International Journal of Heat and Mass Transfer 54 (2011) 2080–2088

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International Journal of Heat and Mass Transfer

journal homepage: www.elsevier .com/locate / i jhmt

Experimental investigation on two-phase flow boiling heat transfer of fiverefrigerants in horizontal small tubes of 0.5, 1.5 and 3.0 mm inner diameters

Jong-Taek Oh a, A.S. Pamitran b, Kwang-Il Choi a, Pega Hrnjak c,⇑a Department of Refrigeration and Air Conditioning Engineering, Chonnam National University, San 96-1, Dunduk-Dong, Yeosu, Chonnam 550-749, Republic of Koreab Department of Mechanical Engineering, University of Indonesia, Kampus Baru UI, Depok 16424, Indonesiac Department of Mechanical Science and Engineering, ACRC, University of Illinois at Urbana-Champaign, 1206 West Green Street, Urbana, IL 61801, USA

a r t i c l e i n f o a b s t r a c t

Article history:Available online 1 February 2011

Keywords:RefrigerantFlow boilingFlow patternHeat transfer coefficientCorrelationHorizontal small tubes

0017-9310/$ - see front matter � 2011 Elsevier Ltd. Adoi:10.1016/j.ijheatmasstransfer.2010.12.021

⇑ Corresponding author. Tel.: +1 217 244 6377; faxE-mail address: pega@uiuc.edu (P. Hrnjak).

An experimental investigation on two-phase flow boiling heat transfer with refrigerants of R-22, R-134a,R-410A, C3H8 and CO2 in horizontal circular small tubes is presented. The experimental data wereobtained over a heat flux range of 5–40 kW m�2, mass flux range of 50–600 kg m�2 s�1, saturation tem-perature range of 0–15 �C, and quality up to 1.0. The test section was made of stainless steel tubes withinner diameters of 0.5, 1.5 and 3.0 mm, and lengths of 330, 1000, 1500, 2000 and 3000 mm. The exper-imental data were mapped on Wang et al. (1997) [5] and Wojtan et al. (2005) [6] flow pattern maps. Theeffects of mass flux, heat flux, saturation temperature and inner tube diameter on the heat transfercoefficient are reported. The experimental heat transfer coefficients were compared with some existingcorrelations. A new boiling heat transfer coefficient correlation that is based on a superposition modelfor refrigerants in small tubes is presented with 15.28% mean deviation and �0.48% average deviation.

� 2011 Elsevier Ltd. All rights reserved.

1. Introduction

Although most HCFCs’ chemicals break down before reachingthe ozone layer, the chlorine produced reaches the stratosphereand depletes the ozone layer. R-22, a HCFC, is still widely used inthe refrigeration and air-conditioning industry even though somecountries have ceased using HCFCs. In an attempt to find a replace-ment for R-22 as an environmental conservation effort, HFC andnatural refrigerants, such as R-134a, R-410A, CO2 and C3H8 havebeen studied extensively because they do not contain chlorine,which can deplete the ozone layer. The zero or very low GlobalWarming Potential (GWP) is another advantage of these naturalrefrigerants. And recently, energy and material efficiencies haveemerged as important topics in refrigeration and air conditioning.Recent awareness of the advantages of process intensification hasled to the demand for smaller evaporators in refrigeration, air con-ditioning and processing due to their energy and material efficien-cies. However, existing methods for predicting the heat transfer oftwo-phase flow in large tubes cannot properly predict same type ofheat transfer in small tubes. Published data related to two-phaseflow and heat transfer in small tubes are limited compared withdata for large tubes. Furthermore, compared with evaporation inconventional tubes, that in a small tube may yield a higher heattransfer coefficient due to the larger contact area per unit volume

ll rights reserved.

: +1 217 333 1942.

of fluid. Several studies dealing with two-phase flow heat transferin small tubes have been published in the past years. In evapora-tion with small tubes, such as reported by Zhang et al. [1], Tranet al. [2], Pettersen [3] and Yun et al. [4], the contribution of nucle-ate boiling is predominant.

The heat transfer coefficients of boiling flows of R-22 and itsalternatives R-134a, R-410A, CO2 and C3H8 in horizontal smoothsmall tubes were measured in this study. The experimental resultswere compared with the predictions given by seven existing heattransfer correlations. Due to the limitations of the correlation forforced convective boiling of refrigerants in small tubes, a new cor-relation for evaporative refrigerants in small tubes was developedin this study based on superposition.

2. Experimental aspects

2.1. Experimental apparatus and method

The experimental facilities are schematically shown in Fig. 1(a)and (b). The test facility consisted of a condenser, subcooler, recei-ver, refrigerant pump, mass flow meter, preheater, and test sec-tions. For the test with the 3.0 and 1.5 mm inner diameter tubes,a variable A.C output motor controller was used to control the flowrate of the refrigerant. And for the test with the 0.5 mm innerdiameter tube in an open-loop system, the flow rate was controlledwith a needle valve. A Coriolis-type mass flow meter was installedin the horizontal layout for the test with the 1.5 and 3.0 mm inner

Nomenclature

AD average deviation, AD ¼ 1n

Pn1ðhpred�hexpÞ�100

hexp

Bo boiling number, Bo = q/GifgC co-current two-phase pressure drop multiplierD diameter (m)F convective two-phase multiplierf friction factorG mass flux (kg m�2 s�1)h heat transfer coefficient (kW m�2 K�1)i enthalpy (kJ kg�1)k thermal conductivity (kW m�1 K�1)L length of test section (m)MD mean deviation, MD ¼ 1

n

Pn1ðhpred�hexpÞ�100

hexp

��� ���Q electric power (kW)q heat flux (kW m�2)Re Reynolds number, Re = GD/lS suppression factor of nucleate boilingT temperature (K)W Mass flow rate (kg s�1)X Martinelli parameterx mass qualityz axial coordinate (m)

Greek lettersl dynamic viscosity (N s m�2)q density (kg m�3)

r surface tension (N/m)u2 two-phase frictional multiplier gradients and differ-

ences(dp/dz) pressure gradient (N m�2 m�1)

Subscriptsexp experimental valuef saturated liquidg saturated vaporin physical property based on inlet temperaturei inner tubenbc nucleate boiling contributionpred prediction valuesat saturationsc subcooledt turbulenttp two-phasev laminarw wall

J.-T. Oh et al. / International Journal of Heat and Mass Transfer 54 (2011) 2080–2088 2081

diameter tubes, whereas a weighing balance was used for the testwith the 0.5 mm inner diameter tube to measure the refrigerantflow rate. A preheater was installed to control the mass qualityby condensing the refrigerant before it entered the test section.For evaporation at the test section, a certain heat flux was con-ducted from a variable A.C voltage controller. The vapor refrigerantfrom the test section was then condensed in the condenser andsubcooler, and then the condensed vapor was supplied to thereceiver.

The test section was made of stainless steel circular smoothtubes with inner tube diameters of 3.0, 1.5 and 0.5 mm. The rateof input electric potential and current were adjusted to controlthe input power and to determine the applied heat flux, whichwas measured by a standard multimeter. The test sections wereuniformly and constantly heated by applying electric current di-rectly to the tube walls, which were well insulated with foamand rubber. The outside tube wall temperatures at the top, bothsides and bottom were measured at 100 mm (for the test sectionwith inner tube diameters of 3.0 and 1.5 mm) and at 30 mm (forthe test section with inner tube diameter of 0.5 mm) axial intervalsfrom where heating was started by using T-type copper–constan-tan thermocouples at each site. The local saturation pressure,which was used to determine the saturation temperature, wasmeasured using bourdon tube type pressure gauges at the inletand the outlet of the test sections. The differential pressure wasmeasured by the bourdon tube type pressure gauges and a differ-ential pressure transducer. Sight glasses with the same inner tubediameter as the test section were installed to visualize the flow.

The experimental conditions used in this study are listed inTable 1. The temperature and flow rate measurements were re-corded by logger software programs. The physical properties ofthe refrigerants were obtained from REFPROP 8.0. The experimen-tal uncertainty associated with all the parameters is tabulated inTable 2. The uncertainties were obtained using both random andsystematic errors, and these values changed according to the flowconditions, so their minimum to maximum ranges are shown.

2.2. Data reduction

The local heat transfer coefficient was defined as follow:

h ¼ qTwi � Tsat

ð1Þ

The inside tube wall temperature, Twi, was the average temper-ature of the top, both right and left sides, and the bottom walltemperatures, and was determined based on the steady-stateone-dimensional radial conduction heat transfer through the wallwith internal heat generation. Power generation is then obtainedfrom the rate of input electric power divided by volume of the cur-rent-carrying medium

_q ¼ QV¼ 4Q

pLðD2o � D2

i Þð2Þ

The heat flux that was used in Eq. (1) is then derived from Eq. (3)

q ¼_qðD2

o � D2i Þ

4Dið3Þ

The vapor quality at the measurement locations were deter-mined based on the thermodynamic properties. The outlet massquality was then determined using the following equation:

xo ¼Diþ ifi � if

ifgð4Þ

The refrigerant flow at the inlet of the test section was not com-pletely saturated. Even though it was almost completely saturated,it was necessary to determine the subcooled length to ensurereduction data accuracy. The subcooled length was calculatedusing Eq. (5)

zsc ¼ Lif � ifi

Di¼ L

if � ifi

ðQ=WÞ ð5Þ

D : 3.0 and 1.5 mm

A

A’

Thermocouple

Subcooler

Receiver

Condenser

Ref. Pump

Preheater

Water Pump

Refrigerator Unit

Sightglass

A’

L : 1.0, 1.5, 2.0, 3.0 m

A

Test Sections

Mass Flowmeter

100 mm

Cooler

Receiver

Condenser

Refrigerantcylinder

Mass Flow meterTest SectionL = 330 mm

Needle valve

Di = 500 µm

A

A’

Thermocouplesat every 30 mm

A’

A

Refrigerator

(a)

(b)

Fig. 1. Experimental test facility: (a) For test section with tubes of Di = 3.0 mm and Di = 1.5 mm; (b) For test section with tube of Di = 0.5 mm.

2082 J.-T. Oh et al. / International Journal of Heat and Mass Transfer 54 (2011) 2080–2088

3. Result and discussion

3.1. Flow pattern

The flow with heat addition, or diabatic flow, is a coupled ther-modynamic problem. Heat transfer causes phase changes andhence, changes of phase distribution and flow pattern. The presentexperimental results were mapped on Wang et al. [5] and Wojtanet al. [6] flow pattern maps, which were developed for diabatictwo-phase flow. The Wang et al. [5] flow pattern map is a modifiedBaker [7] map, developed using R-22, R-134a, and R-407C inside a6.5 mm-diameter horizontal smooth tube. The Wang et al. [5]study showed that the flow transition of mixture refrigerant is con-siderably delayed compared with those of pure refrigerants. For R-410A, at the initial stage of evaporation, R-32 evaporated fasterthan R-125. Therefore, R-32 increased the concentration of the va-por phase, and R-125 increased the concentration of the liquidphase throughout the evaporation process at the liquid–vaporinterface. This resulted in a higher mean vapor velocity and a lowermean liquid velocity during evaporation. The physical propertiessuch as density, viscosity and surface tension of the other workingrefrigerants have strong effect on the flow pattern. Shown on theexisting flow pattern maps of Wang et al. [5] and Wojtan et al.[6], the predicted flow pattern resulting from selected data of thecurrent experiment can be seen in Figs. 2 and 3, respectively. TheWang et al. [5] map showed better prediction of the flow pattern

of the current experimental data for the beginning of annular flowthan the Wojtan et al. [6] map. Stratified wavy flow appeared ear-lier for higher mass flux and its regime was longer for the low massflux condition. Annular flow appeared earlier for higher mass fluxand its regime was longer for the high mass flux condition.The Wojtan et al. [6] flow pattern map is a modified version ofthe Kattan et al. [8] map, which was developed by using R-22and R-410A inside a 13.6 mm horizontal smooth tube. Kattanet al. [8] used five refrigerants inside 12 and 10.92 mm tubes.Because the Wojtan et al. [6] map was developed using aconventional tube, the flow pattern transition of the present exper-imental data showed a delay on this map. However, the Wojtanet al. [6] map can be used to predict the dry-out condition. TheWojtan et al. [6] map better predicted the dry-out condition forthe R-410A experimental data. Overall, the Wang et al. [5] flowpattern map provided a better flow pattern prediction for the cur-rent experimental results than the Wojtan et al. [6] map.

3.2. Effect on heat transfer coefficient

Fig. 4 shows the effect of mass flux on the heat transfer coeffi-cient. Mass flux had an insignificant effect on the heat transfercoefficient in the low quality region. This indicates that nucleateboiling heat transfer was predominant. The high nucleate boilingheat transfer occurred because of the physical properties of therefrigerants, namely their surface tension and pressure, and the

Table 1Experimental conditions.

Test section Horizontal circular smooth small tubesQuality up to 1.0

Working refrigerant R-22Inner diameter (mm) 3.0 1.5Tube length (mm) 2000 2000Mass flux (kg/(m2 s)) 400–600 300–600Heat flux (kW m�2) 20–40 10–20Inlet Tsat (�C) 10 10

Working refrigerant R-134aInner diameter (mm) 3.0 1.5 0.5Tube length (mm) 2000 2000 330Mass flux (kg/(m2 s)) 200–600 200–400 100Heat flux (kW m�2) 10–40 10 5–20Inlet Tsat (�C) 10 10 6–10

Working refrigerant R-410AInner diameter (mm) 3.0 1.5 0.5Tube length (mm) 3000 1500 330Mass flux (kg/(m2 s)) 300–600 300–600 70–400Heat flux (kW m�2) 10–40 10–30 5–20Inlet Tsat (C) 10–15 10–15 1–11

Working refrigerant C3H8

Inner diameter (mm) 3.0 1.5Tube length (mm) 2000 2000Mass flux (kg/(m2 s)) 50–240 100–400Heat flux (kW m�2) 5–25 5–20Inlet Tsat (C) 0–11 0–12

Working refrigerant CO2

Inner diameter (mm) 3.0 1.5Tube length (mm) 2000 2000Mass flux (kg/(m2 s)) 200–600 300–600Heat flux (kW m�2) 20–30 10–30Inlet Tsat (�C) 1–10 0–11

J.-T. Oh et al. / International Journal of Heat and Mass Transfer 54 (2011) 2080–2088 2083

geometric effect of the small tubes. A higher mass flux corre-sponded to a higher heat transfer coefficient at the moderate-highvapor quality region, due to an increase of the convective boilingheat transfer contribution; this result is similar to the results re-ported by Kuo–Wang [9], who used R-22 in a 9.52 mm tube. Inthe high quality region, the heat transfer coefficient dropped atlower qualities for a relatively higher mass flux; this trend agreeswith the results of Pettersen [3] and Yun et al. [4]. For the highermass flux condition in the convective evaporation region, the in-crease in the heat transfer coefficient appeared at a lower quality,which can be explained by the annular flow becoming dominant.Nucleate boiling suppression appeared earlier for the higher massflux, and this means that convective heat transfer appeared earlierunder the higher mass flux condition. The lower mass flux condi-tion results show smaller increases in the heat transfer coefficientin the convective region. The heat transfer coefficients suddenly in-crease in the annular flow region before the initial dry-out, and thiscan be explained by the fact that as quality was increased in theannular flow, the effective wall superheat decreased due to thethinner liquid film or less thermal resistance. The steep decreaseof the heat transfer coefficient at high qualities was due to the

Table 2Summary of the estimated uncertainty.

Parameters Uncertainty (%)

R-22 R-134a

Twi ±0.44 to ±3.90 ±0.29 to ±3.92P ±2.5 kPa ±2.5 kPaG ±1.84 to ±3.16 ±1.85 to ±3.80q ±1.78 to ±2.26 ±1.79 to ±2.59x ±1.88 to ±3.39 ±2.23 to ±4.19h ±2.74 to ±9.07 ±2.75 to ±9.24

effect of the small diameter on the boiling flow pattern; dry-patchoccurs easier in smaller diameter tubes and at higher mass fluxes.Several previous studies such as those of Kew and Cornwell [10]and Wambsganss et al. [11] used small tubes and showed thatnucleate boiling in small tubes tends to be predominant.

Fig. 5 depicts the dependence of heat flux on heat transfer coef-ficients in the low-moderate quality region. Nucleate boiling isknown to be dominant in the initial stage of evaporation, particu-larly under high heat flux conditions. The large effect of heat fluxon the heat transfer coefficient shows the dominance of nucleateboiling heat transfer contribution. At the higher quality region,nucleate boiling was suppressed, or convective heat transfer con-tribution was predominant; this is indicated by the low effect ofheat flux on the heat transfer coefficient. As the heat flux increasedat high qualities, the evaporation was more active and the dry-outquality reduced.

The effect of saturation temperature on heat transfer coefficientis depicted in Fig. 6. The heat transfer coefficient increases with theincrease in saturation temperature because of the large effect ofnucleate boiling. A higher saturation temperature leads to lowersurface tension and higher pressure, as shown in Table 3. The vaporformation in the boiling process indicates that a lower surface ten-sion and higher pressures provide a higher heat transfer coefficient[12].

Fig. 7 shows that a smaller inner tube diameter has a higher heattransfer coefficient at low quality regions. This is due to the more ac-tive nucleate boiling in a smaller diameter tube. As the tube diameterbecomes smaller, the contact surface area for heat transfer increases.More active nucleate boiling causes dry-patches to appear earlier.The quality for a rapid decrease in the heat transfer coefficient canbe lowered for a smaller tube. It is supposed that annular flow wouldappear at a lower quality in a smaller tube, and therefore, the dry-outquality would be relatively lower for a smaller tube.

3.3. Comparison of heat transfer coefficient

Fig. 8 shows the comparisons of the heat transfer coefficients of R-22, R-134a, R-410A, C3H8 and CO2 at some experimental conditions.The mean heat transfer coefficient ratio of R-22, R-134a, R-410A,C3H8 and CO2 was approximately 1.0:0.8:1.8:0.7:2.0. The heat trans-fer coefficient of CO2 was higher than that of the other workingrefrigerants during evaporation under all test conditions. The higherheat transfer coefficient of CO2 is believed to be due to its high boilingnucleation. CO2 has much lower surface tension and applies muchhigher pressure than the other working refrigerants. The heat trans-fer coefficients of R-22, R-134a, and C3H8 are similar due to their sim-ilar physical properties. The comparison of the physical properties ofthe present working refrigerants are given in Table 3. CO2 has a muchlower viscosity ratio lf/lg than the other working refrigerants,which means that the liquid film of CO2 can break easier than thoseof the other refrigerants. CO2 has also a much lower density ratio qf/qg than the other working refrigerants, which leads to a lower vaporvelocity, which in turn causes less suppression of nucleate boiling.

R-410A C3H8 CO2

±0.23 to ±6.53 ±0.18 to ±5.58 ±2.10 to ±4.56±2.5 kPa ±2.5 kPa ±2.5 kPa±1.84 to ±9.48 ±3.24 to ±9.78 ±1.85 to ±9.48±1.67 to ±3.20 ±2.07 to ±3.58 ±1.67 to ±2.70±1.78 to ±9.85 ±4.27 to ±9.82 ±1.79 to ±9.71±2.59 to ±10.33 ±1.78 to ±9.89 ±4.46 to ±8.23

(a)

(b)

Fig. 2. Present experimental results mapped on the Wang et al. [5] flow patternmap: (a) R-22 at q = 10 kW m�2, Di = 1.5 mm and Tsat = 10 �C; (b) R-410A atq = 20 kW m�2, Di = 3.0 mm and Tsat = 10 �C.

0

200

400

600

800

1000

0 0.2 0.4 0.6 0.8 1Str.

Intermittent

G(k

g/m

²s)

R-22q = 10 kW/m²Di = 1.5 mmTsat = 10 °C

Slug + Str.WavySlug

Annular

Str.Wavy

x

Dry-out

The current experimental data:Intermittent to AnnularAnnular to Dry-out

(a)

0

200

400

600

800

1000

0 0.2 0.4 0.6 0.8 1Str.

Intermittent

G(k

g/m

²s)

R-410Aq = 20 kW/m²Di = 3.0 mmTsat = 10 °C

Slug + Str.WavySlug

Annular

Str.Wavy

x

Dry-outMist

The current experimental data:Intermittent to AnnularAnnular to Dry-out

(b)

Fig. 3. Present experimental results mapped on the Wojtan et al. [6] flow patternmap: (a) R-22 at q = 10 kW m�2, Di = 1.5 mm and Tsat = 10 �C; (b) R-410A atq = 20 kW m�2, Di = 3.0 mm and Tsat = 10 �C.

2084 J.-T. Oh et al. / International Journal of Heat and Mass Transfer 54 (2011) 2080–2088

The heat transfer coefficients of the present study are comparedwith the results given by seven correlations in Table 4. The Gun-gor–Winterton [13], Jung et al. [14], Shah [15] and Tran et al. [2]correlations provided better predictions, with mean deviations oflower than 30%, than the other correlations. The Gungor–Winter-ton [13] correlation was a modification of the superposition model;it was developed using fluids in several small and conventionaltubes under various test conditions. The Jung et al. [14] correlationwas developed with pure and mixture refrigerants in conventionalchannels; its F factor contributed a big calculation deviation withthe current experimental data. The Shah [15] correlation wasdeveloped using a large data set for conventional tubes. The predic-tion with Shah [15] correlation was fair under prediction at lowquality region. The Tran et al. [2] correlation was developed forR-12 and R-113 in small tubes. The correlations of Chen [16] andWattelat et al. [17], which were developed for large tubes, have ahigh prediction deviation. The correlations of Kandlikar [18] andZhang et al. [1] were developed for small tubes; however, the cor-relations could not predict well the present experimental data. Thecorrelations of Wattelat et al. [17], Kandlikar [18] and Zhang et al.[1] showed a large deviation in the prediction of the CO2 data. TheKandlikar [18] correlation failed to predict the heat transfer coeffi-cient at the high quality region.

4. Development of a new correlation

Heat transfer of flow boiling is mainly governed by two impor-tant mechanisms, namely nucleate boiling and forced convectiveheat transfer. In two-phase flow boiling heat transfer, the nucleateboiling heat transfer contribution is suppressed by the two-phase

flow. Therefore, the nucleate boiling heat transfer contributionmay be correlated with the nucleate boiling suppression factor, S.Another contribution of convective heat transfer may be correlatedwith a liquid single phase heat transfer. The F factor is introducedas a convective two-phase multiplier to account for enhanced con-vection due to co-current two-phase flow. A superposition modelof the heat transfer coefficient may be written as follows:

htp ¼ Shnbc þ Fhf ð6Þ

The appearance of convective heat transfer in boiling in small tubesoccurs later than it does in larger tubes because of high boilingnucleation. Chen [17] introduced a multiplier factor, F = fn(Xtt), toaccount for the increase in the convective turbulence due to thepresence of the vapor phase. The function should be physically eval-uated again for flow boiling heat transfer in a small tube that haslaminar flow condition due to the small diameter effect. By consid-ering the flow conditions (laminar or turbulent) in relation to theReynolds number factor, F, Zhang et al. [1] introduced a relationshipbetween the factor F and the two-phase frictional multiplier that isbased on the pressure gradient of a liquid alone flow, /2

f . This rela-tionship is (F ¼ fnð/2

f Þ), where /2f is the general form for four flow

conditions according to Chisholm [19]. For the liquid–vapor flowconditions of turbulent–turbulent (tt), laminar–turbulent (vt), tur-bulent–laminar (tv) and laminar–laminar (vv), the values of theChisholm parameter, C, are 20, 12, 10, and 5, respectively. The valueof C in this study was obtained by considering the flow conditions oflaminar and turbulent with thresholds of Re = 2300 and Re = 3000for the laminar and turbulent flows, respectively. The laminar-tur-bulent transition Reynolds number was referred from Yang andLin [20].

0

1

2

3

4

5

6

0 0.2 0.4 0.6 0.8

300400500600

x

h tp

(kW

/m2 K

)R-22q = 10 kW/m2

Di = 1.5 mmTsat = 10oC

G (kg/m2s)

0

2

4

6

8

10

0 0.1 0.2 0.3 0.4 0.5 0.6 0.7

200300400500600

x

h tp

(kW

/m2 K

)

R-134aq = 10 kW/m2

Di = 3.0 mmTsat = 10oC

G (kg/m2s)

0

4

8

12

16

20

0 0.2 0.4 0.6 0.8 1

300400500600

x

h tp

(kW

/m2 K

)

R-410Aq = 20 kW/m2

Di = 3.0 mmTsat = 10oC

G (kg/m2s)(a) (b)

0

2

4

6

8

10

12

0 0.2 0.4 0.6 0.8 1

100120150180

xh t

p(k

W/m

2 K)

C3H8q = 15 kW/m2

Di = 3.0 mmTsat = 10oC

G (kg/m2s)(c) (d)

Fig. 4. Effect of mass flux on heat transfer coefficient: (a) R-22, (b) R-410A, (c) R-134a and (d) C3H8.

0

2

4

6

8

0 0.2 0.4 0.6 0.8

510

x

h tp

(kW

/m2 K

)

R-134aG = 120 kg/m2sDi = 0.5 mmTsat = 8oC

q (kW/m2)

0

5

10

15

20

25

0 0.2 0.4 0.6 0.8

102030

x

h tp

(kW

/m2 K

)

R-410AG = 500 kg/m2sDi = 1.5 mmTsat = 15oC

q (kW/m2)

0

2

4

6

8

10

0 0.2 0.4 0.6 0.8 1

5101520

x

h tp

(kW

/m2 K

)

C3H8G = 100 kg/m2sDi = 3.0 mmTsat = 10oC

q (kW/m2)

0

2

4

6

8

10

12

14

0 0.2 0.4 0.6 0.8 1

2030

x

h tp

(kW

/m2 K

)

CO2G = 300 kg/m2sDi = 3.0 mmTsat = 10oC

q (kW/m2)

(b)(a)

(d)(c)

Fig. 5. Effect of heat flux on heat transfer coefficient: (a) R-134a, (b) R-410A, (c) C3H8 and (d) CO2.

J.-T. Oh et al. / International Journal of Heat and Mass Transfer 54 (2011) 2080–2088 2085

The F factor in this study is developed as a function of /2f , where

/2f is obtained from Eq. (7)

/2f ¼

� dpdz F

� �tp

� dpdz F

� �f

¼ 1þ C� dp

dz F� �

g

� dpdz F

� �f

264

375

1=2

þ� dp

dz F� �

g

� dpdz F

� �f

¼ 1þ CXþ 1

X2

ð7Þ

The Martinelli parameter, X, is defined by the following equation:

X ¼� dp

dz F� �

f

� dpdz F

� �g

264

375

1=2

¼2f f G

2ð1� xÞ2qg=D

2f gG2x2qf=D

" #1=2

¼ ff

fg

� �1=2 1� xx

� � qg

qf

� �1=2

ð8Þ

The friction factor in Eq. (8) was obtained by considering laminar–turbulent flows, where f = 16Re�1 for laminar flow (Re < 2300) andf = 0.079Re�0.25 for turbulent flow (Re > 3000). Also, the liquid heattransfer is defined by existing liquid heat transfer coefficient corre-

0

5

10

15

20

25

0 0.2 0.4 0.6 0.8 1

689

x

h tp

(kW

/m2 K

)

R-134aG = 120 kg/m2sq = 20 kW/m2

Di = 0.5 mm

Tsat (oC)

0

5

10

15

20

25

30

0 0.1 0.2 0.3 0.4 0.5 0.6 0.7

110

x

h tp

(kW

/m2 K

)

R-410AG = 180 kg/m2sq = 10 kW/m2

Di = 0.5 mm

Tsat (oC)

3

4

5

6

7

0 0.2 0.4 0.6 0.8 1

09

x

h tp

(kW

/m2 K

)

C3H8G = 150 kg/m2sq = 25 kW/m2

Di = 3.0 mm

Tsat (oC)

0

2

4

6

8

10

12

14

0 0.2 0.4 0.6 0.8 1

110

xh t

p(k

W/m

2 K)

CO2G = 300 kg/m2sq = 30 kW/m2

Di = 3.0 mm

Tsat (oC)

(b)(a)

(d)(c)

Fig. 6. Effect of saturation temperature on heat transfer coefficient: (a) R-134a, (b) R-410A, (c) C3H8 and (d) CO2.

Table 3Physical properties of R-22, R-134a, R-410A, C3H8 and CO2 at 10 and 5 �C.

T (�C) Refrigerant P (MPa) qf (kg/m3) qg (kg/m3) qf/qg lf (10�6 Pa s) lg (10�6 Pa s) lf/lg r (10�3 N/m)

10 R-22 0.681 1247 28.82 43.27 195.7 11.96 16.36 10.22R-134a 0.415 1261 20.23 62.33 238.8 11.15 21.42 10.14R-410A 1.085 1130 41.74 27.07 146.6 12.91 11.36 7.16C3H8 0.636 515 13.8 37.32 113.8 8.15 13.96 8.85CO2 4.497 861.7 134.4 6.41 86.3 15.46 5.59 2.77

5 R-22 0.584 1264 24.79 50.99 206.7 11.73 17.62 10.95R-134a 0.350 1278 17.13 74.61 254.4 10.94 23.25 10.84R-410A 0.934 1151 35.43 32.21 156.0 12.60 12.38 7.90C3H8 0.551 522 11.98 43.57 119.8 7.97 15.03 9.48CO2 3.965 896.0 114.1 7.85 95.84 14.83 6.46 3.64

0

3

6

9

12

0 0.2 0.4 0.6 0.8

1.53.0

x

h tp

(kW

/m2 K

)

R-22G = 500 kg/m2sq = 20 kW/m2

Tsat = 10oC

Di (mm)

0

2

4

6

8

10

0 0.2 0.4 0.6 0.8 1

1.53.0

x

h tp

(kW

/m2 K

)

C3H8G = 200 kg/m2sq = 15 kW/m2

Tsat = 10oC

Di (mm)(b) (a)

Fig. 7. Effect of inner tube diameter on heat transfer coefficient: (a) R-22 and (b) C3H8.

2086 J.-T. Oh et al. / International Journal of Heat and Mass Transfer 54 (2011) 2080–2088

lations by considering flow conditions of laminar and turbulent. Forlaminar flow, Ref < 2300, the liquid heat transfer coefficient is ob-tained from the following correlation:

hf ¼ 4:36kf=D ð9Þ

For flow with 3000 6 Ref 6 104, the liquid heat transfer coefficient isobtained from Gnielinski [21],

hf ¼ðRef � 1000ÞPrf ðff=2Þðkf=DÞ1þ 12:7 Pr2=3

f � 1� �

ðff=2Þ0:5ð10Þ

For turbulent flow with 1046 Ref 6 5 � 106, the liquid heat transfer

coefficient is obtained from Petukhov and Popov [22],

hf ¼Ref Prf ðff=2Þðkf=DÞ

1þ 12:7 Pr2=3f � 1

� �ðff=2Þ0:5

ð11Þ

0

4

8

12

16

20

0 0.2 0.4 0.6 0.8 1x

h tp

(kW

/m2 K

)

G = 400 kg/m2sq = 20 kW/m2

Di = 3.0 mmTsat = 10oC

RefrigerantsR-22R-134aR-410ACO2

1

2

3

4

5

6

0 0.2 0.4 0.6 0.8 1x

h tp

(kW

/m2 K

)

G = 200 kg/m2sq = 10 kW/m2

Di = 1.5 mmTsat = 10oC

RefrigerantsR-134aC3H8

(b)(a)

Fig. 8. Comparison heat transfer coefficient for the present working refrigerants.

Table 4Deviation of the heat transfer coefficient comparison with some existing correlations.

Previous correlations Mean deviation (%) Average deviation (%)

Gungor–Winterton [13] 25.76 16.66Jung et al. [14] 26.83 18.01Shah [15] 27.31 �1.70Tran et al. [2] 27.57 0.56Chen [16] 38.62 28.78Wattelat et al. [17] 44.90 43.29Kandlikar [18] 51.55 �37.22Zhang et al. [1] 84.48 84.22

Table 5Summary of the new heat transfer coefficient correlation.

htp = Shnbc + Fhf

S ¼ 0:279ð/2f Þ�0:029Bo�0:098

J.-T. Oh et al. / International Journal of Heat and Mass Transfer 54 (2011) 2080–2088 2087

Dittus and Boelter [23] correlation is used for turbulent flow withRef P 5 � 106

hf ¼ 0:023Re0:8f Pr0:4

f kf=D ð12Þ

And then a new factor F, as is shown in Fig. 9, is developed using aregression method.

F ¼MAX 0:0:23/2:2f þ 0:76

� �;1

ð13Þ

The prediction of the nucleate boiling heat transfer for the pres-ent experimental data used the Cooper [24] correlation, which is apool boiling correlation

h ¼ 55P0:12red ð�0:4343 ln PredÞ�0:55M�0:5q0:67 ð14Þ

where the heat flux, q, is in W m�2. Kew and Cornwell [10] and Junget al. [25] showed that the Cooper [24] pool boiling correlation bestpredicted their experimental data.

Chen [16] defined the nucleate boiling suppression factor, S, as aratio of the mean superheat, DTe, to the wall superheat, DTsat. Jung

1

10

100

1000

1 10 100 1000

h tp/

h lo

Experimental DataZhang et al. (2004) Factor FNew Factor F

F=MAX(0.64φf, 1)

F=MAX[(0.023φf2.2+0.76), 1]

2fφ

Fig. 9. Convective heat transfer multiplier as a function of /2f .

et al. [14] proposed a convective boiling heat transfer multiplier fac-tor, N, as a function of quality, heat flux, and mass flow rate (repre-sented by employing Xtt and Bo) to represent the strong effect ofnucleate boiling in flow boiling in comparison with that in nucleatepool boiling, hnbc/hnb. The Martinelli parameter, Xtt, is replaced by atwo-phase frictional multiplier, /2

f , in order to consider laminar flowin small tubes. By using the experimental data of this study, a newnucleate boiling suppression factor, as a ratio of hnbc/hnb, is proposedas follows:

S ¼ 0:279ð/2f Þ�0:029Bo�0:098 ð15Þ

A new heat transfer coefficient correlation, as summarized inTable 5, is developed using a regression method with 1588 datapoints. The experimental heat transfer coefficient, htp,exp, and thepredicted heat transfer coefficient, htp,pred, are compared inFig. 10. The new correlation agrees closely with the comparisonwith a mean deviation of 15.28% and an average deviation of�0.48%.

5. Concluding remarks

Convective boiling pressure drop and heat transfer experimentswere performed in horizontal small tubes with refrigerants R-22,R-134a, R-410A, C3H8 and CO2. Wang et al. [5] predicted the flow

F ¼MAX ð0:023/2:2f þ 0:76Þ;1

h ihnbc ¼ 55P2:12

r ð�0:4343lnPrÞ�0:55M�0:5q0:67, where q is in W m�2

hf

¼ 4:36 kfD if Ref < 2300

¼ ðRef�1000ÞPrfff2

� �kfD

� �1þ12:7 Pr2=3

f�1ð Þ Ff

2

� �0:5 if 3000 � Ref � 104

¼ Ref Prfff2

� �kfD

� �1þ12:7 Pr2=3

f�1ð Þ Ff

2

� �0:5 if 104 � Ref � �106

¼ 0:023 kfD

Gð1�xÞDlf

h i0:8 Cpf lfkf

� �0:4if Ref � �106

8>>>>>>>>>>>>>>><>>>>>>>>>>>>>>>:

/2f ¼ 1þ C

F þ 1X2

X ¼ fffg

� �1=21þx

x

� � qgqf

� �1=2

C

¼ 5 for Ref < 2300 and Reg < 2300¼ 10 for Ref > 3000 and Reg < 2300¼ 12 for Ref < 2300 and Reg > 3000¼ 20 for Ref > 3000 and Reg > 3000

8>><>>:

f ¼ 16Re�1 for Re < 2300¼ 0:079Re�0:25 for Re > 3000

�

0

3

6

9

12

15

0 3 6 9 12 15

hexp (kW/m2K)

h pre

d(k

W/m

2 K)

+20%

-20%

MD = 15.28% AD = -0.48%

Fig. 10. Diagram of the experimental heat transfer coefficient, hexp, vs predictedheat transfer coefficient, hpred, from the new heat transfer coefficient correlation.

2088 J.-T. Oh et al. / International Journal of Heat and Mass Transfer 54 (2011) 2080–2088

pattern based on the current experimental results better than Woj-tan et al. [6]. Wang et al. [5] map showed a better prediction offlow pattern for the beginning of annular flow. Wojtan et al. [6]map showed a better prediction of flow pattern for the dry-outcondition for the R-410A experimental data. Compared with theflow pattern predicted by Wang et al. [5] and Wojtan et al. [6] flowpattern maps, the current experimental data show that annularflow and dry-out occur at lower vapor quality for the evaporationcondition with higher heat flux, higher mass flux and smaller innertube diameter.

Mass flux had an insignificant effect on the heat transfer coeffi-cient at low quality region, but had a significant effect on the heattransfer coefficient at high quality region. Heat flux had a signifi-cant effect on the heat transfer coefficient at the low quality region,but had an insignificant effect on the heat transfer coefficient at thehigh quality region. Heat transfer coefficient increased with in-creased saturation temperature. Heat transfer coefficient was high-er in the smaller inner diameter tube, especially at the low qualityregion. The heat transfer coefficient in the evaporation conditionwith small tubes was dominated by the nucleate boiling heattransfer contribution. CO2 had the highest heat transfer coefficientamong the working refrigerants. The mean heat transfer coefficientratio of R-22, R-134a, R-410A, C3H8 and CO2 was approximately1.0:0.8:1.8:0.7:2.0. The average heat transfer coefficient increasedwith a smaller inner tube diameter and with increased saturationtemperature. The Gungor–Winterton [13], Jung et al. [14], Shah[15] and Tran et al. [2] models predicted the current experimentalheat transfer coefficient relatively well.

The geometric effect of a small tube must be considered in thedevelopment of a new heat transfer coefficient correlation. Lami-nar flow appeared for flow boiling in small tubes, so the modifiedcorrelation of the multiplier factor for the convective boiling con-tribution, F, and the nucleate boiling suppression factor, S, are

developed in this study using laminar and turbulent flows consid-eration. The new boiling heat transfer coefficient correlation basedon a superposition model for refrigerants in small tubes had 15.28%mean deviation and �0.48% average deviation.

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