Experimental study of evaporation heat transfer
characteristics of refrigerants R-134a and R-407C
in horizontal small tubes
Y.M. Lie, F.Q. Su, R.L. Lai, T.F. Lin
*Department of Mechanical Engineering, National Chiao Tung University, Hsinchu 30010, Taiwan, ROC Received 2 December 2004; received in revised form 30 July 2005
Available online 27 September 2005
Abstract
An experiment is carried out here to investigate the characteristics of the evaporation heat transfer for refrigerants R-134a and R-407C flowing in horizontal small tubes having the same inside diameter of 0.83 or 2.0 mm. In the exper-iment for the 2.0-mm tubes, the refrigerant mass flux G is varied from 200 to 400 kg/m2s, imposed heat flux q from 5 to 15 kW/m2, inlet vapor quality xinfrom 0.2 to 0.8 and refrigerant saturation temperature Tsatfrom 5 to 15C. While for
the 0.83-mm tubes, G is varied from 800 to 1500 kg/m2s with the other parameters varied in the same ranges as those for Di= 2.0 mm. In the study the effects of the refrigerant vapor quality, mass flux, saturation temperature and imposed
heat flux on the measured evaporation heat transfer coefficient hrare examined in detail. The experimental data clearly
show that both the R-134a and R-407C evaporation heat transfer coefficients increase almost linearly and significantly with the vapor quality of the refrigerant, except at low mass flux and high heat flux. Besides, the evaporation heat trans-fer coefficients also increase substantially with the rises in the imposed heat flux, refrigerant mass flux and saturation temperature. At low R-134a mass flux and high imposed heat flux the evaporation heat transfer coefficient in the smaller tubes (Di= 0.83 mm) may decline at increasing vapor quality when the quality is high, due to the partial dryout of the
refrigerant flow in the smaller tubes at these conditions. We also note that under the same xin, Tsat, G, q and Di,
refrig-erant R-407C has a higher hrwhen compared with that for R-134a. Finally, an empirical correlation for the R-134a and
R-407C evaporation heat transfer coefficients in the small tubes is proposed. 2005 Elsevier Ltd. All rights reserved.
1. Introduction
In the past decade following the signing of the Mon-treal Protocol in 1996, extensive research has been undertaken to search for the alternatives that can
re-place the chlorofluorocarbons (CFCs) refrigerants. It was well known at that time that the use of the CFCs refrigerants, which contain chlorine and carbon, would lead to the ozone depletion and a consequent in-crease in ultraviolet radiation, and would cause global warming. Some hydrochlorofluorocarbons (HCFCs) and hydrofluorocarbons (HFCs) refrigerants have been developed. The HCFCs refrigerants contain less chlorine than the CFCs and have shorter atmospheric life time. They were considered as interim for the CFCs. The HFCs refrigerants have zero ozone depletion potential
0017-9310/$ - see front matter 2005 Elsevier Ltd. All rights reserved. doi:10.1016/j.ijheatmasstransfer.2005.07.018
*
Corresponding author. Tel.: +886 35 712121 55118; fax: +886 35 726 440.
E-mail address:tflin@mail.nctu.edu.tw(T.F. Lin).
and can replace the CFCs for some period of time. In order to properly use these new refrigerants, we need to understand their thermodynamic, flow and heat transfer properties. In particular, a detailed understand-ing of the characteristics of the evaporation and conden-sation heat transfer for the HFCs refrigerants is very important in the design of evaporators and condensers for many refrigeration and air conditioning systems.
Recently, there is growing interest in the use of ultra-compact heat exchangers in various thermal systems be-cause of their very high heat transfer density. Thus the flow and heat transfer characteristics in small and capil-lary tubes have been extensively investigated for various fluids such as air, water and some refrigerants[1]. But the heat transfer characteristics associated with the evaporation and condensation of the HFCs refrigerants in the small tubes are less explored. Data for these two-phase heat transfer coefficients in the small tubes are still scarce. In the present study we conduct experiments to measure the evaporation heat transfer coefficients of the HFCs refrigerants R-134a and R-407C in small tubes. We choose these two refrigerants in this study be-cause they are regarded as the major substitutes for refrigerants R-12 and R-22.
In what follows the relevant literature on the present study is briefly reviewed. Yu et al.[2]recently examined flow boiling heat transfer for water in a 2.98-mm diam-eter channel and found that the measured boiling heat transfer coefficients for the vapor quality above 0.5
depended on the heat flux but were independent of the water mass flux. They concluded that the nucleate boil-ing was dominant over the convective boilboil-ing in small channels. The results are significantly different from these for the conventional channels, where the mass flux effects can be substantial. Similarly, Sumith et al. [3]
measured the saturated flow boiling heat transfer and pressure drop of water in a test section made of a stain-less steel tube with an inner diameter of 1.45 mm. They indicated that the dominant flow pattern in the tube was a slug-annular or an annular flow, and liquid film evap-oration dominated the heat transfer.
An experiment carried out by Yan and Lin [4,5]to study the evaporation heat transfer and pressure drop of R-134a in a tube bank forming by 28 small side-by-side contacting pipes (Di= 2.0 mm) revealed that both
the refrigerant mass flux and imposed heat flux were important and the evaporation heat transfer in the small pipes was significantly higher than that in large tubes. A similar study for subcooled flow boiling of R-134a in a vertical multiport parallel rectangular channels (Dh= 2.01 mm) was carried out by Agostini and
Bon-temps [6]. In a visualization investigation Nino et al.
[7]examined R-134a flow boiling in a multiport mini-channel tube with Dh= 1.5 mm. They proposed a
meth-od to describe the fraction of time or the probability that a flow pattern existed in a particular flow condition. A recent study from Fujita et al.[8]for R-123 boiling in a horizontal small tube with an inside diameter of Nomenclature
As inside surface area of the small tubes, m 2
Bo Boiling number, Bo¼Giq
fg, dimensionless
Dh hydraulic diameter, mm
Di diameter of small tube, mm
f friction factor, dimensionless g gravitational acceleration, m/s2
G mass flux, kg/m2s
hl single-phase liquid heat transfer coefficient,
W/m2C
hr evaporation heat transfer coefficient, W/
m2C
ifg enthalpy of vaporization, J/kg
k thermal conductivity, W/mK Nconf confinement number, Nconf¼
r gðqlqgÞ
h i0:5
Dh ,
dimensionless
Nul Nusselt number for liquid flow, Nul¼hlkDli,
dimensionless
Nur Nusselt number for evaporation, Nur¼hrkDli,
dimensionless
Pr Prandtl number, dimensionless q average imposed heat flux, W/m2
Q heat transfer rate, W Re Reynolds number, Re¼GDi
ll , dimensionless
T temperature,C
Tr,sat saturated temperature of refrigerant,C
x vapor quality Greek symbols
Dx total quality change in the small tubes l viscosity, N s/m2
q density, kg/m3 r surface tension, N/m Subscripts
g vapor phase
in at inlet of the test section
l liquid phase
n net power input to the refrigerant R-134a or R-407C
r refrigerant sat saturation
1.12 mm suggested that heat transfer in the flow was dominated by the nucleate boiling and the effects of the refrigerant mass flux and vapor quality to the boiling heat transfer were very weak. A similar study for R-113 boiling conducted by Lazarek and Black[9] noted the negligible variation of the boiling heat transfer coeffi-cient with the local vapor quality, which implied that the wall heat transfer processes were again controlled by nucleate boiling. In a vertical small tube (Di= 1 mm)
with refrigerant R-141b flowing in it, Lin et al.[10]found that at low quality, nucleate boiling dominated. But at higher quality, convection dominated. In a further study
[11], they examined the same refrigerant in four circular tubes with diameters 1.1, 1.8, 2.8, 3.6 mm and one square tube of cross-section 2· 2 mm2. Their results indicate
that the mean heat transfer coefficient in a tube or chan-nel is independent of the mass flux and tube diameter but is a function of the imposed heat flux. Cornwell and Kew
[12]observed boiling in refrigerant R-113 flow and mea-sured the heat transfer for two geometries: one had 75 channels with Dh= 1.03 mm and other had 36 channels
with Dh= 1.64 mm. Their experimental work suggested
the presence of three two-phase flow patterns in the channels: isolated bubble, confined bubble and annu-lar-slug flow. In a continuing study [13], they investi-gated refrigerant R-141b boiling in a horizontal tube with its inner diameter ranging from 1.39 to 3.69 mm and proposed that flow boiling in a narrow channel might be through one of four mechanisms: nucleate boil-ing, confined bubble boilboil-ing, convective boiling and par-tial dry-out. They further indicated that, except at very low heat flux, the boiling showed a strong dependence on the heat flux, a weak dependence on the mass flux, and independence of the quality. Besides, they intro-duced a new dimensionless group named as the confine-ment number Nconf, which represented the importance of
the restriction of the flow by the small size of the chan-nel. The dimensionless number Nconfcan be used to find
the transition from the isolated to confined bubble re-gimes. To a first approximation, they showed that the confined boiling occurred when Nconf> 0.5.
By examining boiling of refrigerants in a small cir-cular tube (Di= 2.46 mm) and a rectangular duct
(Dh= 2.40 mm) with nearly the same hydraulic
diame-ters, Tran et al.[14]showed that there was no signifi-cant geometry effect for the two channels tested. Furthermore, their results imply that the nucleation mechanism dominates over the convection mechanism in small-channel evaporators over the full range of quality (0.2–0.8), which is contrary to the situations in larger channels where the convection mechanism dominates at qualities typically above 0.2. Bao et al.
[15]also found that the boiling heat transfer coefficient was a strong function of the heat flux and system pres-sure in a 1.95 mm diameter tube with R-11 and R-123, while the effects of the mass flux and vapor quality
were very small, suggesting that the heat transfer was mainly through the nucleate boiling. Wambsganss et al. [16] studied boiling heat transfer of refrigerant R-113 in a small diameter (2.92 mm) tube and evalu-ated 10 different heat transfer correlations. They found that the high boiling number and slug flow pattern led to the domination by the nucleation mechanism and the two-phase correlations based on this dominance also predicted the data best. Moreover, Warrier et al.
[17] tested FC-84 in five parallel channels with each channel having a hydraulic diameter of 0.75 mm and compared their results with five widely used correla-tions. They then proposed two new correlations, one for subcooled flow boiling heat transfer and another for saturated flow boiling heat transfer. Oh et al. [18]
examined R-134a in capillary tubes of 500-mm long and 2, 1 and 0.75-mm inside diameters. They concluded that the heat transfer in the forced convective boiling was more influenced by the refrigerant mass flux than by the boiling number and the heat transfer coefficient was controlled by the Reynolds number of the flow. Besides, their results also showed that the dry-out point moved to the lower quality with decreasing size of the tubes. Vaporization of CO2 in 25 flowchannels each
having an inside diameter of 0.8 mm was recently examined by Pettersen [19]. He also observed the two-phase flow pattern with another separate test rig with a 0.98-mm heated glass tube. The results showed that nucleate boiling dominated at low/moderate vapor fractions, where the boiling heat transfer coefficient in-creased with the heat flux and refrigerant saturated temperature but was less affected by the mass flux and vapor fraction. Moreover, the dryout effects be-came very important at higher mass flux and tempera-ture, where the boiling heat transfer coefficient dropped rapidly at increasing quality. And the two-phase flow was in intermittent and annular flow regimes, the latter becoming more important at high mass flux. More complete information on various aspects of the two phase flow and boiling heat transfer in ultra-compact evaporators and microchannels is available from the re-cent critical review conducted by Kim et al.[1], Ghiaa-siaan and Abdel-khalik [20], Thome [21], Sobhan and Garimella [22], Kandlikar [23,24], Watel [25], and Vlasie et al.[26].
The above literature review clearly indicates that the experimental data for the evaporation heat transfer of the HFC refrigerants in small tubes are still in urgent need. To complement our earlier investigations[4,5], in this study we move further to measure the evaporation heat transfer coefficients of refrigerants R-134a and R-407C in horizontal small tubes of inside diameter 2.0 and 0.83 mm. The effects of the vapor quality, refrig-erant mass flux, imposed heat flux and system pressure on the evaporation heat transfer in the small tubes will be examined in detail.
2. Experimental apparatus and procedures
The experimental system modified slightly from that used in the previous study[4]is employed here to inves-tigate the evaporation heat transfer of the HFC refriger-ants in small tubes. It is schematically depicted inFig. 1. The experimental apparatus consists of three main loops, namely, a refrigerant loop, a water-glycol loop, and a hot-water loop. The refrigerant R-134a or R-407C is circulated in the refrigerant loop. In order to control various test conditions of the refrigerants in the test section, we need to control the temperature and flow rate in the other two loops. The detailed description of the apparatus is available from our earlier study[4]. Here only the modified test section employed in the experiment is described in detail.
The modified test section along with the entry and exit sections attached to it are schematically shown in
Fig. 2. Due to the tubes to be tested being relatively small, the refrigerant flow rates in them are very low and direct measurement of evaporation heat transfer coefficient in the tubes is difficult and can be subject to large error. Thus 28 small tubes all made of copper, each having the same diameter and length, are put together side by side to form a plane tube bundle acting as the
test section, as shown in Fig. 2. Each small tube has the same diameter of 0.83 or 2.0 mm, outside diameter of 1.83 or 3.0 mm, and length of 150 mm. In order to allow the refrigerant to flow smoothly into the small tubes, a section including divergent, convergent and straight portions is connected to the inlets of the tubes. Besides, another section including straight and conver-gent parts is attached to the exits of the tubes. Both the entry and exit sections are formed by the stainless steel plates. Note that the addition of the entry and exit sections in the present study is expected to improve the flow distribution among the tubes in the bundle[5]. At the middle axial location of the small tubes 14 thermo-couples are soldered onto the outer surface of the tubes. Specifically, these thermocouples are soldered onto 14 selected tubes at the circumferential position of 45 from the top of the tube or from the bottom of the tube, as shown inFig. 3. Two copper plates of 5-mm thick are respectively soldered onto the upper and lower sides of the tube bundle also shown inFig. 3. The copper plates are heated directly by an electric-resistance heater of 2.6-mm wide, 0.5-2.6-mm thick and 2.5-m long. The heater is connected to a 500 W DC power supply. Mica sheet is placed in the narrow space between the heater and cop-per plates to prevent the electric current leaking to the
TEST SECTION FLOW DIRECTION T T P TTP -50V - 30A DC Power Supper + DEGASED VALVE PUMP THERMOCOUPLE PRESSURE TAP P T DPDIFFERENTIALPRESSURE SIGHT GLASS VALVE RELEASE VALVE WATER LOOP
WATER FLOW METER
WATER-GLYCOL LOOP REFRIGERANT LOOP PUMP SUBCOOLER CONDENSER THERMOSTAT WATER-GLYCOL
MASS FLOWMETER PUMP
THERMOSTAT WATER BY-PASS FLOW FILTER/DRYER PREHEATER N2 ACCUMULATO R BY- PASS F L OW RECEIVER TOWER COOLING T T T T
copper plates. The power input to the heater is measured by a power meter with an accuracy of ±0.5%. In order to reduce the heat loss from the heaters, the whole test section is wrapped with a 10-cm thick polyethylene layer. It should be noted that the heated section of the
tube bundle is only 100-mm long and there are two un-heated sections each having 25-mm in length upstream and downstream of the heated section. Axial heat con-duction in the tube walls can be important in affecting the measured evaporation heat transfer coefficient in
Inlet Section SectionTest Exit Section
Units : mm FLOW
Fig. 2. Schematic diagram of test section along with the inlet and exit sections.
A-A CROSS SECTIONAL VIEW
THERMOCOUPLE TUBES MICA SHEET COPPER PLATE A A HEATER HEATER (a) (b) 100 mm 150 mm 1 mm 5 mm Do Di=0.83 or 2.0 mm Do=1.83 or 3.0 mm
view of the thermal conductivity of the copper being much higher than that of R-134a and R-407C. In the present study, however, the liquid refrigerant flow in the small tubes is at a very high Peclet number (>5000). Thus the conjugation effects between the con-vection in the flow and conduction in the tube walls are expected to be small, as evident from our early stud-ies[27,28].
Before a test is started, the system temperature is compared with the saturation temperature of refrigerant R-134a or R-407C corresponding to the measured satu-ration pressure of the refrigerant and the allowable dif-ference is kept in the range of 0.2–0.3 K. Otherwise, the system is evacuated and then charged to re-move some noncondensible gases possibly existing in the refrigerant loop. In each test the liquid refrigerant leaving the subcooler is first maintained at a specified temperature by adjusting the water-glycol temperature and flow rate. In addition, we adjust the thermostat in the water loop to stabilize the refrigerant temperature at the test section inlet. Next, the temperature and flow rate of the water loop for the preheater are adjusted to keep the vapor quality of R-134a or R-407C at the test section inlet at the desired value. Then, we regulate the refrigerant pressure at the test section inlet by adjusting the gate valve locating right after the exit of the test sec-tion. Meanwhile, by changing the current of the DC motor connecting to the refrigerant pump, the refriger-ant flow rate can be varied. The imposed heat flux from the heater to the refrigerant is adjusted by varying the electric current delivered from the DC power supply to the heater. By measuring the current delivered to and voltage drop across the heater, we can calculate the heat transfer rate to the refrigerant. All tests are run when the experimental system has reached statistically steady state. Finally, all the data channels are scanned every 5 s for a period of 50 s.
3. Data reduction
Before the two-phase experiments, the total heat loss from the test section is evaluated by comparing the total power input from the power supply with that calculated from the energy balance in the single phase refrigerant flow. The measured results indicated that for all runs in the energy balance test the heat loss was within 2%. The average single-phase liquid refrigerant convection heat transfer coefficient in the small tubes is defined as hl¼
Qn
As ðTwall Tr;aveÞ
ð1Þ Here Qnis the net power input to the liquid refrigerant
R134a or R-407C, Asis the total inside surface area of
the small tubes in the test section, Twallis the average
of the measured tube wall temperatures at all detected
locations, and Tr,aveis the average refrigerant
tempera-ture in the tubes, which in turn is estimated from the measured refrigerant temperatures at the inlet and exit of the test section as (Tr,i+ Tr,o)/2.
The vapor quality of refrigerant R-134a or R-407C entering the test section is evaluated from the energy bal-ance for the preheater. The total change of the refriger-ant vapor quality in the test section is deduced from the net heat transfer rate from the heater to the refrigerant in the test section. Finally, the average heat transfer coefficient for the evaporation of refrigerant R-134a or R-407C in the test section is determined from the definition
hr
Qn
AsðTwall Tr;satÞ
ð2Þ More detailed description of the data reduction is available from our earlier study [4]. Uncertainties of the measured heat transfer coefficients are estimated according to the procedures proposed by Kline and McClintock[29]. The detailed results from this uncer-tainty analysis are summarized inTable 1.
4. Results and discussion
In what follows selected data obtained here are pre-sented to illustrate the evaporation heat transfer of R-134a or R-407C in the horizontal small circular tubes. The present experiments are performed for refrigerant R-134a or R-407C in the tube bank forming by the 2.0-mm diameter tubes with the refrigerant mass flux G varied from 200 to 400 kg/m2s, imposed heat flux q from 5 to 15 kW/m2, inlet vapor quality xin from 0.2
to 0.8, and refrigerant saturated temperature Tsatfrom
5 to 15C. While for the other tube bank forming by
Table 1
Summary of the uncertainty analysis
Parameter Uncertainty
Small tubes geometry
Length, width and thickness (%) ±1.5
Area (%) ±3.0
Parameter measurement
Temperature, T (C) ±0.2
Temperature difference, DT (C) ±0.4
System pressure, P (kPa) ±2
Mass flux of refrigerant, G (%) ±5
Single-phase heat transfer in small tubes
Imposed heat flux, q (%) ±4.5
Heat transfer coefficient, hr,l(%) ±12.0
Evaporation heat transfer in small tubes
Imposed heat flux, q (%) ±4.5
Inlet vapor quality, xin(%) ±9.5
the 0.83-mm diameter tubes, G is varied from 800 to 1500 kg/m2s with the other parameters varied in the same ranges as those for Di= 2.0 mm. Note that the
dif-ferent ranges of the refrigerant mass flux are chosen for the different sizes of the tubes. Since at the low mass flow rate the evaporating refrigerant flow in the smaller tubes for Di= 0.83 mm is somewhat unsteady, leading to the
unstable intermittent flow in the system. In the following the effects of the vapor quality, imposed heat flux and refrigerant mass flux and saturated temperature on the R-134a and R-407C evaporation heat transfer coeffi-cients are to be examined in detail.
4.1. Single phase heat transfer
Before measuring the R-134a and R-407C evapora-tion heat transfer coefficients, single-phase liquid R-134a and R-407C convection heat transfer coefficients in the small tubes were obtained first for the refrigerant inlet temperature fixed at 15C and imposed heat flux of 5 kW/m2with the refrigerant mass flux respectively
var-ied from 200 to 800 kg/m2s (corresponding to Rel
rang-ing from 1786 to 9227) and from 400 to 3000 kg/m2s
(corresponding to Relranging from 1482 to 14,360) in
the 2.0-mm and 0.83-mm diameter tubes. Note that the single-phase heat transfer test is conducted here to check the energy balance in the test section and the suitability of the experimental system for the present measurement. The measured single-phase liquid R-134a and R-407C heat transfer coefficients are compared with the correlations from Dittus–Boelter[30]and Gnielinski
[31]inFigs. 4 and 5.
The well known Dittus–Boelter correlation is Nul¼ 0:023 Re0:8l Pr
0:4
l ð3Þ
applicable for Rel> 105and 0.7 < Prl< 16,700 and the
Gnielinski correlation is Nul¼
ðf =2ÞðRel 1000ÞPrl
1:07þ 12:7pffiffiffiffiffiffiffiffif =2ðPr2=3l 1Þ ð4Þ applicable for 2300 < Rel< 106 and 0.6 < Prl< 105,
where
f¼ ð1:58 ln Rel 3:28Þ2 ð5Þ
It is of interest to note from the results inFigs. 4 and 5that for the 2.0-mm tubes the measured single-phase heat transfer coefficients are close to the Dittus–Boelter correlation. While for the 0.83-mm tubes the data are well fitted with the Gnielinski correlation. Finally, in the single-phase heat transfer tests to check the energy balance in the test section the relative heat loss is found to be within 2% for all runs. Thus the heat loss from the test section is small.
4.2. Evaporation heat transfer in 2.0-mm tubes
The measured heat transfer data for the R-134a and R-407C evaporation in the 2.0-mm tubes are presented first.Figs. 6 and 7respectively illustrate how the refrig-erant saturated temperature, mass flux and imposed heat flux affect the heat transfer coefficients for R-134a and R-407C evaporation in the 2.0-mm tubes. The measured heat transfer coefficients are examined by checking their variations with the vapor quality at the test section inlet xin. Since the tubes are short, the total quality change Dx
between the inlet and exit of the test section is relatively
0 200 400 600 800 1000 G (kg/m2s) 0 1000 2000 3000 hl (W/m 2 oC )
Single-Phase Heat Transfer in 2.0mm small tubes R-407C R-134a D.B. for R-407C ---Gn. for R-407C D.B. for R-134a ---Gn. for R-134a
Fig. 4. Comparison of the present data for the liquid R-134a and R-407C heat transfer coefficients in 2.0-mm small tubes with the Dittus–Boelter and Gnielinski correlations.
0 1000 2000 3000 4000 G (kg/m2s) 0 2000 4000 6000 8000 10000 hl (W/m 2 oC)
Single-Phase Heat Transfer in 0.83mm Small tubes R-407C R-134a D.B. for R-407C --- Gn. for R-407C D.B. for R-134a --- Gn. for R-134a
Fig. 5. Comparison of the present data for the liquid R-134a and R-407C heat transfer coefficients in 0.83-mm small tubes with the Dittus–Boelter and Gnielinski correlations.
small, ranging from 0.01 to 0.03 in the present study. The results given inFigs. 6 and 7show that for given q, Tsat and G the evaporation heat transfer coefficients
for both R-134a and R-407C increase almost linearly with the inlet quality. Moreover, at higher Tsat, q and
G the increase of hrwith xinis more significant. The
sig-nificant increase of the evaporation heat transfer coeffi-cients with the inlet vapor quality is considered to result from the fact that at a higher inlet quality the li-quid film thickness of the refrigerant on the inside sur-face of the tubes becomes thinner. Hence the thermal resistance of the liquid film is reduced and heat transfer
across the film is improved. Besides, at a higher vapor quality the mass flux of the vapor and the velocity of vapor flow are faster. This also improves the interfacial heat transfer. More specifically for the case with R-134a at Tsat= 15C, G = 400 kg/m
2
s and q = 15 kW/m2, an increase of 31% in hroccurs for xinraised from 0.2 to 0.8
(Fig. 6(a)). The effects of each parameter on hrare
exam-ined in the following. At first, the effects of the saturated temperature Tsat are presented inFig. 6(a) by showing
the variations of the R-134a evaporation heat transfer coefficients with the inlet vapor quality at Tsat= 5, 10,
15C for given refrigerant mass flux and imposed heat flux. The results in Fig. 6(a) indicate that at fixed q hr (W/m 2 oC) hr (W/m 2 oC) hr (W/m 2 oC) 3000 4000 5000 6000 7000
Evaporation Heat transfer Coefficient of R-134a (Di = 2.0 mm) at G =400 k g/m2s, q= 15 kW/m2 Tsat=5 o C Tsat=10 o C Tsat=15 o C 3000 4000 5000 6000 7000
Evaporation Heat transfer Coefficient of R-134a (Di = 2.0 mm) at Tsat=15 o C, q=15 kW/m2 G =200 kg/m2s G =300 kg/m2s G =400 kg/m2s 0 0.2 0.4 0.6 0.8 1 xin 0 0.2 0.4 0.6 0.8 1 xin 0 a b c 0.2 0.4 0.6 0.8 1 xin 2000 4000 6000 8000
Evaporation Heat transfer Coefficient of R-134a (Di = 2.0 mm), at Tsat=15oC, G=400 kg/m2s
q= 5 kW/m2
q= 10 kW/m2
q= 15 kW/m2
Fig. 6. Variations of R-134a evaporation heat transfer coeffi-cient with inlet vapor quality in 2.0-mm small tubes: (a) for various Tsatat G = 400 kg/m2s and q = 15 kW/m2, (b) for var-ious G at Tsat= 15C and q = 15 kW/m2, and (c) for various q at Tsat= 15C and G = 400 kg/m2s. 0 a b c 0.2 0.4 0.6 0.8 1 xin 0 0.2 0.4 0.6 0.8 1 xin 0 0.2 0.4 0.6 0.8 1 x in 4000 6000 8000 10000 12000 14000
Evaporation Heat transfer Coefficient of R-407C (Di = 2.0 mm) at G =400 k g/m2s, q =15 kW/m2 Tsat=5 o C Tsat=10 o C Tsat=15 o C 3000 5000 7000 9000 11000
13000 Evaporation Heat transfer Coefficient of R-407C (Di =2.0 mm) at Tsat=15 o C, q=15 kW/m2 G =200 kg/m2s G =300 k g/m2s G =400 k g/m2s 2000 4000 6000 8000 10000 12000
14000 Evaporation Heat transfer Coefficient of R-407C (Di = 2.0 mm) at Tsat=15 o C, G =400 kg/m2s q=5 kW/m2 q=10 kW/m2 q=15 kW/m2 hr (W/m 2 oC) hr (W/m 2 oC) hr (W/m 2 oC)
Fig. 7. Variations of R-407C evaporation heat transfer coeffi-cient with inlet vapor quality in 2.0-mm small tubes: (a) for various Tsatat G = 400 kg/m2s and q = 15 kW/m2, (b) for var-ious G at Tsat= 15C and q = 15 kW/m2, and (c) for various q at Tsat=15C and G = 400 kg/m2s.
and G the R-134a evaporation heat transfer coefficients rise with the saturated temperature of the refrigerant. A similar trend is also noted by Agostini and Bontemps[6]. This increase in hrwith Tsatis ascribed to the fact that at
a higher Tsatthe latent heat of vaporization ifgis lower,
which in turn results in a higher evaporation rate of the liquid R-134a for a fixed q. Hence the R-134a vapor in the tubes flows at a higher speed, producing a higher convection effect and therefore a higher hr. To be more
quantitative on the effects of Tsaton the R-134a
evapo-ration heat transfer coefficient, the quality-averaged evaporation heat transfer coefficients hr at G = 400 kg/
m2s and q = 15 kW/m2are calculated from the data in
Fig. 6(a). For Tsat raised from 5C to 15 C, hr is
in-creased by 19%.
Next, the effects of the refrigerant mass flux on the R-134a evaporation in the 2.0-mm tubes are shown in
Fig. 6(b). The results indicate that the increase of hrwith
the R-134a mass flux is rather significant, suggesting that the interfacial evaporation is effectively enhanced by the rise in the refrigerant mass flux. Hence the convection mechanism is important in the flow. Quantitatively according to the data inFig. 6(b) for Tsat= 15C and
q = 15 kW/m2, the quality-averaged evaporation heat
transfer coefficient is increased by 27% for G raised from 200 to 400 kg/m2s for R-134a.
Then, the results presented inFig. 6(c) indicate that the R-134a evaporation heat transfer coefficient in-creases rather significantly with the imposed heat flux. This significant increase of hr with q reflects that the
evaporation at the liquid–vapor interface in the refriger-ant flow is substrefriger-antially augmented by the increase in the imposed heat flux. According to the data inFig. 6(c) for Tsat= 15C and G = 400 kg/m2s, with q raised from 5
to 15 kW/m2hr is increased by 44%.
Checking with the heat transfer data given inFig. 7
for R-407C, we note that the effects of the vapor quality, refrigerant saturated temperature and mass flux, and im-posed heat flux on the evaporation heat transfer coeffi-cients of R-407C are qualitatively similar to those for R-134a. A close inspection of the data in Figs. 6 and 7, however, reveals some differences. The R-407C evap-oration heat transfer coefficient is noticeably higher. Be-sides, the effects of Tsat, G and q on hrfor R-407C are
stronger for most cases. Moreover, for R-134a the evap-oration heat transfer coefficient varies with G more sig-nificantly at a high refrigerant mass flux (Fig. 6(b)). While the opposite is the case for R-407C, as evident fromFig. 7(b).
4.3. Evaporation heat transfer in the smaller tubes (Di= 0.83 mm)
The measured heat transfer data for the smaller tubes with Di= 0.83 mm are illustrated in Fig. 8 for the
R-134a evaporation and inFig. 9for the R-407C
evap-oration, covering the effects of various parameters on the evaporation heat transfer coefficients for these two refrigerants. First, it is noted from the results inFig. 8
that for given Tsat, G and q the R-134a evaporation heat
transfer coefficient also increases noticeably with the inlet vapor quality except at high vapor quality for some cases at low G and high q (Fig. 8(b)). A close inspection of data inFig. 8reveals that at low mass flux and high imposed heat flux, hr even decreases with a rise in xin
at a high vapor quality for xin> 0.6. This is conjectured
to result from the partial dryout of the refrigerant on the tube wall at high xinat these conditions. In fact, the data
0 a b c 0.2 0.4 0.6 0.8 1 xin 0 0.2 0.4 0.6 0.8 1 xin 0 0.2 0.4 0.6 0.8 1 xin 3000 4000 5000 6000 7000 8000 hr (W/m 2 o C) hr (W/m 2 o C) hr (W/m 2 o C)
Evaporation Heat transfer Coefficient of R-134a (Di=0.83 mm) at G=1500 kg/m2s, q=15 kW/m2 Tsat=5 oC Tsat=10 oC Tsat=15 oC 3000 5000 7000 9000
Evaporation Heat transfer Coefficient of R-134a (Di = 0.83 mm) at Tsat=15 oC, q=15 kW/m2 G =800 k g/m2s G =1150 kg/m2s G =1500 kg/m2s 3000 4000 5000 6000 7000 8000
Evaporation Heat transfer Coefficient of R-134a (Di = 0.83 mm) at Tsat=15 o C, G =1500 kg/m2s q= 5 kW/m2 q=10 kW/m2 q=15 kW/m2
Fig. 8. Variations of R-134a evaporation heat transfer coefficient with inlet vapor quality in 0.83-mm small tubes: (a) for various Tsat at G = 1500 kg/m
2
s and q = 15 kW/m2, (b) for various G at Tsat= 15C and q = 15 kW/m
2
, and (c) for various q at Tsat= 15C and G = 1500 kg/m
2 s.
from Yan and Lin[4]for R-134a evaporation in the 2.0-mm tubes show the decline of hrwith a rise in xin for
many cases. These earlier poor data result from the sig-nificant refrigerant dryout in the tubes, as already men-tioned here. This is apparently due to the bad refrigerant mass flux distribution at the inlet of the previous test sec-tion. In the present study, this bad flow distribution has been improved by modifying the entry and exit sections of the test section. Hence the partial refrigerant dryout only occurs in the smaller tubes with Di= 0.83 mm at
high xinand q and low G (Fig. 8(b)). Moreover, at an
intermediate quality the increase of hrwith xinis rather
large for the cases at high Tsat, q and G. However, for
these cases the increase is rather mild at low quality. Quantitatively for the case with Tsat= 15C, G = 1500
kg/m2s and q = 15 kW/m2, the increase in hr is 24%
for xinraised from 0.2 to 0.8 (Fig. 8(a)).
According to the results inFig. 9, the R-407C evap-oration heat transfer coefficients increase almost linearly with the inlet vapor quality for most cases. The increase is also rather substantial at high Tsat, q and G. Checking
the numerical values for hrinFig. 9reveals the
quanti-tative increase of hrwith xinfor R-407C. For example,
at Tsat= 15C, G = 1500 kg/m2s and q = 15 kW/m2
we have a much smaller hrincrease of 11% for R-407C
for the same rise in xin(Fig. 9(a)).
An overall inspection of the data presented inFig. 8
discloses that for R-134a evaporation in the smaller tubes with Di= 0.83 mm the evaporation heat transfer
in the flow can be significantly increased by raising the refrigerant saturated temperature, mass flux and heat flux. These trends are similar to that for the large tubes with Di= 2.0 mm already examined above. A
fur-ther inspection of the measured data for R-134a evap-oration in the smaller tubes reveals that the effects of the refrigerant saturated temperature and mass flux on hrare more pronounced at a higher heat flux.
Final-ly, the heat flux variation on hr is more important at
high Tsatand G.
For R-407C evaporation in the smaller tubes a sub-stantial increase in hr with the increasing refrigerant
saturated temperature, mass flux and heat flux is also noted from the results in Fig. 9. It should be men-tioned that for R-407C evaporation at high vapor quality hrdoes not decline for a rise in xin, unlike that
for R-134a. This is attributed to the fact that R-407C has a high latent heat of vaporization and the partial dryout on the tube wall is less likely to occur in the R-407C evaporation. Moreover, in the smaller tubes the effects of Tsat, G and q on the evaporation heat
transfer coefficients for R-407C are also stronger than that for R-134a. Again in the small tubes R-407C has a higher hr.
To be more quantitative on the effects of various parameters on the heat transfer data in the smaller tubes, the quality-averaged evaporation heat transfer coefficients for various cases are evaluated. The results from this evaluation show that for R-134a at G = 1500 kg/m2s and q = 15 kW/m2, hr experiences a 21%
increase when Tsat is raised from 5C to 15 C
(Fig. 8(a)). While for R-407C the corresponding increase is 37% (Fig. 9(a)). Next, we note that for R-134a at Tsat= 15C and q = 15 kW/m2, there is a 41% increase
in hr for G raised from 800 to 1500 kg/m 2
s (Fig. 8(b)). The corresponding increase for R-407C is 48% (Fig. 9(b)). Finally, a 33% increase in hr is noted for q
raised from 5 to 15 kW/m2 for R-134a at T
sat= 15C
and G = 1500 kg/m2s (Fig. 8(c)). For R-407C the corre-sponding increase is 67% (Fig. 9(c)).
hr (W/m 2 o C) hr (W/m 2 o C) hr (W/m 2 o C) 5000 7000 9000 11000 13000
Evaporation Heat transfer Coefficient of R-407C (Di = 0.83 mm) at G =1500 kg/m2s, q =15 kW/m2 Tsat=5 o C Tsat=10 o C Tsat=15 o C 0 0.2 0.4 0.6 0.8 1 xin 0 0.2 0.4 0.6 0.8 1 xin c b a 0.2 0 0.4 0.6 0.8 1 xin 5000 7000 9000 11000
13000 (D Evaporation Heat transfer Coefficient of R-407C
i=0.83 mm) at Tsat=15 o C, q=15 kW/m2 G=800 kg/m2s G=1150 kg/m2s G=1500 kg/m2s 4000 6000 8000 10000 12000 14000
Evaporation Heat transfer Coefficient of R-407C (Di = 0.83 mm) at Tsat=15 oC, G =1500 k g/m2s
q= 5 kW/m2
q=10 kW/m2
q=15 kW/m2
Fig. 9. Variations of R-407C evaporation heat transfer coeffi-cient with inlet vapor quality in 0.83-mm small tubes: (a) for various Tsat at G = 1500 kg/m
2
s and q = 15 kW/m2, (b) for various G at Tsat= 15C and q = 15 kW/m
2
, and (c) for various q at Tsat= 15C and G = 1500 kg/m
2 s.
4.4. Correlation equation for evaporation heat transfer coefficients
For practical application the present data for the R-134a and R-407C evaporation in the 2.0-mm and 0.83-mm tubes need to be correlated empirically. The data presented above indicate that hrvaries linearly with
the vapor quality for most cases and the correlation is thus expressed as
Nur
hr Di
kl
¼ m1xinþ m2 ð6Þ
where m1and m2can be correlated as
m1¼ f ðBo; ReÞ ¼ a1þ b1Boc1Red1 ð7Þ
m2¼ f ðBo; ReÞ ¼ a2Bob2Rec2 ð8Þ
The values for the coefficients in m1 and m2 are a1=
1.39, b1= 1.87· 103, c1= 1.82, d1= 3.14, a2= 28.6,
b2= 0.706, c2= 0.888. The Boiling number and
Reynolds number are defined respectively as Bo¼ q G ifg ð9Þ Re¼ G Di ll ð10Þ
Comparison of the above correlation with the present experimental data shown inFig. 10indicates that more
than 80% of the present data for hr fall within ±35%
of Eq.(6), and the mean absolute error (MAE) between the present data for hrand the proposed correlation is
19%.
5. Concluding remarks
Experiments have been conducted here to investigate the evaporation heat transfer of R-134a and R-407C in the small tubes with Di= 0.83 and 2.0 mm. The effects
of the refrigerant saturated temperature, mass flux, imposed heat flux, and vapor quality of R-134a and R-407C on the evaporation heat transfer coefficients have been examined in detail. The results show that the R-134a and R-407C evaporation heat transfer coef-ficients in the small tubes increase almost linearly with the vapor quality and the increases are significant except at low imposed heat flux and low refrigerant mass flux. Moreover, the increases of the R-134a and R-407C evaporation heat transfer coefficients in both tubes with the imposed heat flux, refrigerant mass flux and satu-rated temperature are also substantial. Besides, the evaporation heat transfer coefficients for R-134a are noticeably lower than that for R-407C at the same Tsat,
G and q. Furthermore, for R-134a in the smaller tubes (Di= 0.83 mm) partial refrigerant dryout may occur,
resulting in the decline of the evaporation heat transfer coefficient at increasing inlet vapor quality at high xin.
This is normally seen at high imposed heat flux and sat-urated temperature and low mass flux. Finally, an empirical correlation is proposed to correlate the present data for R-134a and R-407C evaporation heat transfer coefficients in the small tubes.
Acknowledgements
The financial support of this study by the engineering division of National Science Council of Taiwan, ROC through the contract NSC 92-2212-E-009-016 is greatly appreciated.
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