Investigation of the two-phase convective boiling of HFO-1234yf in a 3.9 mm diameter tube

Investigation of the two-phase convective boiling of HFO-1234yf

in a 3.9 mm diameter tube

Ming-Chang Lu, Jing-Rei Tong, Chi-Chuan Wang

⇑

Department of Mechanical Engineering, National Chiao Tung University, Hsinchu 300, Taiwan

a r t i c l e

i n f o

Article history:

Received 26 February 2013

Received in revised form 30 May 2013 Accepted 1 June 2013

Available online 13 July 2013 Keywords:

HFO-1234yf Convective boiling Heat transfer coefficient Mini-channel Microchannel

a b s t r a c t

In this study, the influences of heat flux and mass flux on the two-phase convective boiling heat transfer performance are reported for refrigerants HFO-1234yf and HFC-134a in a 3.9 mm smooth diameter tube. Tests are performed with a saturation temperature of 10 °C. It is found that at lower vapor quality region the nucleate boiling is the dominant heat transfer mechanism while the convective evaporation mecha-nism takes control at the higher vapor quality region. Both HFC-134a and HFO-1234yf shows similar trend and the difference in heat transfer coefficient between HFO-1234 and HFC-134a is quite small. The comparable heat transfer performance between HFC-134a and HFO-1234yf is attributed to similar physical properties and nucleate boiling contribution. The present test results are in line with some exist-ing reports but are inconsistent with one other study havexist-ing a tube diameter of 1.1 mm. It is found that the departure of heat transfer coefficients between the available publications is mainly attributed to the different flow phenomena caused by the difference of the channel size and channel geometry. A notice-able deterioration of the heat transfer coefficient for HFO-1234yf is encountered in the microchannel. The pressure drops for HFC-134a is about 5–15% higher than that of HFO-1234yf.

Ó 2013 Elsevier Ltd. All rights reserved.

1. Introduction

Concerns for chlorofluorocarbons (CFCs) and hydrochloroflu-orocarbans (HCFCs) refrigerants casting impact on environment lead to the advent of hydrofluorocarbans (HFCs). Despite HFC refrigerants have no ozone depletion potential (ODP), many of them have a relatively high global warming potential (GWP) which casts significant impact on the environment. For example, HFC-134a is the extensively used refrigerant in air-conditioning and automobile air conditioners (MACs), it has a GWP of 1300 (time horizon of 100 years). As a result, efforts were made to search for new refrigerants that are environmentally benign and can be used to in future air-conditioning and mobile air conditioning systems. Among the candidates, HFO-1234yf is regarded as one of the promising candidates for its GWP is as low as four. The thermo-physical properties, cycle performance, and two-phase heat trans-fer performance of HFO-1234yf are the key parameters to assess the feasibility of using this new refrigerant in air conditioners. The thermophysical properties of refrigerant mixture were similar to those of HFC-134a (Arakawa et al. [1]), thereby offering an opportunity as a drop-in solution for current mobile air condition-ers. Normally a drop-in solution yields a lower system

performance for lacking optimization in the design process. For in-stance, Lee and Jung[2]had shown that the coefficient of perfor-mance and capacity of HFO-1234yf are up to 2.7% and 4.0% lower than those of HFC-134a, respectively during a typical drop-in experiment. The compressor discharge temperature and the amount of refrigerant charge of HFO-1234yf are 6.5 °C and 10% lower than those of HFC-134a. Analogous results were also re-ported by Zilio et al. [3] and Navarro-Esbri et al. [4], they also showed a slight decrease in COP for HFO-1234yf system at a same cooling capacity with HFC-134a.

For further optimizing the system performance, further details in designing the heat exchangers (condenser or evaporator) are imperative. As a consequence, information about the two-phase heat transfer convective performance in the evaporator plays a crucial role in optimizing the heat exchangers. However, the pub-lished results regarding to the convective evaporation performance of HFO-1234yf is still limited and inconsistent. For instance, Saitoh et al.[5]conducted study for boiling heat transfer of the refrigerant HFO-1234yf flowing in a smooth small-diameter horizontal tube (inner diameter (ID): 2 mm) and Li et al.[6]used similar test facil-ity and identical test tube for comparing the HTC between HFC-32 and HFO-1234yf. Their test results showed that from the low to the high vapor quality region the difference between the heat transfer coefficients of HFO-1234yf and HFC-134a is small, Saitoh et al.[5] attributed this to the small differences in their thermodynamic properties. Recently, Col et al. [7] performed flow boiling of 0017-9310/$ - see front matter Ó 2013 Elsevier Ltd. All rights reserved.

http://dx.doi.org/10.1016/j.ijheatmasstransfer.2013.06.004 ⇑ Corresponding author.

E-mail address:ccwang@mail.nctu.edu.tw(C.-C. Wang).

Contents lists available atSciVerse ScienceDirect

International Journal of Heat and Mass Transfer

j o u r n a l h o m e p a g e : w w w . e l s e v i e r . c o m / l o c a t e / i j h m t

HFO-1234yf in a 1mm diameter circular microchannel and com-pared to R134a with a saturation temperature of 31 °C. They found that there were no significant differences between the flow boiling performance of R1234yf and R134a. On the other hand, Mortada et al.[8]performed an experiment for HFO-1234yf and HFC-134a in a 1.1 mm rectangular channel with rather small mass flux of 20–100 kg m2_s1_{and heat flux from 2 to 15 kW m}2_{. However,} their results showed that the HTC for HFO-1234yf is lower than that of HFC-134a as much as 40%. The results are contradictory to the findings of Saitoh et al.[4]and Col et al.[7]. A recent over-view about the general two-phase heat transfer characteristics for HFO-1234yf by Wang[9]also reported some inconsistent data in condensation.

In view of the relatively few data associated with the connective boiling performance of HFO-1234yf, it is the objective of this study to report some newly tested data concerning the two-phase vective heat transfer performance. Moreover, there are some con-tradictory results about the HFO-1234yf and HFC-134a. The present study also aims to elaborate some possible causes about the differences of the existing data.

2. Experimental setup

The schematic of the experimental apparatus is depicted in Fig. 1(a). The test rig is composed of three independent flow loops. Namely, a refrigerant loop, a heating water flow loop and a glycol flow loop. The refrigerant flow loop consists of a variable speed gear pump which delivers subcooled refrigerant to the preheater. The refrigerant pump can provide refrigerant mass fluxes ranging from 100 to 500 kg m2_s1_{. A very accurate mass flowmeter is} in-stalled between the refrigerant pump and the preheater. Note that the accuracy of the mass flowmeter is generally 0.3% of the test span. The subcooled refrigerant liquid was heated in the preheater to achieve a prescribed evaporator inlet quality before entering the test section. Then, the refrigerant went into the test section to

vaporize. Finally, the two-phase refrigerant was condensed in a shell-and-coil condenser. The horizontal test section is a double-pipe heat exchanger with effective heat transfer length of 0.6 m. Its detailed configuration can be seen fromFig. 1(b). Note that an electrically heated preheater is installed at the upstream of the test section, and the generated two-phase mixtures from the preheater flows into the test section. A 50-mm-thick rubber insulation is wrapped around the double-pipe test section to ensure heat loss to the ambient to be less than 10 W (less than 2% of the heat input) for the test tube. As seen fromFig. 1(b), inside the double-pipe heat exchanger, water flows countercurrently in the test section annu-lus, while refrigerant is evaporated inside the test tube. The pres-sure drop of the refrigerant across the test tube was meapres-sured by a differential pressure transducer with 10 Pa precision. A magnetic flowmeter was used to record the flowrates of water in the annulus of the test section. The magnetic flowmeter was calibrated in ad-vance with a calibrated accuracy of 0.002 L/s. An absolute pressure transducer was installed at the inlet and exit of the test section with resolution up to 0.1 kPa. During each experiment, the heat flux in the test section is maintained at a desired constant value. Experiments were conducted using a smooth copper tube having an internal diameter of 3.9 mm. Tests were conducted at an evap-oration temperature of 10 °C. All of the water and refrigerant tem-peratures, were measured by RTDs (Pt 100X) having a calibrated accuracy of 0.05 °C. The refrigerant leaving the test section was condensed and subcooled by a glycol circuit. The inlet temperature of the glycol is controlled by a 3 kW low-temperature thermostat. All of the data signals were collected and converted by a data acquisition system (Hybrid recorder). The data acquisition system then transmits the converted signals through USB interface to a host computer for further operation. Uncertainties of the heat transfer coefficients and reported in the present investigation, fol-lowing the single-sample analysis proposed by Moffat [10], are within ±7.1% of the measured values. The working fluids in this study are HFO-1234yf and HFC-134a.

Nomenclature

Ao outside heat transfer area of the tube, m2 Ai nominal inside heat transfer area of the tube, m2 cp specific heat of water, J kg1K1

Co confinement number, Co ¼ r gðqL qG Þ  0:5

d

D hydraulic diameter at annulus side, m Di inside diameter of the tube, m Do outside diameter of the test tube, m G mass flux, kg m2_s1

g gravitation constant, N/m

hi inside heat transfer coefficient, W m2K1

ho heat transfer coefficient on the annulus side, W m2K1 ifg latent heat of evaporating vapor, J kg1

k thermal conductivity, W m1_K-1 L effective heating length, m LMTD log mean temperature difference, K

_

mwater mass flow rate of the refrigerant, kg s1

_

mwater mass flow rate of coolant water, kg s1

Nu Nusselt number, dimensionless p pressure, kPa

Pr reduced pressure

Pr Prandtl mumber, dimensionless q heat flux, W m2

_

Q heat transfer rate, W

Re Reynolds number, dimensionless Rw wall resistance, K W1

Twater,in inlet temperature of water at annulus side, K Twater,out outlet temperature of water at annulus side, K Tsat saturation temperature of the refrigerant, K DT temperature rise on the water coolant, K DT1 temperature difference,DT1= Tsat,in Twater,out, K DT2 temperature difference,DT2= Tsat,out Twater,in, K Uo overall heat transfer coefficient, W m2K1 v velocity of water at annulus side, m s1 x vapor quality

Greek symbols

q

l density of refrigerant, kg m3

l

dynamic viscosity of refrigerant, N s m2

r

surface tension of refrigerant, N m1 Subscript

L liquid phase

G gas phase

f water fluid at annulus side

i inside in inlet o outside out outlet w wall water water

3. Data reduction

The heat duty for the test section was obtained from the flow rate and temperature drop of the water on the annulus according to the relation

_

Q ¼ _mwatercp

D

T ð1Þ

where cpis the specific heat of water andDT is the temperature dif-ference between water inlet and outlet. _mwateris the mass flowrate of water. The overall heat transfer coefficient was then computed from Uo¼ _ Q LMTD  Ao ð2Þ where LMTD ¼

D

T1

D

T2 ln DT1 DT2   ð3Þ

D

T1¼ Tsat;in Twater;out ð4Þ

D

T2¼ Tsat;out Twater;in ð5Þ

where Tsat,inand Tsat,outare the saturation temperatures of the refrig-erant in the test section at the inlet and outlet, respectively while Twater,inand Twater,out denote the inlet and outlet temperature of the water coolant on the annulus. The in-tube heat transfer coeffi-cient was obtained from the thermal resistance equation:

1 UoAo¼ 1 hoAoþ R wþ 1 hiAi ð6Þ

where hoand hirepresent the average outside and inside heat trans-fer coefficients, and Rw denotes the wall resistance. In the present calculation, the overall resistance is based on the outer surface area, which is evaluated as DoL, where Dois the outside diameter of the test tube and L is the effective heat transfer length. The properties on the water side were calculated using average temperature of in-let and outin-let bulk fluid temperatures. Note that the heat transfer coefficient and heat flux is based on the inside surface area (DiL). The determination of the inside heat transfer coefficient, hi, requires knowledge of the outside heat transfer coefficient, ho. This was accomplished by means of a separate water-to-water tests on the Transducer Magnetic Magnetic Meter Flow Thermostat T Water Subcooler Mass Flow Meter T Thermostat Water Sight glass Flow Meter Preheater Gauge Pressure T P Inlet P Pressure Referigerant pump Differential Transducer Tin Test section P Thermostat T Low temperature

(a) Test setup

Heat transfer tube Differential Pressure Transducer Inside diameter of outer tube Outside diameter test tube Cooling water Inlet pressure Transducer m ( m / h ) outer tube G ( kg / m s )2 2 refrigerant 2 refrigerant 2 test tube 3 t P P T T t C1 r1 r2 2 1 L = 60 cmc

(b) Test section

Fig. 1. Schematic of the test facility.

same apparatus, with subsequent Wilson-plot analyses yielding the individual heat transfer coefficient relationships. The subsequent outline is a typical Wilson plot method applicable to obtain the heat transfer coefficient in the water side.

1. Conduct heat transfer experiments for the test tube (double-tube heat exchanger) with water running in both (double-tube and annulus side. Hence one can calculate the associate heat trans-fer rate and the logarithmic mean temperature diftrans-ference, respectively. Notice that it is suggested that the temperature difference in both sides side should be at least 2 °C as suggested for common heat exchanger test to minimize the measurement uncertainty (Wang et al.[11], Wang[12]).

2. The overall resistance can be easily estimated as: Rt¼ 1 UA¼ 1 hiAi þ Rwþ 1 hoAo ð7Þ 3. To obtain the annulus side resistance (heat transfer coefficient) via Wilson plot, it is a must to maintain a fixed thermal resis-tance in the tube side throughout the test. This can be made possible by fixing an average tube side temperature and a fixed flow rate.

4. The heat transfer performance in annulus side is generally assumed to be the following form:

Nu ¼ C0ReaPrb ð8Þ

Note that an iteration process must be carried out to obtain the cor-rect exponent values of a and b. Initially, one can presume some val-ues of a and b. As a consequence, the heat transfer coefficient in the annulus side is obtained:

Rt¼ 1 UA¼ 1 hiAi þ Rwþ 1 hoAo ¼ 1 hoAo þ C1 ¼ 1 C0ReaPrb k_DfAo þ C1 ð9Þ

5. Rearranging the overall resistance equation in the following, Rt¼ 1 UA¼ 1 C2

v

af þ C1 ð10Þ

The above equation takes the form as

y ¼ mx þ c ð11Þ

With y = 1/UA, m = 1/C2, x ¼

v

af , and c = C1.

6. By changing the velocity in the annulus side, one can have a plot of y vs. x on a linear scale. However, if the plot is not linear, one should adjust the corresponding values of a and b to achieve the linearity. The slope m and the intercept c are then determined from the plot. With m, the heat transfer coefficient in the annu-lus side can be obtained.

The vapor quality entering the test section (xin) is calculated from the energy balance of the preheater and the quality change in each test section is given by the energy balance

D

x ¼ _ Q _ mrifg ð12Þ

where ifgis the latent heat of refrigerant and _mr is the refrigerant mass flowrate. The average quality in each test section is given by:

xave¼ xinþ

D

x

2 ð13Þ

4. Results and discussion

Fig. 2shows the measured heat transfer coefficients between HFC-134a and HFO-1234yf when Ts= 10 °C and G = 200 kg m2s1 with supplied heat flux being, 5.67, 11.35, and 18.9 kW m2_, respectively. As expected, the heat transfer coefficients (HTC) rise with the vapor quality, indicating appreciably influence of nucleate boiling. On the other hand, the dependence of heat flux for HTCs against vapor quality start to decline when the vapor quality is fur-ther increased, suggesting the convective evaporation is in control. The results are applicable for both HFC-134a and HFO-1234yf. In addition, one can see that the measured HTCs for HFC-134a and HFO-1234yf are virtually the same. The results are in line with the measurements of Saitoh et al. [5] and Li et al. [6]. Saitoh et al.[5]conducted convective boiling heat transfer of the refriger-ant HFO-1234yf flowing in a smooth small-diameter horizontal tube (inner diameter (ID): 2 mm) and Li et al.[6]used similar test facility and identical test tube for comparing the HTC between HFC-32 and HFO-1234yf. The test tube was heated by direct elec-trification using a DC power supply connected to two electrodes soldered at the flanges of the two ends of the test tube. Their experimental conditions are Ts= 15 °C, q = 6–24 kW m2, and G = 100–400 kg m2_s1_{. Their measurements also showed that} the HTCs are increased with the supplied heat flux at low vapor quality; thus, nucleate boiling is the dominant heat transfer mech-anism at the low vapor quality regime. On the other hand, the detectable rise of HTC vs. vapor quality for a low heat flux around 5.7 kW m2_{is associated with the change of flow pattern. This is} because the annular flow may prevail at high quality region. How-ever, as claimed by the authors[5]who argued that the nucleate boiling is the dominant heat transfer mechanism when q is about 11.5 or 19 kW m2_{, thereby showing a moderate change of HTC} as the vapor quality is increased. This seems feasible but the rela-tive effect of heat flux, based on the test results of Saitoh et al.[5], is in fact lower. A rough estimation of the HTC with the heat flux dependency is about h  q0.42which is generally appreciably lower than the pure nucleate boiling where h  q06–0.7_{. In this sense, it is} expected that convective evaporation still plays certain minor role rather than pure nucleate boiling. Basically our measurements for

HFO-1234yf and HFC-134a in φ3.9mm smooth tube

Vapor Quality

0.0 0.2 0.4 0.6 0.8 1.0

Heat Transfer Coefficient (W m

-2K -1) 0 2000 4000 6000 8000 10000 HFO-1234yf, G = 200 kg m-2 s-1 , q = 5.73 kW m-2 HFO-1234yf, G = 200 kg m-2 s-1 , q = 11.46 kW m-2 HFO-1234yf, G = 200 kg m-2 s, q = 19.2 kW m-2 HFC-134a, G = 200 kg m-2 s-1 , q = 5.67 kW m-2 HFC-134a, G = 200 kg m-2 s-1 , q = 11.35 kW m-2 HFC-134a, G = 200 kg m-2 s-1 , q = 18.91 kW m-2 T=10oC

Fig. 2. Test results for heat transfer coefficient for HFC-134a and HFO-1234yf at a saturation temperature of 10 °C and G = 200 kg m2

s1 .

both HFC-134a and HFO-1234yf also reveal a similar heat flux dependence at the lower vapor quality region.

It is interesting to note that the HTCs for HFC-134a and HFO-1234yf are virtually the same regardless the change of heat flux as depicted inFig. 2. To explain this phenomenon, one must resort to the basic heat transfer mechanisms associated with convective boiling, namely the nucleate boiling and convective evaporation. For nucleate boiling, it is often recognized that three mechanisms, namely bubble agitation (including bubble frequency and site den-sity), vapor–liquid change phenomenon, and evaporation are asso-ciated with basic mechanisms of the nucleate boiling heat transfer (Thome,[13]). As shown inTable 1a(Akasaka et al.[1]and Tanaka et al.[14]), the HFO-1234yf has a higher reduced pressure at the same saturation temperature. This is because its critical pressure for HFO-1234yf is about 17% lower than that of HFC-134a. In fact, at a saturation temperature of 40 °C the reduced pressure Pris approximately 20% higher than that of HFC-134a, thereby leading to a larger activation sites for HFO-1234yf that would boost the heat transfer coefficient.

On the other hand, the smaller bubble departure diameter

 r

gðqLqGÞ

 0:5

;

r

:surface tension;

q

:density

 

of HFO-1234yf implies a lower bubble agitation, and a smaller vapor–liquid change contribution which offset the positive contribution from the higher reduced pressure. A summation of the foregoing effects, conse-quently, HFC-134a and HFO-1234yf shows nearly identical nucleate boiling HTCs. Therefore, the HTCs for HFC-134a and HFO-1234yf are about the same. To further elaborate about the foregoing qualitative argument, calculation of the HTCs is made using the well-known Chen’s correlation[15]applicable to conven-tional channels for HFC-134a and HFO-1234yf with Tsat= 10 °C in a smooth tube having an inner diameter of 10 mm, G = 200 and 400 kg m2_s1 _{and q = 20 kW m}2 _{as shown inFig. 3. As clearly} shown in the figure, the calculated HTCs for HFC-134a and HFO-1234yf is nearly the same when the vapor quality is less than 0.6, and HFC-134a is marginally higher than those of HFO-1234yf when the vapor quality exceeds 0.6. The results suggest that the differ-ence in HTC between HFC-134a and HFO-1234yf is rather small.

However, as mentioned in the introduction, test results from Mortada et al.[6]for HFO-1234yf and HFC-134a in a 1.1 mm rect-angular channel at some rather small mass flux of 20–100 kg m2 -s1_{and heat flux from 2 to 15 kW m}2_{showed an opposite trend.} Their data indicated the dominance of convective boiling rather than nucleate boiling. There are two possible explanations for this difference. The first is associated with the difference in tube geom-etry. Notice that present test tube is a round tube while it is rect-angular configuration for Mortada et al.[6], and the influence of the liquid film on the heat transfer in a rectangular tube is different from that in a circular tube. Another possible explanation of their results is attributed to the difference in flow phenomenon. As pro-posed by Kew and Cornwell[16]who took a rigorous approach to define the micro-channels in two-phase flow. They had come out a so-called confinement number (Co), defined as the ratio of the departing bubble diameter to the channel diameter, to determine the transition to micro-channels flow, as below:

Co ¼ r

gðqLqGÞ

 0:5

d ð11Þ

where

r

is the surface tension,

q

Lis the liquid density,

q

Gis the va-por density and d is the inner tube diameter. Based on analysis from some previous data, Kew and Cornwell[16]found that the mea-sured heat transfer coefficients departed appreciably from the con-ventional channel correlations. They thus chose Co = 0.5 as the transition criteria for microchannel. In this study, the associated Co is around 0.2–0.22 whereas it is about 0.74–0.81 for Mortada et al.[8]. Hence, it explains in part about the different heat transfer behavior between Mortada et al.[8]and the present results. Note that the test results from[8]showed an early dry out at a vapor quality near 0.4 irrespective of the supplied heat flux. The results implied that the annular flow pattern may prevail at an even lower vapor quality, thereby revealing a dominance of the convective boil-ing. In the meantime, Mortada et al.[8]reported that the measured HTC for HFO-1234yf is lower than that of HFC-134a as much as 40%. The results are contradictory to the findings of Saitoh et al.[5]and the present results where the HTCs for HFC-134a and HFO-1234yf are virtually the same. Again, the difference may be due to the dif-ference in heat transfer characteristics between conventional chan-nels and microchanchan-nels. Also, since nucleate boiling is in control in their case, suggesting the aforementioned positive contribution by nucleate boiling for HFO-1234yf has been lifted. In this regard, the flow pattern and physical properties play essential role in the heat transfer performance. The agitation contribution for HFO-1234yf, as depicted in aforementioned discussion, is lowered than that of HFC-134a. In addition, as shown inTable 1b, typically the vapor density of HFC-134a is about 20% lower than that of HFO-1234yf. Hence, the effective vapor velocity for HFC-134a is

Table 1a

Fundamental constants of HFO-1234yf. Molecular weight (g mol1 ) Critical temperature (K) Critical pressure (MPa) HFC-134a 102 374.13 4.07 R-1234yf 114.042 367.85 3.382 x 0.0 0.2 0.4 0.6 0.8 1.0 h (W m -2K -1) 0 5000 10000 15000 20000 25000 30000 G = 200 kg m-2 s-1 , HFC-134a G = 200 kg m-2 s-1 , HFO-1234yf G = 400 kg m-2 s-1 , HFC-134a G = 400 kg m-2 s-1 , HFO-1234yf

Fig. 3. Comparison of the calculated convective heat transfer coefficient using the Chen’s correlation between HFC-134a and HFO-1234yf at a saturation temperature of 10 °C and q = 20 kW m2_.

Table 1b

Comparison of the physical properties of HFC-134a and HFO-1234yf at Tsat= 6 °C and 15 °C.

Property Ts= 6 °C Ts= 15 °C

R134a HFO-1234yf R134a HFO-1234yf

p (kPa) 363 384.8 493.15 513 qL(kg m3) 1274.6 1156.8 1243 1127.5 qG(kg m3) 17.758 21.52 24.005 28.7 qL(lPa s) 251.28 203.6 224.75 182.5 qG(lPa s) 10.982 11.716 11.365 12.13 kL(W m1K1) 0.08936 0.07266 0.08545 0.06925 k(kJ kg1_{) 193.98 158.32 186.46 151.95} r(N m1 ) 0.010708 0.008564 0.00945 0.0074 cp,L(kJ kg1K1) 1.358 1.2786 1.3875 1.3125 cp,G(kJ kg1K1) 0.926 0.9622 0.9735 1.012

20% higher than that of HFO-1234yf, implicating a larger convective contribution. In summary of these two effects, it can explain why the heat transfer performance of HFC-134a is superior to that of HFO-1234yf in microchannels.Fig. 4shows the effect of mass flux on the measured heat transfer coefficients between HFC-134a and HFO-1234yf when Ts= 10 °C and q  11.4 kW m2with mass flux being 200, 300, and 400 kW m2_{, respectively. At a lower vapor} quality of 0.2, the effect of mass flux on HTCs is negligible. The re-sults accord withFig. 2where nucleate boiling is the dominant heat transfer mechanism at a lower vapor quality region. Yet the influ-ence of mass flux become more and more pronounced as the vapor quality is increased. Apparently the effect of convective evaporation

starts to gain control. On the other hand, the HTCs for HFC-134a and HFO-1234yf remain about the same subject to change of mass flux. The test results also indicate an identical heat transfer performance between HFC-134a and HFO-1234yf. The adiabatic two-phase pressure drop of Ts= 10 °C for both HFO-1234yf and HFC-134a is de-picted inFig. 5. As expected, the pressure drop is increased with the mass flux and vapor quality. Yet the pressure drops for HFC-134a exceeds those of HFO-1234yf by approximately 5–15%. The higher pressure drops of HFC-134a may be mainly associated with the den-sity difference. It appears that the vapor denden-sity for HFC-134a is about 20% lower than that of HFO-1234yf as shown inTable 1b, implying a higher vapor velocity of HFC-134a and a higher pressure drop accordingly.

5. Conclusions

In this study, convective boiling heat transfer performance for refrigerants HFO-1234yf and HFC-134a in a 3.9 mm smooth diam-eter tube is reported. Tests are performed with a saturation tem-perature of 10 °C. The influences of heat flux and mass flux on the HTCs are reported in this study. It is found that at lower vapor quality region the nucleate boiling is the controlled mechanism while the convective evaporation mechanism takes control at the higher vapor quality region. Both HFC-134a and HFO-1234yf shows similar trend. Test results also show that the difference in heat transfer coefficient between HFO-1234 and HFC-134a is quite small. The comparable heat transfer performance between HFC-134a and HFO-1234yf is attributed to similar physical properties and the nucleate boiling contribution. This is because the nucleate boiling and convective evaporation contributions of these two refrigerants are about the same provided that the channel size is not too small. Despite the present test results are in line with some existing publications like[4–6], the test results are opposite to those reported by[7]. The departure between the present mea-surements and[7]is mainly due to the difference in channel geom-etry and the flow phenomenon. The tube geomgeom-etry for[4–6]and the present one is of round tube configuration whereas[7]is of rectangular configuration and it is operated in the microchannel regime. The difference in flow phenomenon subject to channel size and channel configuration results in an opposite heat transfer per-formance. On the other hand, for the present test tube, the effect of mass flux on the HTC becomes more pronounced at higher quality region. This is applicable for both HFC-134a and HFO-1234yf. The pressure drops for HFC-134a is about 5–15% higher than that of HFO-1234yf due to its higher superficial vapor velocity.

Acknowledgements

The author would like to express gratitude for supporting funding from the National Science Council of Taiwan (100-2221-E-009-087-MY3 and 102-ET-E-009-006-ET). Refrigerant HFO-1234yf provided by Dr. Lawrence Chin from Honeywell is highly appreciated.

References

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bench tester for automobile applications, Appl. Therm. Eng. 35 (2012) 240– 242.

[3]C. Zilio, J.S. Brown, G. Schiochet, A. Cavallini, The refrigerant R1234yf in air conditioning systems, Energy 36 (2011) 6110–6120.

[4]J. Navarro-Esbri, J.M. Mendoza-Miranda, A. Mota-Babiloni, A. Barragan-Cervera, J.M. Belman-Flores, Experimental analysis of R1234yf as a drop-in replacement for R134a in a vapor compression system, Int. J. Refrig. 36 (2013) 870–880.

HFO-1234yf and HFC-134a in _{φ3.9mm smooth tube}

Vapor Quality

0.0 0.2 0.4 0.6 0.8 1.0

Heat Transfer Coefficient (W m

-2 K -1 ) 2000 4000 6000 8000 10000 HFO-1234yf, G = 200 kg m-2 s-1 , q = 11.46 kW m-2 HFO-1234yf, G = 300 kg m-2 s-1 , q = 11.46 kW m-2 HFO-1234yf, G = 400 kg -2 s-1 , q = 11.46 kW m-2 HFC-134a, G = 200kg m-2 s-1 , q = 11.35 kW m-2 HFC-134a, G = 300 kg m-2 s-1 , q = 11.35 kW m-2 HFC-134a, G = 400kg m-2 s-1 , q = 11.35 kW m-2 T=10oC

Fig. 4. Effect of mass flux on HTCs for HFC-134a and HFO-1234yf at a saturation temperature of 10 °C and q  11.4 kW m2_.

HFO-1234yf and R-134a in _{φ3.9mm smooth tube}

Vapor Quality 0.0 0.2 0.4 0.6 0.8 1.0 dP/dZ (kPa/m) 0 10 20 30 40 HFO-1234yf, G = 200 kg m-2 s-1 HFO-1234yf, G = 300 kg m-2 s-1 HFO-1234yf, G = 400 kg m-2 s-1 HFO-1234yf, G = 500 kg m-2s-1 R-134a G = 200 kg m-2 s-1 R-134a G = 300 kg m-2 s-1 R-134a G = 400 kg m-2 s-1 R-134a G = 500 kg m-2 s-1 T=10oC

Fig. 5. Effect of mass flux on the pressure drops for HFC-134a and HFO-1234yf at a saturation temperature of 10 °C.

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