transfer coefficient at the certain measurement point of different position, respectively. The expression of the frictional pressure drop,Dpf, was:
Dpf ¼ Dp Dpi þ Dpo Dpa
ð8Þ
where Dp represents the tall pressure drop measured by the differential pressure transducer, Dpa is the acceleration pressure drop, Dpi and Dpo are inlet and outlet pressure losses, respectively. Estimation for Dpa, Dpi and Dpo was proposed by Collier [18]. The standard error analysis presented by Moffat [19] was introduced to estimate the uncertainties of the heat transfer coefficient and other calculated parameters, for instance the heat flux, mass flux, vapor quality and the frictional pressure drop, the results are shown in Table 3. Taking the heat transfer coefficient for example, the uncertainty of it is estimated as follows:
@hev ¼ hev
sffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 2 2 2 2 2 @U @I @Ai @T wi @T eV þ þ þ þ U I Ai T wi T ev T wi T eV ð9Þ
According to the natural uncertainties of the measurement equipments and the actual measured data during the experiment, the relative uncertainties of the voltage difference U and the current I are estimated at 0.27% and 0.39%, respectively, the relative uncertainty of the inside wall area is assumed at 9.2%. the inside wall temperature Twi and the refrigerant evaporation temperature Tev are assumed to be typical, 3.6 °C and 0 °C (absolute value), respectively, and their relative uncertainties with the temperature difference Twi Tev are estimated at 9.0% and 0.83%, respectively, so the uncertainty of heat transfer coefficient is 12.9%. The transport and physical properties of R22 and R161 were obtained by using REFPROP 7.0 [20].
Table 3 Estimated uncertainties. Parameter
Uncertainties
T Tev P 4P _r m U I G q x hev hav 4Pf
±0.1 °C ±0.03 °C ±0.2% ±0.1% ±0.1% ±0.2% ±0.2% ±1.2% ±1.2% ±2.2% ±12.9% ±15.5% ±9.4%
Note: T in this table refers to all the temperatures measured by the thermocouples.
4. Results and discussion The heat transfer coefficient and the pressure drop were obtained for the mass flux range of (100–250) kg/m2 s, the heat flux range of (28.15–49.29) kW/m2 and the evaporation temperature range of (5 to 8) °C. 4.1. Comparison with existing correlations and previous experimental data 4.1.1. Convergence verification by comparison between the experimental data of this work and results from existing correlations It was found that many correlations about the refrigerant boiling heat transfer have been proposed for the smooth tubes. Due to lack of the related database as well as the difficulties in correlating the complicated parameters, models for micro-fin tubes are limited in open literatures [21]. Two enhancement factors were introduced by Kandlikar and Rayoff [22] into a smooth tube correlation to predict the boiling heat transfer coefficient in micro-fin tube, but for different micro-fin tubes and different test refrigerants, the enhancement factors were specifically empirical constants. Cavallini et al. [23] proposed a micro-fin tube model based on their condensation correlation, including parameters liking the surface tension and the fin height. A semi-empirical model for estimating the evaporation heat transfer inside micro-fin tubes were developed by Chamra and Mago [24], which considered the nucleate boiling and the forced convention. Yun et al. [21] generalized a boiling correlation for horizontal micro-fin tubes based on a smooth tube model, and non-dimensional parameters were introduced to explain the heat transfer enhancement of micro-fin tubes in this correlation. The present heat transfer coefficient of R22 was compared with the results from the above mentioned correlations, shown in Fig. 5. It can be observed from Fig. 5 that the correlation of Kandlikar and Rayoff [22], and the correlation of Cavallini et al. [23] overpredicted the heat transfer coefficient by more than 35% and 46%, respectively, while the Chamra and Mago [24] correlation under-predicted the heat transfer coefficient by more than 30%. Within an error band of ±30%, the predicted values by Yun et al. [21] correlation cover 74% of the experimental data, which revealed a relatively reasonable agreement. The similar conclusions can be found in literature [21] that this correlation caused an average deviation of 11.7%, when compared with 1333 data points. The comparison with above-mentioned correlations suggested that our experimental data had a relatively good convergence at the certain conditions. 4.1.2. Comparison between the experimental data of this work and the cited data The evaporation heat transfer data of R22 in a micro-fin tube by Seo and Kim [25] was compared to those in this study, shown in Fig. 6. The vapor quality variations from the inlet to the outlet are about 0.2–0.8 and 0.1–1, respectively in the cited literature and in this work. From the compared results, most of results had the similar trend within the common mass flux ranges. At the mass fluxes of 100 and 250 kg/m2 s, the conclusions of the measured heat transfer coefficients at 0 °C and 5 °C were different from those of literature [25], because the measurement ranges were changed. When the conditions were changed, the temperature influence on the heat transfer might also be different. For example, the decrease of temperature lowers the density of the two-phase refrigerant, correspondingly the flow rate will be increased and the convective heat transfer is promoted, the explanation supported the results from literature [25]. However, with the mass flux increases, that phenomenon will be restrained, which might
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8000
10000
8000
6000
4000
Tsat= -5-8 ºC 2
G = 100-250 kg/m s 2000
2
q = 28.15-49.29 kW/m
2
G = 175 kg/m s 6000
4000
R161, Tsat= - 5 º C R161, Tsat= 0 º C R161, Tsat= 8 º C
2000
R22, Tsat= - 5 º C R22, Tsat= 0 º C R22, Tsat= 8 º C
0
0 0
2000
4000
6000
8000
10000
12000
2
Experimental heat transfer coefficient, W/m K Fig. 5. Comparison between the present heat transfer coefficient of R22 with those predicted from some correlations.
8000 Seo and Kim (2000) : Tsat= -15 º C, q = 5kW/m
2
2
Average heat transfer coefficient , W/m K
2
q = 38.62 kW/m
2
Local heat transfer coefficient, W/m K
Yun et al.(2002) Cavallini et al. (1999) Kandlikar-Rayoff (1997) Chamra-Mago (2006)
2
Predicted heat transfer coefficient , W/m K
12000
Seo and Kim (2000) : Tsat=
-5 º C,
2
q = 5kW/m
Seo and Kim (2000) : Tsat= 5 º C, q = 5kW/m
2
6000
4000
Present data: Tsat = -5 º C, q = 28.15kW/m
2
2000
Present data: Tsat =
0 ºC,
2
q = 28.15kW/m
Present data: Tsat = 8 º C, q = 28.15kW/m
2
0 100
200
300
2
Mass flux,kg/m s Fig. 6. Comparison between the present heat transfer coefficient of R22 with those obtained by Seo and Kim [16].
lead to the similar results with the literature [27] and our work. Thus, within the same measurement ranges, the comparison results between the cited literature and our data can support the reliability of experimental apparatus to some extent. 4.2. Evaporation heat transfer coefficient Fig. 7 showed the variation of local heat transfer coefficients of R22 and R161 with the vapor quality for various evaporation temperatures when the heat flux and mass flux are fixed at 38.62 kW/m2 and 175 kg/m2 s, respectively. It can be observed that the heat transfer coefficient of R161 grew with the vapor quality, and reaches a maximum at x = (0.7–0.8). The region corresponding to x > (0.7–0.8) was usually referred to the dry-out region. But R22 did not apparently reveal the phenomenon of dry-out in present study. Reasons were that, in the dry-out region of R161, the liquid film had disappeared and heat transfer would be weakened for the low heat conductivity of the vapor, while R22 had significantly higher liquid viscosity than R161, which was beneficial for the formation and preserving of the liquid film, and heat transfer would not be rapidly deteriorated in high quality region. As shown in Fig. 7, the heat transfer coefficient of R22 rose with the decrease
0.0
0.2
0.4
0.6
0.8
1.0
Vapor quality Fig. 7. Local heat transfer coefficient for various saturation temperatures, at q = 38.62 kW/m2 and G = 175 kg/m2 s.
of the saturation temperature during the total quality range. But for R161, the heat transfer coefficient under different vapor qualities were influenced by the saturation temperatures unpredictably. This trend might be supported by Seo and Kim [25], who reported that the increase of heat transfer coefficient with the saturation temperatures was lowered in micro-fin tubes. Kuo et al. [26] concluded that the heat transfer coefficient increased when the saturation temperature rose from 5 to 15 °C in their micro-fin tube. It should be pointed out that the increase of the saturation temperature would lead to the enhancement of the nucleate boiling. But on the other hand, as the saturation temperature decreases, the growth of the two-phase Reynolds number might also boost the heat transfer coefficient [27]. Thus, the influence of saturation temperatures on the heat transfer performance was relatively intricate according to different test fluids and test conditions. Fig. 8 showed the variation of local heat transfer coefficient of R22 and R161 with the vapor quality for different heat flux when the saturation temperature and mass flux are fixed at –5 °C and 175 kg/m2 s, respectively. Generally, the heat transfer coefficient went up with the heat flux. The growth in the heat transfer coefficient with increasing heat flux seemed to be more obvious in low vapor quality region. This might be ascribed to the fact that the nucleate boiling was much stronger than the forced convection at low vapor qualities, and it would be promoted as the imposed heat flux increased. However, when the convective evaporation dominated heat transfer process at high qualities, the effect of the heat flux was very slight. By further observation from Fig. 8, the growth of heat transfer coefficient with increasing heat flux for R161 seemed to be more pronounced than that for R22. Fig. 9 showed the average heat transfer coefficient versus the heat flux for various saturation temperatures at a mass flux of 175 kg/m2 s. Generally, the average heat transfer coefficient showed a higher value at lower saturation temperatures. The average heat transfer coefficient of R161 increased by about 10% when the saturation temperature dropped from 8 °C to 5 °C. But for R22, the influence of saturation temperature on average heat transfer coefficient did not present an obvious regularity at the high heat flux region. Fig. 9 also indicated that the average heat transfer coefficient increased with the heat flux. As the saturation temperature was fixed at 0 °C, the average values of the heat transfer coefficients at 49.29 kW/m2 were about 14% and 5% higher than those at 28.15 kW/m2 for R22 and R161, respectively. For R161, the heat transfer coefficient at lower saturation temperatures was more sensitive to the heat flux. As to R22, Seo et al. [27] reported that for a micro-fin tube, the heat transfer coefficient reduced slightly
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8000
Tsat= -5 °C
Local heat transfer coefficient , W/m K
2
2
2
Local heat transfer coefficient, W/m K
8000
G = 175 kg/m s 6000
4000
2
R161, q =28.15 kW/m 2 R161, q =38.62 kW/m 2 R161, q =49.29 kW/m 2 R22, q =28.15 kW/m 2 R22, q =38.62 kW/m 2 R22, q =49.29 kW/m
2000
6000
4000 2
2000
Tsat = 8 °C 2
q = 28.15 kW/m 0
0
0.0
0.2
0.4
0.6
0.8
0.0
1.0
0.2
0.4
0.6
0.8
1.0
Vapor quality
Vapor quality Fig. 8. Local heat transfer coefficient for various heat flux, at Tsat = 5 °C and G = 175 kg/m2 s.
Fig. 10. Local heat transfer coefficient for various mass flux, at Tsat = 8 °C and q = 28.15 kW/m2.
8000
2
2
Average heat transfer coefficient, W/m K
8000
Average heat transfer coefficient, W/m K
R161, G =100 kg/m s 2 R161, G =175 kg/m s 2 R161, G =250 kg/m s 2 R22, G =100 kg/m s 2 R22, G =175 kg/m s 2 R22, G =250 kg/m s
6000
4000
R161, Tsat =-5 °C R161, Tsat
= 0 °C
R161, Tsat = 8 °C
2
G = 175kg/m s
R22, Tsat =-5 °C
2000
R22, Tsat = 0 °C R22, Tsat = 8 °C
6000
4000
R161, Tsat= -5 °C R161, Tsat= 0 °C R161, Tsat= 8 °C
2
R22, Tsat= -5 °C
q =28.15 kW/m
2000
R22, Tsat= 0 °C R22, Tsat= 8 °C 0
0 30
40
50 2
Heat flux, kW/m
100
200
300
2
Mass flux, kg/m s
Fig. 9. Average heat transfer coefficient versus the heat flux for various saturation temperatures, at G = 175 kg/m2 s.
Fig. 11. Average heat transfer coefficient versus the mass flux for various saturation temperatures, at q = 28.15 kW/m2.
as the heat flux increased. The considerably higher heat flux together with higher saturation temperatures in this study might account for the opposite conclusion. The influence of the heat flux on the heat transfer is very complicated. For one hand, the high heat flux is beneficial for the nucleate boiling at low vapor qualities, and the heat transfer will be promoted, especially at high saturation temperatures; for another, the flow regime transition at a high heat flux is easier than at a low one, and the dry-out flow at high qualities will result in the decrease of average heat transfer coefficients. Fig. 10 presented variation of the local heat transfer coefficients of R22 and R161 with the vapor quality for different mass fluxes at a saturation temperature of 8 °C and a heat flux of 28.15 kW/m2. As expected, the heat transfer coefficient increased strongly with the increase of the mass flux. Most previous studies [28,29] also reported the same trend. Furthermore, the increase of the heat transfer coefficient with the mass flux seemed more pronounced in high quality region. The convective evaporation dominated the heat transfer process at high qualities, and it would be boosted by the higher mass flux. Fig. 11 presented the average heat transfer coefficient versus mass flux for various saturation temperatures at a heat flux of 28.15 kW/m2. As indicated in Fig. 11, the heat transfer coefficient
was strongly dependent on the mass flux. As the saturation temperature was fixed at 0 °C, the average heat transfer coefficients at 250 kg/m2 s were about 72% and 84% higher than those at 100 kg/m2 s, respectively for R22 and R161. Furthermore, the effect of mass flux on the heat transfer coefficient was more significant in the lower mass flux region, and R161 seemed to be more sensitive to the mass flux than R22. Fig. 12 presented the ratio of average heat transfer coefficient between the two refrigerants (R161/R22) versus the heat flux at various saturation temperatures and mass fluxes. According to Fig. 12, the heat transfer coefficient of R161 was significantly enhanced based on R22. The enhancing percentage generally reduced with the increased heat flux and the reduced mass flux, floating between 0% and 50%. Except for the point with a saturation temperature of 8 °C and a mass flux of 100 kg/m2 s, the ratios under different test conditions merged together around 1.23, indicating that the heat transfer coefficient of R161 in present tube was about 23% higher than those of R22. According to the REFPROP [20], R161 had better thermophysical properties than R22 in aspect of the boiling heat transfer. For example, at the saturation temperature of 25 °C,the liquid heat conductivity of R161 is 0.13 W/m K, while that of R22 is only 0.084 W/m K; the vaporization latent heats of R161 and 22 are 338.47 kJ/kg and 182.89 kJ/kg, respectively.
645
2.5
1.8
Tsat =-5 °C, G = 175 kg/m s
Tsat = 0 °C, G = 100 kg/m s
Tsat = 0 °C, G = 175 kg/m s
Tsat = 8 °C, G = 100 kg/m s
Tsat = 8 °C, G = 175 kg/m s
2
2
2
2
2
1.5
Tsat =-5 °C, q = 28.15 kW/m
2
Tsat = -5 °C, G = 100 kg/m s
2
1.23 1.2
0.9
Ratio of pressure drop, R22/R161
Ratio of average heat transfer coefficient, R161/R22
X. Han et al. / International Journal of Heat and Mass Transfer 62 (2013) 638–646
Tsat = 0 °C, q = 28.15 kW/m
2
2.0
Tsat = 8 °C, q = 28.15 kW/m
2
1.65 1.5
1.0 0.88
Tsat = -5 °C, q =38.62 kW/m
2
0.5
Tsat = 0 °C, q = 38.62 kW/m
2
Tsat = 8 °C, q =38.62 kW/m
2
0.0
0.6 30
40
100
50
200
Mass flux, kg/m s
Heat flux, kW/m
10
R22, Tsat =- 5 °C
6
R22, Tsat = 0 °C
Fig. 14. Ratio of pressure drop versus the mass flux, at q = 28.15 and 38.62 kW/m2.
Frictional pressure drop of per unit length, kPa/m
Frictional pressure drop of per unit length, kPa/m
Fig. 12. Ratio of average heat transfer coefficient versus the heat flux, at G = 100 and 175 kg/m2 s.
8
2
q =28.15kW/m
R22, Tsat = 8 °C R161, Tsat = -5 °C R161, Tsat = 0 °C R161, Tsat = 8 °C
4
2
3.5 3.0
2
G =100 kg/m s
R161, Tsat = 0 °C
R22, Tsat = 0 °C
R161, Tsat = 8 °C
2.0 1.5 1.0 0.5 0.0 30
200
R161, Tsat = -5 °C
R22, Tsat = -5 °C R22, Tsat = 8 °C
2.5
0 100
300
2
2
300
40
50 2
Heat flux, kW/m
2
Mass flux, kg/m s Fig. 13. Pressure drop versus the mass flux for various saturation temperatures at q = 28.15 kW/m2.
4.3. Pressure drop Fig. 13 showed the pressure drop of per unit length versus the mass flux for various saturation temperatures as the heat flux fixed at 28.15 kW/m2. Generally, the pressure drop strongly increase with the mass flux both for R22 and R161. Taking R161 with the saturation temperature of 5 °C for example, the pressure drop at G = 175 kg/m2 s was about 4.7 kPa/m, more than twice that at G = 100 kg/m2 s. According to Fig. 11, the pressure drop was also significantly dependent on the saturation temperatures, it dropped with the increasing saturation temperatures. Furthermore, the pressure drop difference in different saturation temperatures increased with the mass flux, and the influence of the mass flux on the pressure drop was reduced for higher saturation temperatures. The increase of the pressure drop with the decreasing saturation temperatures was mainly attributed to the augmentation of the specific volume and the liquid viscosity of the evaporating refrigerant. It still can be found from Fig. 13 that R161 displayed a lower pressure drop compared with R22, but the difference reduced gradually with the mass flux. Even when the mass flux exceeded a certain value (close to 250 kg/m2 s), the pressure drop of R161 became slightly higher than that of R22 at the saturation temperature of 5 and 0 °C.
Fig. 15. Pressure drop versus the heat flux for various saturation temperatures, at G = 100 kg/m2 s.
Fig. 14 showed the ratio of the pressure drop between R22 and R161 versus the mass flux for various saturation temperatures at the heat flux of 28.15 and 38.62 kW/m2. Generally, the ratio dropped with the mass flux, ranging from 0.88 to 1.65. For another, the ratio of the pressure drop was affected by the heat flux, generally the higher heat flux caused a relatively higher growth in the pressure drop. The previous study by Seo et al. [27] suggested that the pressure drop difference between R22 and R410A decreased as the saturation temperature dropped. However, this tendency was not significant in present study for R161 and R22. The difference in the thermophysical properties for R22 and R161, liking the viscosity and specific volume, were deemed to be responsible for the difference of pressure drop. Fig. 15 presented the pressure drop versus heat flux for various saturation temperatures at a mass flux of 100 kg/m2 s, it suggested that the increase in heat flux resulted in a mild increase of the pressure drop. 5. Conclusions R161 is regarded as one of the promising alternative refrigerants of R22. The evaporation heat transfer characteristics of R161 in a horizontal micro-fin tube with an outside diameter
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of 7 mm were investigated by experiment. The heat transfer coefficient and the pressure drop were measured when the heat flux ranged from 28.15 to 49.29 kW/m2, the mass flux ranged from 100 to 250 kg/m2 s, and the saturation temperature ranged from 5 to 8 °C. The heat transfer coefficient of R161 went up with the vapor quality before the dry-out region. The heat transfer coefficient grew with the mass flux and the heat flux, and dropped with the saturation temperature within the experimental conditions. The heat transfer coefficient of R161 was enhanced by about 23% by comparison with R22 on the same test conditions. The pressure drop rose with the mass flux and dropped with the saturation temperature. A rise in the imposed heat flux resulted in a mild increase of the pressure drop. The ratio of pressure drop (R22/R161) ranged from 0.88 to 1.65. The difference in the thermophysical properties of R22 and R161 had an important effect on the enhancement of the heat transfer coefficient and the pressure drop. Acknowledgments This work has been supported by the Nation Natural Science Foundation of China (Grant No. 51176166), and the Program for Key Innovative Research Team of Zhejiang Province (No. 2009R50036), China. We still are very full of appreciation for the help and direction of this project from Prof. Jacobi M. Anthony. References [1] Y.M. Xuan, G.M. Chen, Experimental study on HFC-161 mixture as an alternative refrigerant to R502, Int. J. Refrig. 28 (2005) 436–441. [2] X.H. Han, G.M. Chen, C.S. Li, X.G. Qiao, X.L. Cui, Isothermal vapor–liquid equilibrium data for the binary mixture refrigerant pentafluoroethane (R125) + fluoroethane (R161) at 265.15, 275.15, 283.15, 293.15, 303.15 and 303.15 K with a recirculating still, J. Chem. Eng. Data 51 (4) (2006) 1232–1235. [3] X.H. Han, Q. Wang, Z.W. Zhu, G.M. Chen, Cycle performance study on R32/ R125/R161 as an alternative refrigerant to R407C, Appl. Therm. Eng. 27 (2007) 2559–2565. [4] X.H. Han, Z.W. Zhu, F.S. Chen, Y.J. Xu, Z.J. Gao, G.M. Chen, Solubility and miscibility for the mixture of (Ethyl Fluoride + Polyol Ester Oil), J. Chem. Eng. Data 55 (9) (2010) 3200–3207. [5] X.H. Han, Z.J. Gao, Y.J. Xu, Y. Qiu, X.W. Ming, Q. Wang, G.M. Chen, Isothermal vapor–liquid equilibrium data for the binary mixture difluoromethane (R-32) + ethyl fluoride (R-161) over a temperature range from 253.15 to 303.15 K, Fluid Phase Equilib. 299 (1) (2010) 116–121. [6] X.H. Han, Y.J. Xu, X.W. Min, Z.J. Gao, Q. Wang, G.M. Chen, Density data for the refrigerant ethyl fluoride (HFC-161) over a temperature range from (233 to 343) K, J. Chem. Eng. Data 56 (7) (2011) 3038–3042. [7] X.H. Han, Y. Qiu, P. Li, Y.J. Xu, Q. Wang, G.M. Chen, Cycle performance studies on HFC-161 in a small-scale refrigeration system as an alternative refrigerant to HFC-410A, Energy Build. 44 (2012) 33–38.
[8] L.M. Schlager, M.B. Pate, A.E. Bergles, Evaporation and condensation heat transfer and pressure drop in horizontal, 12.7-mm micro-fin tubes with refrigerant 22, J. Heat Transfer 112 (1990) 1041–1045. [9] C.B. Dang, S.Z. Saitoh, Y.T. Nakamura, M.X. Lia, E.J. Hiharaa, Boiling heat transfer of HFO-1234yf flowing in smooth small-diameter horizontal tube, in: Proc. 2010 International Symposium on Next-generation Air Conditioning and Refrigeration Technology, Tokyo, Japan, February 17–19, 2010. [10] A. Greco, G.P. Vanoli, Flow-boiling of R22, R134a, R507, R404A and R410A inside a smooth horizontal tube, Int. J. Refrig. 28 (2005) 872–880. [11] R.L. Zeng, X. Wang, B.J. Chen, Y.P. Zhang, J.L. Niu, X.C. Wang, H.F. Di, Heat transfer characteristics of microencapsulated phase change material slurry in laminar flow under constant heat flux, Appl. Energy 86 (2009) 2661–2670. [12] Y. Kim, K. Seo, J.T. Chung, Evaporation heat transfer characteristics of R-410A in 7 and 9.52 mm smooth/micro-fin tubes, Int. J. Refrig. 25 (2002) 716–730. [13] M. Yu, T. Lin, C. Tseng, Heat transfer and flow pattern during two-phase flow boiling of R-134a in horizontal smooth and micro-fin tubes, Int. J. Refrig. 25 (2002) 789–798. [14] M. Lallemand, C. Branescu, P. Haberschill, Local heat transfer coefficients during boiling of R22 and R407C in horizontal smooth and micron tubes, Int. J. Refrig. 24 (2001) 57–72. [15] E. Torrella, J. Navarro-Esbri, R. Cabello, Boiling heat-transfer coefficient variation for R407C inside horizontal tubes of a refrigerating vapourcompression plants shell-and-tube evaporator, Appl. Energy 83 (2006) 239– 252. [16] H.T. Hu, G.L. Ding, K.J. Wang, Heat transfer characteristics of R410A – oil mixture flow boiling inside a 7 mm straight microfin tube, Int. J. Refrig. 31 (2008) 1081–1093. [17] S.M. Yang, W.Q. Tao, Heat Transfer, third ed., High Education Press, China, 1998. pp. 162–168. [18] J.G. Collier, Convective Boiling and Condensation, second ed., McGraw-Hill International Book Company, 1981. pp. 26–108. [19] R.J. Moffat, Describing the uncertainties in experimental results, Exp. Therm. Fluid Sci. 1 (1988) 3–17. [20] E.W. Lemmon, M.O. McLinden, M.L. Huber, NIST reference fluid thermodynamic and transport properties – REFPROP, Version 7.0. NIST, 2002. [21] R. Yun, Y. Kim, K. Seo, H.Y. Kim, A generalized correlation for evaporation heat transfer of refrigerants in micro-fin tubes, Int. J. Heat Mass Transfer 45 (2002) 2003–2010. [22] G. Kandlikar, T. Raykoff, Predicting flow boiling heat transfer of refrigerants in micro-fin tubes, J. Enhanced Heat Transfer 4 (1997) 257–268. [23] A. Cavallini, D. Del Col, L. Doretti, G.A. Longo, L. Rossetto, Enhanced in tube heat transfer with refrigerants, in: Proc. 20th International Congress of Refrigeration, IIR/IIF, Sydney, Australia, 1999, pp. 1–14. [24] L.M. Chamra, P.J. Mago, Modelling of evaporation heat transfer of pure refrigerants and refrigerant mixtures in micro-fin tubes, J. Mech. Eng. Sci. 221 (4) (2007) 443–447. [25] K. Seo, Y. Kim, Evaporation heat transfer and pressure drop of R-22 in 7 and 9.52 mm smooth/micro-fin tubes, Int. J. Heat Mass Transfer 43 (2000) 2869– 2882. [26] C.S. Kuo, C.C. Wang, W.Y. Cheng, D.C. Lu, Evaporation of R-22 in a 7 mm microfin tube, ASHRAE Trans. 95 (1995) 1055–1061. [27] K. Seo, Y. Kim, K.J. Lee, Y.C. Park, An experimental study on convective boiling of R-22 and R-410A in horizontal smooth and micro-fin tubes, J. Mech. Sci. Technol. 15 (2001) 1156–1164. [28] C.C. Wang, C.S. Kuo, Y.J. Chang, D.C. Lu, Two-phase flow heat transfer and friction characteristics of R-22 and R-407C, ASHRAE Trans. 102 (1996) 830– 838. [29] H. Wijaya, M.W. Spatz, Two phase flow heat transfer and pressure drop characteristics of R-22 and R-32/125, ASHRAE Trans. 101 (1995) 1020–1027.
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International Journal of Heat and Mass Transfer journal homepage: www.elsevier.com/locate/ijhmt
Evaporation heat transfer and pressure drop of R161 in a 7 mm micro-fin tube Xiaohong Han, Peng Li, Zheng Wang, Xuehui Wang, Xuejun Zhang ⇑, Guangming Chen Institute of Refrigeration and Cryogenics, State Key Laboratory of Clean Energy Utilization, Zhejiang University, Hangzhou 310027, China
a r t i c l e
i n f o
Article history: Received 14 November 2012 Received in revised form 28 February 2013 Accepted 2 March 2013 Available online 11 April 2013 Keywords: R161 Evaporation heat transfer Micro-fin tube
a b s t r a c t The evaporation characteristics of refrigerant fluoroethane (R161) in a horizontal micro-fin tube with an outside diameter of 7 mm were investigated by experiment. The heat transfer coefficient and the pressure drop were measured when the heat flux ranged from 28.15 to 49.29 kW/m2, the mass flux ranged from 100 to 250 kg/m2 s, and the saturation temperatures ranged from 5 to 8 °C. According to the results, the heat transfer coefficient went up with the mass flux and the heat flux, and dropped with the saturation temperatures within the experimental conditions. The pressure drop grew with the mass flux and decreased with the saturation temperatures. By comparison, the heat transfer coefficient of R161 was about 23% higher than that of R22. In addition, the ratio of pressure drop (R22/R161) ranged from 0.88 to 1.65. The presented research is helpful in designing more compact and effective heat transfer exchangers for the air-conditioning systems using R161. Ó 2013 Elsevier Ltd. All rights reserved.
1. Introduction As providing comfortable living conditions for us, the air conditioning and heat pump technology exert some negative influences on the environment. The research of alternative refrigerants is recognized as one of the urgent issues to ease the ozone depletion and global warming. For the excellent environmental performance (Ozone Depletion Potential (ODP) = 0, Global Warming Potential (GWP) = 12), high energy efficiency, and good commonality with the existing systems [1], fluoroethane (R161) is another potential substitute of R22 in the small-scale air conditioning. So far, a lot of researches about R161 and the related refrigerant mixtures have been conducted, for example, PVTx (Pressure–Volume–Temperature-Composition), the saturated vapor pressure, the vapor–liquid equilibrium, the solubility with POE (Polyol Ester), cycle performance, etc. [2–7]. On the other hand, micro-fin tubes are widely used in the air-conditioning applications for their better heat transfer ability together with larger heat transfer area. Though inner fins will inevitably cause larger refrigerant pressure drop in comparison with the smooth tubes, the enhancement of heat transfer is found to be more significant [8]. When a new refrigerant is proposed, the heat transfer characteristics are usually regarded as one of the research focuses in order to analyze the performance of the evaporators and condensers in the actual air conditioning systems. For example, Dang et al. [9] studied the boiling heat transfer of HFO1234yf in a small diameter horizontal smooth tube. The results ⇑ Corresponding author. Tel./fax: +86 571 87952446. E-mail address: [email protected] (X. Zhang). 0017-9310/$ - see front matter Ó 2013 Elsevier Ltd. All rights reserved. http://dx.doi.org/10.1016/j.ijheatmasstransfer.2013.03.017
indicated that heat transfer was influenced by the mass flux more significantly at high vapor qualities, and for the low quality region, heat flux showed large effect. Greco and Vanoli [10] measured the boiling heat transfer coefficient for R22, R507, R134a, R404A and R410A in a smooth tube, the results revealed that R134a had the better heat transfer performance than the other refrigerants. Based on research about the convective heat transfer inside an annular tube, Zeng et al. [11] reported that the tube wall temperature and the Nusselt number’s fluctuation profile were mostly influenced by the Stefan number and the dimensionless phase change temperature range in the phase-change heat transfer process. Kim et al. [12] found that the evaporating heat transfer coefficient of R410A in both the smooth tubes and the micro-fin tubes increased with the mass flux and the heat flux. They also reported that at the same test conditions, heat transfer performance of micro-fin tubes was significantly superior to the smooth tubes, and the improvement degree was affected by the tube diameters. Yu et al. [13] investigated the evaporating heat transfer performance of R134a in a 10.7 mm (OD) smooth tube and a 10.7 mm (OD) micro-fin tube by the experiment. They found that for the annular and intermittent flows, the experimental heat transfer coefficient kept an agreement with the Gungor and Winterton correlation with corrected constants. Lallemand et al. [14] studied the boiling heat transfer performance of R407C and R22 in the smooth and micro-fin tubes, respectively. The results showed that the heat transfer coefficient of R407C was 15% and 35% lower than that of R22 in the smooth and micro-fin tubes, respectively. Torrella et al. [15] concluded that boiling heat transfer coefficient of R407C was influenced by the refrigerant flow rate and the evaporation temperature. Hu et al.
X. Han et al. / International Journal of Heat and Mass Transfer 62 (2013) 638–646
639
Nomenclature A D Di,av Di,max e eav G GWP h hi hav H i ilg I L Lx _ m N ODP PVTx POE P 4P Dpa Dpf Dpi
area (m2) diameter (mm) average inside diameter (mm) bottom inside diameter (mm) bottom wall thickness (mm) average wall thickness (mm) mass flux (kg/(m2 s)) Global Warming Potential heat transfer coefficient (W/m2 K) local heat transfer coefficient (W/m2 K) average heat transfer coefficient (W/m2 K) fin height (mm) enthalpy (kJ/kg) latent heat of vaporization (kJ/kg) current (A) length (m) the length from the inlet to the certain measurement point mass flow rate (kg/s) number of fins Ozone Depletion Potential Pressure–Volume–Temperature-Composition Polyol Ester pressure total pressure drop (kPa) acceleration pressure drop (kPa) frictional pressure drop (kPa) inlet pressure drop (kPa)
[16] studied the boiling heat transfer characteristics of two-phase R410A–oil mixture in a micro-fin tube. It was revealed that at lower vapor qualities (x < 0.4), heat transfer was enhanced by the presence of oil. While at the higher qualities (x > 0.65), the heat transfer coefficient dropped significantly as the oil concentration increased. In order to enrich the database of the potential alternative refrigerant and supply reference for designing more compact and effective evaporators for air conditioning systems using R161, evaporation heat transfer of R161 inside a horizontal micro-fin tube was investigated by experimental methods in this work. According to the results, effect of the saturation temperature, the mass flux and the heat flux on the heat transfer coefficient and the pressure drop were analyzed in detail. Moreover, the measured results of R161were compared with those of R22. 2. Experiment 2.1. Experimental apparatus The experimental setup used to study the evaporation heat transfer characteristics of R161 is shown in Fig. 1. It mainly includes an electric heating system, a data collection system and three main loops: the refrigerant loop, the cooling loop, and the pre-heating loop. The electric heating system provides the heat needed to evaporate the refrigerant in the test section, and it mainly contains a DC power supply and an array of heating band. The DC power supply can continuously output the voltage and the current at the range of 0–100 V and 0–30 A, respectively, with high stability and display precision (uncertainty: ±0.2%), and its maximum output power is 3.0 kW. The material of the heating band is silicone rubber and Ni–Cr wire with a temperature resistance of 250 °C, a insulation resistance of 50 MX and a maximum heat capacity of 280 W/m. The data collection system mainly includes an Agilent 34970A and a PC installed with the data acquisition program, the sample
Dpo q Q T U x xi Greek k b
c
outlet pressure drop (kPa) heat flux (kW/m2) total heat input (kW) temperature (°C) voltage (V) vapor quality (–) local vapor quality
thermal conductivity (W/m K) helix angle (°) fin angle (°)
Subscripts bottom the bottom of the tube ev evaporation in inlet i inside l liquid phase out outlet o outside r refrigerant side the side of the tube sat saturation top the top of the tube wo outside wall wi inside wall
frequency of Agilent 34970A is set at 12 time/min in this study. The refrigerant loop consists of a test tube (7 mm O.D., 2 m length), two sight glasses, a condenser, a receiver, a refrigerant pump, an accumulator, a filter/dryer, a Coriolis mass flow meter (CMF) and a pre-evaporator. The refrigerant pump (lift: 0.6 MPa, rated flow: 0.316 mL/r) is a magnetically gear pump with high conveying stability and excellent sealing. Furthermore, it is a variable-speed pump working together with a frequency converter (0–600 Hz) to transfer the liquid refrigerant and adjust the mass flux for the test section. At the outlet of the pump, an accumulator is accessed in order to alleviate the fluctuation in pressure and flow rate of the refrigerant loop. The mass flow meter (range: 0–2180 kg/h, uncertainty: ±0.1%, sample frequency: 60 time/min) records the refrigerant mass flow rate in real time with high precision and low pressure loss. Considering the necessary overhaul and preventing the possible impurity and lubricant oil from accumulating in the mass flow meter when the refrigerant loop is cleaned, a by-pass loop is installed in parallel with the flow meter. The pre-evaporator, installed before the test section, is used to evaporate the sub-cooled refrigerant liquid to the two-phase state. Actually, in order to obtain the local two-phase heat transfer coefficient of test refrigerant at the limited tube length and heating power, the inlet refrigerant is expected to be in the saturated state with a certain vapor quality and a specified saturation temperature. A thermostatic water tank with the capacity of 8.0 kW supplies the hot water which is used to heat the refrigerant liquid in the pre-evaporator. An electromagnetic flow meter (range: 0–3.5 m3/h, uncertainty: ±0.2%, sample frequency: 60 time/min) is installed in the pre-heating loop to record the flow rate of hot water passing through the pre-evaporator and the inlet/outlet temperatures are measured by thermocouples. Indeed, after the pre-evaporator, there still exists a calming section before the test section. The calming section is to guarantee the inlet refrigerant to be in the fully developed flow regime, and it has a total length
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Fig. 1. Schematic diagram of the experimental system.
of 45 cm. For the turbulent flow in this study, the calming section with a length of >60Di (38 cm) can fully satisfy the need for fully developed flow [17] . The condenser installed after the test section is used to condense the evaporated refrigerant coming from the test section to be in the sub-cooled state. The water–alcohol mixture (40%:60% in mass fraction) works as the cooling medium in another side of the condenser. 2.2. Test section The detailed information about the test section and measurement positions for temperatures and pressures is shown in Fig. 2. Table 1 and Fig. 3 give the geometric parameters of the test micro-fin tube. The tube was heated by the DC power supply through silicone rubber heating band glued on the tube outside surface. A thin layer of heat conduction silicone was coated between the heating band and the tube surface to spread the heat transferred to the refrigerant. The test section was insulated to the ambient with three layers of rubber foam, and the total thickness of the insulation is 24 mm with the thermal conduction ability of
no more the 1 W/m2 K. At twenty locations along the axial direction of the test tube (seen in Fig. 2), T-type thermocouples with the uncertainties of ±0.1 °C were used to measure the local outside surface temperature. And for each location, there were four thermocouples fixed on the top, bottom, right, left sides of the tested tube, respectively. To avoid being disturbed by the current, each thermocouple was separated with the tube outside surface by a layer of 0.03 mm thick Teflon sheet. A high-precision resistance thermometer detector (RTD) with the uncertainty of ±0.02 °C was used to calibrate all the thermocouples before the measurement. The layout of Teflon sheet, thermocouples, heat conduction silicone, heating band and rubber foam are shown in Fig. 4. Applying Fourier’s heat conduction law for the one dimensional steady state, the inside surface temperature used in calculating of local heat transfer coefficient was obtained from the measured outside surface temperatures. At the inlet and the outlet of the test section, one four-wire Pt100 RTD with the uncertainty of ±0.03 °C was inserted to the flow stream to directly measure the refrigerant temperature (Tin and Tout), respectively. At the same time, the corresponding refrigerant pressures were recorded by the pressure
Fig. 2. Schematic diagram of the test section.
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X. Han et al. / International Journal of Heat and Mass Transfer 62 (2013) 638–646 Table 1 Dimension of the test tube.
Table 2 Experimental conditions.
Test tube parameter
Micro-fin
Outside diameter, Do (mm) Tube length, L (mm) Bottom wall thickness, e (mm) Average wall thickness, eav (mm) Average inside diameter, Di,av (mm) Bottom inside diameter, Di,max (mm) Helix angle, b (°) Fin angle , c (°) Fin-height, H (mm) Number of fins, N Inside surface area, Ai (cm2/m)
7.0 2000 0.28 0.33 6.34 6.41 15 34 0.1 65 286.7
Refrigerant
R161 R22
Conditions Saturation temperature (°C)
Mass flux (kg/m2 s)
Heat flux (kW/m2)
5, 0, 8
100, 175, 250
28.15, 38.62, 49.29
R22 and R161, respectively, shown in Table 2. Furthermore, in order to obtain the heat transfer characteristics in a wide quality range of 0.1–1, the length of micro-fin was chosen for 2.0 m. Thus, a relative high outlet quality will be acquired and heat transfer performance at the fully developed flow regimes is accessible. 3. Data reduction The heat transfer coefficient was given by:
q¼
Fig. 3. Detailed information of the microfin tube.
Fig. 4. Cross-section view of the test tube.
transducers with the uncertainties of ±0.2% F.S. (full scale: 0–2.5 MPa), and the whole uncertainties were within 5 kPa. By comparison, the measured inlet and outlet refrigerant temperatures were consistent with those calculated from the measured saturation pressures within ±0.3 °C. A differential pressure transducer with high-precision (±0.1% F.S., the whole uncertainty is within 0.037 kPa), was used to monitor the pressure drop. T-type thermocouples were used to record the inlet and outlet temperatures of the cold and hot liquids in the cooling loop and the preheating loop. The measurement signals of all the thermocouples, RTD, pressures transducers, differential pressure transducer are collected by the Agilent data acquisition unit. When all the parameters (temperatures, pressures, flow rate, output voltage/current, etc.) of different positions have kept stable for more than one hour, it is believed that the steady state was reached, and the data acquisition starts. 2.3. Test conditions The tests were conducted for three saturation temperature points, three mass flux points and three heat flux points with
Q ¼ hev ðT wi T ev Þ Ai
ð1Þ
where hev is the evaporation heat transfer coefficient, Twi is the inside wall temperature of the test tube, and Tev is the refrigerant evaporation temperature. The heat flux q over the whole test section was supposed to be uniform for the negligible heat conduction along the axial direction of the tube and the excellent heat transfer between the tube surface and the heating band. Due to the good insulation for the test tube, heat losses were neglected and the heat flux was estimated from the total heat input (Q(Q = U I), U and I are the output voltage and current of the DC power supply, respectively). The inside wall area Ai was calculated according to the average inside diameter Di,av of present tube (Di,av was a equivalent diameter, and it was calculated according to the wetted cross-section area of the microfin tube, and the wetted cross-section area could be calculated based on the Fig. 3). Applying the heat conduction equation (one-dimensional, radial, and steady-state), the local inside wall temperature in Eq. (1), Twi, was calculated by:
T wi ¼ T wo
Q ln DDi;aov
ð2Þ
2pkL
where k and L are the material thermal conductivity and the length of the test tube, respectively. Two is the average value of measured temperatures at the top, the two sides and the bottom of the tube outside surface:
T wo ¼
T wo;top þ 2 T wo;side þ T wo;bottom 4
ð3Þ
Due to the pressure drop, the local refrigerant temperature, Tev gradually reduced from the tube inlet to the tube outlet [12], and it was interpolated with a linear equation [12]:
T eV ¼ T in þ ðT out T in Þ
Lx L
ð4Þ
where Tin and Tout are the measured inlet and outlet refrigerant temperatures, respectively. Lx is the length from the inlet to the certain measurement point. The saturation temperature Tsat is defined as the average value of Tin and Tout. The refrigerant vapor quality at every measurement point of different location was obtained based on the local refrigerant temperature:
x¼
i il ilg
ð5Þ
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where il is the specific enthalpy of the saturated liquid phase at the certain temperature, ilg is the latent heat of vaporization, i is the specific enthalpy of two-phase refrigerant, determined by:
i ¼ iin þ
q Ai _r m
ð6Þ
where iin is obtained from the sub-cooled liquid through an energy _ r is the refrigerant mass flow rate. balance in the pre-evaporator, m The average heat transfer coefficient was defined by:
P20 h i xi hav ¼ Pi¼1 20 i¼1 xi
ð7Þ
where xi and hi are the local refrigerant vapor quality and heat transfer coefficient at the certain measurement point of different position, respectively. The expression of the frictional pressure drop,Dpf, was:
Dpf ¼ Dp Dpi þ Dpo Dpa
ð8Þ
where Dp represents the tall pressure drop measured by the differential pressure transducer, Dpa is the acceleration pressure drop, Dpi and Dpo are inlet and outlet pressure losses, respectively. Estimation for Dpa, Dpi and Dpo was proposed by Collier [18]. The standard error analysis presented by Moffat [19] was introduced to estimate the uncertainties of the heat transfer coefficient and other calculated parameters, for instance the heat flux, mass flux, vapor quality and the frictional pressure drop, the results are shown in Table 3. Taking the heat transfer coefficient for example, the uncertainty of it is estimated as follows:
@hev ¼ hev
sffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 2 2 2 2 2 @U @I @Ai @T wi @T eV þ þ þ þ U I Ai T wi T ev T wi T eV ð9Þ
According to the natural uncertainties of the measurement equipments and the actual measured data during the experiment, the relative uncertainties of the voltage difference U and the current I are estimated at 0.27% and 0.39%, respectively, the relative uncertainty of the inside wall area is assumed at 9.2%. the inside wall temperature Twi and the refrigerant evaporation temperature Tev are assumed to be typical, 3.6 °C and 0 °C (absolute value), respectively, and their relative uncertainties with the temperature difference Twi Tev are estimated at 9.0% and 0.83%, respectively, so the uncertainty of heat transfer coefficient is 12.9%. The transport and physical properties of R22 and R161 were obtained by using REFPROP 7.0 [20].
Table 3 Estimated uncertainties. Parameter
Uncertainties
T Tev P 4P _r m U I G q x hev hav 4Pf
±0.1 °C ±0.03 °C ±0.2% ±0.1% ±0.1% ±0.2% ±0.2% ±1.2% ±1.2% ±2.2% ±12.9% ±15.5% ±9.4%
Note: T in this table refers to all the temperatures measured by the thermocouples.
4. Results and discussion The heat transfer coefficient and the pressure drop were obtained for the mass flux range of (100–250) kg/m2 s, the heat flux range of (28.15–49.29) kW/m2 and the evaporation temperature range of (5 to 8) °C. 4.1. Comparison with existing correlations and previous experimental data 4.1.1. Convergence verification by comparison between the experimental data of this work and results from existing correlations It was found that many correlations about the refrigerant boiling heat transfer have been proposed for the smooth tubes. Due to lack of the related database as well as the difficulties in correlating the complicated parameters, models for micro-fin tubes are limited in open literatures [21]. Two enhancement factors were introduced by Kandlikar and Rayoff [22] into a smooth tube correlation to predict the boiling heat transfer coefficient in micro-fin tube, but for different micro-fin tubes and different test refrigerants, the enhancement factors were specifically empirical constants. Cavallini et al. [23] proposed a micro-fin tube model based on their condensation correlation, including parameters liking the surface tension and the fin height. A semi-empirical model for estimating the evaporation heat transfer inside micro-fin tubes were developed by Chamra and Mago [24], which considered the nucleate boiling and the forced convention. Yun et al. [21] generalized a boiling correlation for horizontal micro-fin tubes based on a smooth tube model, and non-dimensional parameters were introduced to explain the heat transfer enhancement of micro-fin tubes in this correlation. The present heat transfer coefficient of R22 was compared with the results from the above mentioned correlations, shown in Fig. 5. It can be observed from Fig. 5 that the correlation of Kandlikar and Rayoff [22], and the correlation of Cavallini et al. [23] overpredicted the heat transfer coefficient by more than 35% and 46%, respectively, while the Chamra and Mago [24] correlation under-predicted the heat transfer coefficient by more than 30%. Within an error band of ±30%, the predicted values by Yun et al. [21] correlation cover 74% of the experimental data, which revealed a relatively reasonable agreement. The similar conclusions can be found in literature [21] that this correlation caused an average deviation of 11.7%, when compared with 1333 data points. The comparison with above-mentioned correlations suggested that our experimental data had a relatively good convergence at the certain conditions. 4.1.2. Comparison between the experimental data of this work and the cited data The evaporation heat transfer data of R22 in a micro-fin tube by Seo and Kim [25] was compared to those in this study, shown in Fig. 6. The vapor quality variations from the inlet to the outlet are about 0.2–0.8 and 0.1–1, respectively in the cited literature and in this work. From the compared results, most of results had the similar trend within the common mass flux ranges. At the mass fluxes of 100 and 250 kg/m2 s, the conclusions of the measured heat transfer coefficients at 0 °C and 5 °C were different from those of literature [25], because the measurement ranges were changed. When the conditions were changed, the temperature influence on the heat transfer might also be different. For example, the decrease of temperature lowers the density of the two-phase refrigerant, correspondingly the flow rate will be increased and the convective heat transfer is promoted, the explanation supported the results from literature [25]. However, with the mass flux increases, that phenomenon will be restrained, which might
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8000
10000
8000
6000
4000
Tsat= -5-8 ºC 2
G = 100-250 kg/m s 2000
2
q = 28.15-49.29 kW/m
2
G = 175 kg/m s 6000
4000
R161, Tsat= - 5 º C R161, Tsat= 0 º C R161, Tsat= 8 º C
2000
R22, Tsat= - 5 º C R22, Tsat= 0 º C R22, Tsat= 8 º C
0
0 0
2000
4000
6000
8000
10000
12000
2
Experimental heat transfer coefficient, W/m K Fig. 5. Comparison between the present heat transfer coefficient of R22 with those predicted from some correlations.
8000 Seo and Kim (2000) : Tsat= -15 º C, q = 5kW/m
2
2
Average heat transfer coefficient , W/m K
2
q = 38.62 kW/m
2
Local heat transfer coefficient, W/m K
Yun et al.(2002) Cavallini et al. (1999) Kandlikar-Rayoff (1997) Chamra-Mago (2006)
2
Predicted heat transfer coefficient , W/m K
12000
Seo and Kim (2000) : Tsat=
-5 º C,
2
q = 5kW/m
Seo and Kim (2000) : Tsat= 5 º C, q = 5kW/m
2
6000
4000
Present data: Tsat = -5 º C, q = 28.15kW/m
2
2000
Present data: Tsat =
0 ºC,
2
q = 28.15kW/m
Present data: Tsat = 8 º C, q = 28.15kW/m
2
0 100
200
300
2
Mass flux,kg/m s Fig. 6. Comparison between the present heat transfer coefficient of R22 with those obtained by Seo and Kim [16].
lead to the similar results with the literature [27] and our work. Thus, within the same measurement ranges, the comparison results between the cited literature and our data can support the reliability of experimental apparatus to some extent. 4.2. Evaporation heat transfer coefficient Fig. 7 showed the variation of local heat transfer coefficients of R22 and R161 with the vapor quality for various evaporation temperatures when the heat flux and mass flux are fixed at 38.62 kW/m2 and 175 kg/m2 s, respectively. It can be observed that the heat transfer coefficient of R161 grew with the vapor quality, and reaches a maximum at x = (0.7–0.8). The region corresponding to x > (0.7–0.8) was usually referred to the dry-out region. But R22 did not apparently reveal the phenomenon of dry-out in present study. Reasons were that, in the dry-out region of R161, the liquid film had disappeared and heat transfer would be weakened for the low heat conductivity of the vapor, while R22 had significantly higher liquid viscosity than R161, which was beneficial for the formation and preserving of the liquid film, and heat transfer would not be rapidly deteriorated in high quality region. As shown in Fig. 7, the heat transfer coefficient of R22 rose with the decrease
0.0
0.2
0.4
0.6
0.8
1.0
Vapor quality Fig. 7. Local heat transfer coefficient for various saturation temperatures, at q = 38.62 kW/m2 and G = 175 kg/m2 s.
of the saturation temperature during the total quality range. But for R161, the heat transfer coefficient under different vapor qualities were influenced by the saturation temperatures unpredictably. This trend might be supported by Seo and Kim [25], who reported that the increase of heat transfer coefficient with the saturation temperatures was lowered in micro-fin tubes. Kuo et al. [26] concluded that the heat transfer coefficient increased when the saturation temperature rose from 5 to 15 °C in their micro-fin tube. It should be pointed out that the increase of the saturation temperature would lead to the enhancement of the nucleate boiling. But on the other hand, as the saturation temperature decreases, the growth of the two-phase Reynolds number might also boost the heat transfer coefficient [27]. Thus, the influence of saturation temperatures on the heat transfer performance was relatively intricate according to different test fluids and test conditions. Fig. 8 showed the variation of local heat transfer coefficient of R22 and R161 with the vapor quality for different heat flux when the saturation temperature and mass flux are fixed at –5 °C and 175 kg/m2 s, respectively. Generally, the heat transfer coefficient went up with the heat flux. The growth in the heat transfer coefficient with increasing heat flux seemed to be more obvious in low vapor quality region. This might be ascribed to the fact that the nucleate boiling was much stronger than the forced convection at low vapor qualities, and it would be promoted as the imposed heat flux increased. However, when the convective evaporation dominated heat transfer process at high qualities, the effect of the heat flux was very slight. By further observation from Fig. 8, the growth of heat transfer coefficient with increasing heat flux for R161 seemed to be more pronounced than that for R22. Fig. 9 showed the average heat transfer coefficient versus the heat flux for various saturation temperatures at a mass flux of 175 kg/m2 s. Generally, the average heat transfer coefficient showed a higher value at lower saturation temperatures. The average heat transfer coefficient of R161 increased by about 10% when the saturation temperature dropped from 8 °C to 5 °C. But for R22, the influence of saturation temperature on average heat transfer coefficient did not present an obvious regularity at the high heat flux region. Fig. 9 also indicated that the average heat transfer coefficient increased with the heat flux. As the saturation temperature was fixed at 0 °C, the average values of the heat transfer coefficients at 49.29 kW/m2 were about 14% and 5% higher than those at 28.15 kW/m2 for R22 and R161, respectively. For R161, the heat transfer coefficient at lower saturation temperatures was more sensitive to the heat flux. As to R22, Seo et al. [27] reported that for a micro-fin tube, the heat transfer coefficient reduced slightly
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8000
Tsat= -5 °C
Local heat transfer coefficient , W/m K
2
2
2
Local heat transfer coefficient, W/m K
8000
G = 175 kg/m s 6000
4000
2
R161, q =28.15 kW/m 2 R161, q =38.62 kW/m 2 R161, q =49.29 kW/m 2 R22, q =28.15 kW/m 2 R22, q =38.62 kW/m 2 R22, q =49.29 kW/m
2000
6000
4000 2
2000
Tsat = 8 °C 2
q = 28.15 kW/m 0
0
0.0
0.2
0.4
0.6
0.8
0.0
1.0
0.2
0.4
0.6
0.8
1.0
Vapor quality
Vapor quality Fig. 8. Local heat transfer coefficient for various heat flux, at Tsat = 5 °C and G = 175 kg/m2 s.
Fig. 10. Local heat transfer coefficient for various mass flux, at Tsat = 8 °C and q = 28.15 kW/m2.
8000
2
2
Average heat transfer coefficient, W/m K
8000
Average heat transfer coefficient, W/m K
R161, G =100 kg/m s 2 R161, G =175 kg/m s 2 R161, G =250 kg/m s 2 R22, G =100 kg/m s 2 R22, G =175 kg/m s 2 R22, G =250 kg/m s
6000
4000
R161, Tsat =-5 °C R161, Tsat
= 0 °C
R161, Tsat = 8 °C
2
G = 175kg/m s
R22, Tsat =-5 °C
2000
R22, Tsat = 0 °C R22, Tsat = 8 °C
6000
4000
R161, Tsat= -5 °C R161, Tsat= 0 °C R161, Tsat= 8 °C
2
R22, Tsat= -5 °C
q =28.15 kW/m
2000
R22, Tsat= 0 °C R22, Tsat= 8 °C 0
0 30
40
50 2
Heat flux, kW/m
100
200
300
2
Mass flux, kg/m s
Fig. 9. Average heat transfer coefficient versus the heat flux for various saturation temperatures, at G = 175 kg/m2 s.
Fig. 11. Average heat transfer coefficient versus the mass flux for various saturation temperatures, at q = 28.15 kW/m2.
as the heat flux increased. The considerably higher heat flux together with higher saturation temperatures in this study might account for the opposite conclusion. The influence of the heat flux on the heat transfer is very complicated. For one hand, the high heat flux is beneficial for the nucleate boiling at low vapor qualities, and the heat transfer will be promoted, especially at high saturation temperatures; for another, the flow regime transition at a high heat flux is easier than at a low one, and the dry-out flow at high qualities will result in the decrease of average heat transfer coefficients. Fig. 10 presented variation of the local heat transfer coefficients of R22 and R161 with the vapor quality for different mass fluxes at a saturation temperature of 8 °C and a heat flux of 28.15 kW/m2. As expected, the heat transfer coefficient increased strongly with the increase of the mass flux. Most previous studies [28,29] also reported the same trend. Furthermore, the increase of the heat transfer coefficient with the mass flux seemed more pronounced in high quality region. The convective evaporation dominated the heat transfer process at high qualities, and it would be boosted by the higher mass flux. Fig. 11 presented the average heat transfer coefficient versus mass flux for various saturation temperatures at a heat flux of 28.15 kW/m2. As indicated in Fig. 11, the heat transfer coefficient
was strongly dependent on the mass flux. As the saturation temperature was fixed at 0 °C, the average heat transfer coefficients at 250 kg/m2 s were about 72% and 84% higher than those at 100 kg/m2 s, respectively for R22 and R161. Furthermore, the effect of mass flux on the heat transfer coefficient was more significant in the lower mass flux region, and R161 seemed to be more sensitive to the mass flux than R22. Fig. 12 presented the ratio of average heat transfer coefficient between the two refrigerants (R161/R22) versus the heat flux at various saturation temperatures and mass fluxes. According to Fig. 12, the heat transfer coefficient of R161 was significantly enhanced based on R22. The enhancing percentage generally reduced with the increased heat flux and the reduced mass flux, floating between 0% and 50%. Except for the point with a saturation temperature of 8 °C and a mass flux of 100 kg/m2 s, the ratios under different test conditions merged together around 1.23, indicating that the heat transfer coefficient of R161 in present tube was about 23% higher than those of R22. According to the REFPROP [20], R161 had better thermophysical properties than R22 in aspect of the boiling heat transfer. For example, at the saturation temperature of 25 °C,the liquid heat conductivity of R161 is 0.13 W/m K, while that of R22 is only 0.084 W/m K; the vaporization latent heats of R161 and 22 are 338.47 kJ/kg and 182.89 kJ/kg, respectively.
645
2.5
1.8
Tsat =-5 °C, G = 175 kg/m s
Tsat = 0 °C, G = 100 kg/m s
Tsat = 0 °C, G = 175 kg/m s
Tsat = 8 °C, G = 100 kg/m s
Tsat = 8 °C, G = 175 kg/m s
2
2
2
2
2
1.5
Tsat =-5 °C, q = 28.15 kW/m
2
Tsat = -5 °C, G = 100 kg/m s
2
1.23 1.2
0.9
Ratio of pressure drop, R22/R161
Ratio of average heat transfer coefficient, R161/R22
X. Han et al. / International Journal of Heat and Mass Transfer 62 (2013) 638–646
Tsat = 0 °C, q = 28.15 kW/m
2
2.0
Tsat = 8 °C, q = 28.15 kW/m
2
1.65 1.5
1.0 0.88
Tsat = -5 °C, q =38.62 kW/m
2
0.5
Tsat = 0 °C, q = 38.62 kW/m
2
Tsat = 8 °C, q =38.62 kW/m
2
0.0
0.6 30
40
100
50
200
Mass flux, kg/m s
Heat flux, kW/m
10
R22, Tsat =- 5 °C
6
R22, Tsat = 0 °C
Fig. 14. Ratio of pressure drop versus the mass flux, at q = 28.15 and 38.62 kW/m2.
Frictional pressure drop of per unit length, kPa/m
Frictional pressure drop of per unit length, kPa/m
Fig. 12. Ratio of average heat transfer coefficient versus the heat flux, at G = 100 and 175 kg/m2 s.
8
2
q =28.15kW/m
R22, Tsat = 8 °C R161, Tsat = -5 °C R161, Tsat = 0 °C R161, Tsat = 8 °C
4
2
3.5 3.0
2
G =100 kg/m s
R161, Tsat = 0 °C
R22, Tsat = 0 °C
R161, Tsat = 8 °C
2.0 1.5 1.0 0.5 0.0 30
200
R161, Tsat = -5 °C
R22, Tsat = -5 °C R22, Tsat = 8 °C
2.5
0 100
300
2
2
300
40
50 2
Heat flux, kW/m
2
Mass flux, kg/m s Fig. 13. Pressure drop versus the mass flux for various saturation temperatures at q = 28.15 kW/m2.
4.3. Pressure drop Fig. 13 showed the pressure drop of per unit length versus the mass flux for various saturation temperatures as the heat flux fixed at 28.15 kW/m2. Generally, the pressure drop strongly increase with the mass flux both for R22 and R161. Taking R161 with the saturation temperature of 5 °C for example, the pressure drop at G = 175 kg/m2 s was about 4.7 kPa/m, more than twice that at G = 100 kg/m2 s. According to Fig. 11, the pressure drop was also significantly dependent on the saturation temperatures, it dropped with the increasing saturation temperatures. Furthermore, the pressure drop difference in different saturation temperatures increased with the mass flux, and the influence of the mass flux on the pressure drop was reduced for higher saturation temperatures. The increase of the pressure drop with the decreasing saturation temperatures was mainly attributed to the augmentation of the specific volume and the liquid viscosity of the evaporating refrigerant. It still can be found from Fig. 13 that R161 displayed a lower pressure drop compared with R22, but the difference reduced gradually with the mass flux. Even when the mass flux exceeded a certain value (close to 250 kg/m2 s), the pressure drop of R161 became slightly higher than that of R22 at the saturation temperature of 5 and 0 °C.
Fig. 15. Pressure drop versus the heat flux for various saturation temperatures, at G = 100 kg/m2 s.
Fig. 14 showed the ratio of the pressure drop between R22 and R161 versus the mass flux for various saturation temperatures at the heat flux of 28.15 and 38.62 kW/m2. Generally, the ratio dropped with the mass flux, ranging from 0.88 to 1.65. For another, the ratio of the pressure drop was affected by the heat flux, generally the higher heat flux caused a relatively higher growth in the pressure drop. The previous study by Seo et al. [27] suggested that the pressure drop difference between R22 and R410A decreased as the saturation temperature dropped. However, this tendency was not significant in present study for R161 and R22. The difference in the thermophysical properties for R22 and R161, liking the viscosity and specific volume, were deemed to be responsible for the difference of pressure drop. Fig. 15 presented the pressure drop versus heat flux for various saturation temperatures at a mass flux of 100 kg/m2 s, it suggested that the increase in heat flux resulted in a mild increase of the pressure drop. 5. Conclusions R161 is regarded as one of the promising alternative refrigerants of R22. The evaporation heat transfer characteristics of R161 in a horizontal micro-fin tube with an outside diameter
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of 7 mm were investigated by experiment. The heat transfer coefficient and the pressure drop were measured when the heat flux ranged from 28.15 to 49.29 kW/m2, the mass flux ranged from 100 to 250 kg/m2 s, and the saturation temperature ranged from 5 to 8 °C. The heat transfer coefficient of R161 went up with the vapor quality before the dry-out region. The heat transfer coefficient grew with the mass flux and the heat flux, and dropped with the saturation temperature within the experimental conditions. The heat transfer coefficient of R161 was enhanced by about 23% by comparison with R22 on the same test conditions. The pressure drop rose with the mass flux and dropped with the saturation temperature. A rise in the imposed heat flux resulted in a mild increase of the pressure drop. The ratio of pressure drop (R22/R161) ranged from 0.88 to 1.65. The difference in the thermophysical properties of R22 and R161 had an important effect on the enhancement of the heat transfer coefficient and the pressure drop. Acknowledgments This work has been supported by the Nation Natural Science Foundation of China (Grant No. 51176166), and the Program for Key Innovative Research Team of Zhejiang Province (No. 2009R50036), China. We still are very full of appreciation for the help and direction of this project from Prof. Jacobi M. Anthony. References [1] Y.M. Xuan, G.M. Chen, Experimental study on HFC-161 mixture as an alternative refrigerant to R502, Int. J. Refrig. 28 (2005) 436–441. [2] X.H. Han, G.M. Chen, C.S. Li, X.G. Qiao, X.L. Cui, Isothermal vapor–liquid equilibrium data for the binary mixture refrigerant pentafluoroethane (R125) + fluoroethane (R161) at 265.15, 275.15, 283.15, 293.15, 303.15 and 303.15 K with a recirculating still, J. Chem. Eng. Data 51 (4) (2006) 1232–1235. [3] X.H. Han, Q. Wang, Z.W. Zhu, G.M. Chen, Cycle performance study on R32/ R125/R161 as an alternative refrigerant to R407C, Appl. Therm. Eng. 27 (2007) 2559–2565. [4] X.H. Han, Z.W. Zhu, F.S. Chen, Y.J. Xu, Z.J. Gao, G.M. Chen, Solubility and miscibility for the mixture of (Ethyl Fluoride + Polyol Ester Oil), J. Chem. Eng. Data 55 (9) (2010) 3200–3207. [5] X.H. Han, Z.J. Gao, Y.J. Xu, Y. Qiu, X.W. Ming, Q. Wang, G.M. Chen, Isothermal vapor–liquid equilibrium data for the binary mixture difluoromethane (R-32) + ethyl fluoride (R-161) over a temperature range from 253.15 to 303.15 K, Fluid Phase Equilib. 299 (1) (2010) 116–121. [6] X.H. Han, Y.J. Xu, X.W. Min, Z.J. Gao, Q. Wang, G.M. Chen, Density data for the refrigerant ethyl fluoride (HFC-161) over a temperature range from (233 to 343) K, J. Chem. Eng. Data 56 (7) (2011) 3038–3042. [7] X.H. Han, Y. Qiu, P. Li, Y.J. Xu, Q. Wang, G.M. Chen, Cycle performance studies on HFC-161 in a small-scale refrigeration system as an alternative refrigerant to HFC-410A, Energy Build. 44 (2012) 33–38.
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