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Flow boiling heat transfer characteristics of R407Cinside a smooth tube with different tubeinclinations

Arijit Kundu*, Ravi Kumar, Akhilesh Gupta

Mechanical and Industrial Engineering, Indian Institute of Technology, Roorkee 247667, India

a r t i c l e i n f o

Article history:

Received 6 February 2014

Received in revised form

9 June 2014

Accepted 11 June 2014

Available online 19 June 2014

Keywords:

Refrigerants

Flow boiling

R407C

Smooth tube

Inclined

Heat transfer coefficient

* Corresponding author. Tel.: þ91 844 9498 5E-mail addresses: 8449498566ak@gmail.c

http://dx.doi.org/10.1016/j.ijrefrig.2014.06.0090140-7007/© 2014 Elsevier Ltd and IIR. All rig

a b s t r a c t

This paper presents the results of experimental investigations carried out with refrigerant

R407C during flow boiling inside a smooth tube of inside diameter 7.0 mm at different tube

inclinations. The experiments included five different tube inclination angles from 0� to 90�

of the direction of refrigerant flow. The test runs were carried out at inlet temperature

range of 6 �C to 9 �C, refrigerant mass velocities from 100 to 300 kgm�2 s�1 and heat fluxes

between 3 and 6 kW m�2. The average quality of boiling refrigerant was between 0.1 and

0.9. The influences of the above parameters on the boiling heat transfer coefficient with

tube inclinations are presented. The heat transfer coefficients predicted by some available

models and correlations in the open literature are compared with the present data. An

empirical correlation has also been developed to predict the heat transfer coefficient of

R407C during flow boiling inside an inclined plain tube.

© 2014 Elsevier Ltd and IIR. All rights reserved.

Caract�eristiques du transfert de chaleur du R407C �a l'int�erieurd'un tube lisse �a diff�erentes inclinaisons du tube

Mots cl�es : Frigorig�enes ; Ecoulement en �ebullition ; R407C ; Tube lisse ; Inclin�e ; Coefficient de transfert de chaleur

1. Introduction

Depletion of ozone layer shielding the earth's surface by the

chlorine called for a gradual phase-out of halogenated re-

frigerants in the recent past (Montreal Protocol, 1992). As a

consequence, CFCs (chlorofluorocarbons) and HCFCs

66; fax: þ91 1332 285665.om, ak261dme@iitr.ac.in

hts reserved.

(hydrochlorofluorocarbons) have been replaced by a new family

of substances which are chlorine free, hence undisruptive to-

wards the ozone layer. Among these new ancestors of re-

frigerants, the HFC blend like R407C has been found a suitable

replacement for R22 which has been used largely by industries

and home appliances with only 8e14% better performance

(Aprea and Greco, 2003) but with larger Ozone Depletion

(A. Kundu).

Nomenclature

Bo boiling number (q/G hfg, dimensionless)

CP specific heat (J kg�1 K�1)

D diameter (m)

E two phase convection multiplier (dimensionless)

g gravitational acceleration (m s�2)

G mass velocity (kg m�2 s�1)

hev evaporative heat transfer coefficient (W m�2 K�1)

h specific enthalpy (J kg�1)

hfg latent heat (J kg�1)

I electric current (A)

k thermal conductivity (W m�1 K�1)

L length of tube in flow direction (m)_m total mass flow rate (kg s�1)

M molecular weight (dimensionless)

P pressure (Pa)

Pr Prandtl number (CP m/k, dimensionless)

Pr reduced pressure (PSat/PC, dimensionless)

q heat flux (W m�2)

Q heat transfer rate (W)

Re Reynolds number (GD/m, dimensionless)

S nucleate boiling suppression factor

(dimensionless)

T temperature (�C)t time (s)

V electric potential (V)

x vapor quality (dimensionless)

Xtt Martinelli parameter ([(1 � x)/x]0.9[rg/rf]0.5[mf/mg]

0.1,

dimensionless)

z length (m)

Greek letters

a tube inclination (�)b mass transfer coefficient (m s�1)

ε vapor void fraction (dimensionless)

m dynamic viscosity (Pa-s)

h kinematic viscosity (m2 s�1)

r density (kg m�3)

s surface tension (N m�1)

d liquid film thickness (m)

Subscripts

avg average

C critical

f saturated liquid

g saturated gas

IA intermittent to annular flow transition

i inside

id ideal value

in inlet

out outlet

bp boiling point

CB convective boiling

EXP experimental value

MOD modified

NB nucleate boiling

PH pre-heater

PRED predicted value

Sat saturation

TS test section

t top side

b bottom side

sl left side

sr right side

i n t e rn a t i o n a l j o u rn a l o f r e f r i g e r a t i o n 4 5 ( 2 0 1 4 ) 1e1 22

Potential (ODP ¼ 0.05, Bitzer Refrigerant Report, 2007). The

R407C, a ternary blendofR32, R125andR134a (23%, 25%and52%

by weight, respectively) does not contain chlorine (zero ODP)

with modest Global Warming Potential (GWP ¼ 1774, Bitzer

Refrigerant Report, 2007). The little deficiency of R407C with

lower energetic performance (Aprea and Greco, 2003) with

respect to R22 can be overcome with optimizing the design of

evaporators and condensers. To enhance the heat transfer rate

and the cycle efficiency, detailed information of heat transfer

coefficient and pressure drop data is required.

A well designed evaporator improves the performance of

the whole refrigeration system. Thus, an efficient evaporator

is required to be designed to enhance heat transfer charac-

teristics and cycle efficiency of the system. A number of in-

vestigations have carried out to improve the performance of

evaporator using integral fin tubes (Passos et al., 2003),

different types of fin geometry (Chamra et al., 1996), and

refrigerant with lubricating oils (Cho and Tae, 1999). The

experimental conditions under which heat transfer co-

efficients were determined reflect a typical working situation

for the small-scale refrigeration system. In each of the cases

aforesaid use to bear with more pressure drop, and thus more

initial and running costs (Kouhikamali et al., 2012) than that of

smooth tubes. Thorough scrutiny of the literature reveals that

a lot of work (Aprea et al., 2008; Lallemand et al., 2001; Passos

et al., 2003; Torrella et al., 2006; Wang and Chiang, 1997) has

been done to enhance the heat transfer rate of the refrigerant

in the two-phase region through horizontal tubes. Very few

studies have been reported for heat transfer inside micro-fin

tube with different tube inclinations other than the horizon-

tal and vertical positions of the evaporator tube. Akhavan-

Behabadi et al. (2011) showed that the tube inclination af-

fects boiling heat transfer coefficient for pure fluid R134a at

different vapor quality.

A recent review (Cheng and Mewes, 2006) shows that

though many authors have worked to describe heat transfer

characteristics in flow boiling, heat transfer coefficient trends

are yet not comprehended fully. Even the effect of flowpattern

and gravity on the heat transfer coefficient was not expressed

properly for the forced flow boiling in a vertical tube, though

the influence of shear stress and gravitational force is pre-

dominant for different flow regimes in evaporation of re-

frigerants inside small diameter tubes (Hong et al., 1998). Thus

further study on the effect of gravity on flow boiling of

refrigerant mixtures through small channels is indispensable.

Therefore, the purpose of the present study is to provide the

effect imposed heat flux, vapor quality, mass velocity of

refrigerant and gravity on two-phase flow heat transfer

characteristics for evaporation of R407C in a smooth tube at

different tube inclinations.

i n t e r n a t i o n a l j o u r n a l o f r e f r i g e r a t i o n 4 5 ( 2 0 1 4 ) 1e1 2 3

2. Experimental facility and procedure

The schematic diagram of the experimental facility is shown

in Fig. 1. The test apparatus consists of a semi-hermetic

compressor, water-cooled condenser, an expansion device,

pre-heater, post-heater and test-evaporator. A reciprocating

compressor was used for compressing the refrigerant coming

from the evaporator. The refrigerant enters after the

compressor to a condenser which was counter flow water

cooled double pipe heat exchanger. After the condenser,

Coriolis-effect mass flowmeter has been installed to measure

very accurate mass flow rate of the refrigerant. The mass flow

rate has been controlled through hand shut off valves fitted on

the by-pass line incorporated after the condenser. A quality

filter-dryer was accommodated after the flow meter to entrap

lubricating oil, foreign particles and moisture in refrigerant. A

suitable pre-evaporator was designed and installed to control

the vapor quality at the inlet of test evaporator. By regulating

power supplywith variable voltage AC, heat source, heat input

to pre-evaporator has been controlled. An oil separator was

incorporated downstream of the compressor for oil free

refrigerant to enter in the test evaporator section. A suitable

accumulator was also installed upstreamof the compressor to

ensure that the oil free refrigerant vapor should enter the

compressor.

The test section was made of smooth copper tube with

inner diameter of 7 mm and outside diameter of 9.52mm. The

outside tube wall temperatures were measured by T-type

copper-constantan thermocouples at five axial stations,

including inlet and exit to the test evaporator tube. At each

station, the temperatures weremeasured at the top, two sides

of the middle and bottom position. The average of these

temperatures indicates the local wall temperatures. The local

saturation pressures at the inlet and outlet of test evaporator

weremeasured by piezoelectric pressure transducers having a

range of 0e20 bar. The average of these two pressures in-

dicates the saturation pressure at test section. The test

evaporator tube was heated by flexible Nichrome (Nickel-

eChromium 80e20 percent by weight) heater wire of 3.2 kW

capacity at 8m lengthwrapped overmica thin tapewhichwas

wound around the tube over the full test length of 1.2 m. Heat

Fig. 1 e Schematic of the experimental facility.

flow to the heater was monitored with a variety AC voltage

controller and current flow was measured by standard clamp

meter (accuracy ±2.0% of reading ±5digits) to determine

applied heat flux. To calculate enthalpy at entry to pre-

evaporator, another pressure transducer and T-type thermo-

couplewere fitted at a section just prior to pre-evaporator. The

evaporator tube and pre-evaporator were well insulated with

glass wool from outside to ensure lesser heat loss.

The ranges of experimental parameters are listed in Table 1.

The temperature and pressure measurement were recorded

through a data acquisition system. The physical properties of

refrigerantswere obtained fromREFPROP (Lemmon et al., 2007).

3. Data reduction

The ‘quasi-local’ heat transfer coefficient was calculated by

using following Equation (1):

hev ¼"pDiL

�Two;avg � TSat

�QTS

� Di

2

�ln½Do=Di�

k

�#�1

(1)

This is simply based on heat generation by electrical

resistance heating of heater wire outside evaporator tube wall

and conduction heat flow through the tube wall to the

refrigerant flowing inside at steady state condition. The

outside tubewall temperature, Two, is themean of outsidewall

temperature at the top, two sides of the middle and bottom of

the tube at each section in the test evaporator as:

Two ðzÞ ¼TtðzÞ þ TslðzÞ þ TsrðzÞ þ TbðzÞ

4(2)

TSat is the saturation temperature corresponding to the

pressure readings, taking the mean of the dew and bubble

point temperature for refrigerant mixture R407C and Two;avg is

the mean of the local wall temperatures, Two ðzÞ. Accountingthe heat loss to the ambient by the heating coefficient x,

defined as the ratio between the output power and input

power, the resistance heat flow Q, to the refrigerant from

outside of the tube (pre-heater and test section) has been

calculated by using following Equation (3):

Q ¼ x$V$I (3)

x (typically around 0.935) is experimentally determined

using the method proposed by Wambsganss et al. (1993).

Average vapor quality is calculated by following Equation (4):

xavg ¼ ðxin þ xoutÞ=2 (4)

where, xin is the inlet dryness fraction of refrigerant to the

test section, which is calculated by following Equation (5):

xin ¼ hin � hf ;in

hfg;in(5)

Table 1 e Range of experimental parameters.

Refrigerant mass velocity 100e300 kg m�2s�1

Heat flux 3e6 kW m�2

Vapor quality 0.1e0.9

Inlet temperature 6 to 9 �C

Table 2 e Uncertainty of variables.

Primary measurements Derived quantities

Parameter Uncertainty Parameter Uncertainty

Diameter ±2 mm Reynolds number ±0.6%Pressure ±0.5% Mass velocity ±0.6�2%

Temperature ±0.2 �C Heat transfer coefficient ±3.9�11%

Mass flow

rate

±0.2% Vapor quality ±2.0�9.5%

Heat flux ±2.1�3.6% Electrical power ±0.80%

i n t e rn a t i o n a l j o u rn a l o f r e f r i g e r a t i o n 4 5 ( 2 0 1 4 ) 1e1 24

hin is the specific enthalpy of the refrigerant at the entry to

the test section and outlet of the pre-heater, which is deter-

mined by Equation (6) applying an energy balance on the pre-

heater:

hin ¼ h1 þ QPH= _m (6)

Enthalpy h1 is found with respect to the measured tem-

perature and pressure at the entry to the pre-heater and hfg is

the latent heat of vaporization of the refrigerant. xout is the

exit dryness fraction of refrigerant from test section can be

determined from Equation (7):

xout ¼ hout � hf ;out

hfg;out(7)

hout is the exit enthalpy of the refrigerant is determined by

applying an energy balance on the test section as:

hout ¼ hin þ QTS= _m (8)

The error in measurement depends on the operating con-

ditions and mostly on the accuracy of the wall to saturation

temperature difference. The experimental uncertainty anal-

ysis was done by following Equation (9) proposed by Kline and

McClintock (1953):

UR ¼"XN

i¼1

�vRvVi

UVi

�2#0:5

(9)

where UR is the estimated uncertainty in calculating the value

of desired variable R, due to the independent uncertainty UVi

in the primary measurement of N number of variables Vi,

affecting the result. The experimental uncertainties for the

sensors are listed in Table 2. Maximum uncertainty in the

measurement of heat transfer coefficient is shown in Table 3.

Some additional uncertainty should be considered for the

change in composition of the refrigerant mixture during the

time of experiments, especially when phase change occurs, as

the provision for inspection of the compostion has not been

included for the present study.

Table 3 e Maximum uncertainty in heat transfercoefficient.

G (kg m�2 s�1) q (kW m�2) Uncertainty (%)

100 6.0 ±11200 4.5 ±9300 3.0 ±8.5

4. Results and discussion

The latest version of the flow pattern map of Wojtan et al.

(2005) is used for the evaluation of two phase flow regimes.

Since, the heat transfer process depends upon the flow

pattern; one must distinguish the flow pattern regimes during

the flow boiling experiments as shown in Fig. 2 for the present

study. These maps provide the transition boundaries on a

linear relationship of mass velocity versus vapor quality for

the refrigerant R407C at TSat ¼ 5.658 �C. It can be seen from the

figure that the flowpattern is slug and stratifiedwavy from the

inlet section to a quality x < xIA ¼ 0.3615 depending on test

conditions. Clearly, a stratifiedwavy flow is achieved at higher

vapor content. At higher quality values (>40%), flow pattern

continues to the developed annular flow, with a partial dry-

out where the liquid film around the periphery of the tube

goes away.

The variation of flow boiling heat transfer coefficient of

R407C with vapor quality at the horizontal position of the test

evaporator has been shown in Fig. 3. The effect of mass ve-

locity at a fixed heat flux of 3 kW m�2 and 6 kW m�2 can be

observed from the figure. The heat transfer coefficient in-

creases with mass velocity but decreases with a rise in vapor

quality after reaching a maximum value around x ¼ 0.35 to

0.45 for mass velocity of 100 kg m�2 s�1 and around x ¼ 0.42 to

0.48 for mass velocity of 300 kg m�2 s�1. The heat transfer

development in flow boiling results from the interaction be-

tween nucleate boiling and liquid convective boiling (Wang

and Chiang, 1997). Indeed, increasing the refrigerant mass

velocity increase the fluid velocity, thus raising the convective

boiling contribution. In a very low quality region, where

nucleate boiling is dominant, the effect of the refrigerantmass

velocity is feeble and the heat transfer coefficients tend to

come closer. The interfacial shear stress is amplified by

increasing the mass velocity, and the bubbles grown on the

heated wall are carried away from the surface under the in-

fluence of shear stress without developing to the larger di-

ameters as encountered in the case of pool boiling. This

results in a reduction in the nucleate boiling component,

which can be seen from Fig. 3. Due tomass transfer resistance,

nucleate boiling heat transfer coefficient for zeotropic

Fig. 2 e Flow pattern map for R407C at TSat ¼ 5.658 �C.

Fig. 3 e Effect of mass velocity on heat transfer coefficient

in horizontal evaporator tube.

i n t e r n a t i o n a l j o u r n a l o f r e f r i g e r a t i o n 4 5 ( 2 0 1 4 ) 1e1 2 5

mixtures are considerably lower than single component re-

frigerants (Wang and Chiang, 1997). At low mass velocities,

major contribution to the heat transfer mechanism is the

nucleate boiling. The higher fluid velocity, as the mass ve-

locity increases, inducts liquid entrainment to agitate the

mass transfer resistance (Kattan et al., 1995, 1998).

Fig. 4 shows the effect of tube inclination angles on the

heat transfer coefficient at a heat flux of 3 kW m�2. The mass

velocity of R407C remained constant at G¼ 100 kgm�2 s�1. For

all tube inclinations, heat transfer coefficient increases up to

vapor quality x ¼ 0.35 to 0.45 and then decreases because of

dry-out. The same trend was observed by Wang and Chiang

(1997), where they observed the heat transfer coefficient rea-

ches maximum only at 20% average quality and then de-

creases at low heat flux. For horizontal tube, their results vary

from the present study because of, as they had explained, the

overestimation of wall superheat by the calculation of heat

transfer coefficient based on initial bubble point temperature

and suggested to use a linear approximation of modified bulk

fluid temperature: TSat.MOD ¼ TBub,initial þ x$Dq; where Dq is the

temperature glide at the corresponding saturation pressure.

Fig. 4 e Effect of tube inclination on heat transfer coefficient

at mass velocity of 100 kg m¡2 s¡1and heat flux of

3 kW m¡2.

The effect of heat flux with the tube inclination angles on

heat transfer coefficient has been shown in Figs. (5e7). The

convective boiling contribution dominates heat transfer

mechanism at low heat fluxes (Greco, 2008). Nucleate boiling

prevails in the heat transfer process as the heat flux increases.

Figs. 5 and 6 show the variation of the local heat transfer co-

efficient for different tube inclinations at mass velocity of

200 kgm�2 s�1and fixed heat flux of 4.5 kWm�2 and 6 kWm�2,

respectively. In each case, heat transfer coefficient increases

with increase in heat flux for all the tube inclinations.

In the present investigation, at low mass velocity and low

imposed heat flux inside the evaporator tube, heat transfer

coefficient increases with vapor quality up to 30%e35% of the

tube. This matter of increment in heat transfer coefficient

remains the same for all tube inclinations. When mass ve-

locity increases, the trend in increment continues to vapor

quality of 35%e45% of the tube. After that, heat transfer co-

efficient decreases till the fluid leaves the evaporator tube.

Fig. 7 shows that the heat transfer coefficient decreases after

reaching maximum at vapor quality about 40%e46% for mass

velocity G ¼ 300 kg m�2 s�1and heat flux q ¼ 6 kW m�2 for

different tube inclination angles. Nucleate boiling contribu-

tion can be explained by the active nucleation sites AN (Carey,

2007) as shown in Equation (10); where the liquid convection

contribution can be described by the property combination j

(shown in Equation (11), according to the Dittus-Boelter, sin-

gle-phase forced convection correlation).

AN ¼ 2TSatsf

hfgrg(10)

j ¼ CPf

mf

!0:4

$k0:6f (11)

A comprehensive exploration of convective boiling and

nucleate boiling heat transfer component (as shown in Table

4) separately reveals that as evaporation proceeds through

the length of the tube and vapor quality increases, due tomass

transfer resistance, convective part increases; but nucleate

boiling contribution decreases with the increase in mass

Fig. 5 e Effect of tube inclination on heat transfer coefficient

at mass velocity of 200 kg m¡2s¡1and heat flux of 4.5

kW m¡2.

Fig. 6 e Effect of tube inclination on heat transfer coefficient

at mass velocity of 200 kg m¡2s¡1and heat flux of

6 kW m¡2.

Fig. 7 e Effect of tube inclination on heat transfer coefficient

at mass velocity of 300 kg m¡2s¡1and heat flux of

6 kW m¡2.

i n t e rn a t i o n a l j o u rn a l o f r e f r i g e r a t i o n 4 5 ( 2 0 1 4 ) 1e1 26

velocity. The mass transfer resistances to convective evapo-

ration at high mass velocities use to be reduced significantly

due to the disturbances at the mass transfer resistance

boundary created by the highly turbulent interaction between

liquid and vapor interface. It is clear from Fig. 8 and Fig. 9 that

to increase in mass velocity from 200 kg m�2 s�1 to

300 kg m�2 s�1 with the same imposed heat flux, heat transfer

coefficient increases much more hurriedly due to more

Table 4 e Thermo-physical properties of test refrigerant.

G P sf hfg rg C

100 618,000 0.01257 209,820 26.30 142

200 647,000 0.01003 208,370 27.53 143

300 686,000 0.00975 206,480 29.19 143

a units: [K-(N-m)/(J kg�1)-(kg m�3)].b units: [(J kg�1-K)-(W m�1-K)/(Pa-s)].

contribution of the convective part in the heat transfer pro-

gression process for all the tube inclinations.

During flow boiling in an inclined tube the mass velocity

and tube inclination affect the vapor quality at which boiling

crisis occurs (Akhavan-Behabadi et al., 2011). The same can be

observed from the present study as shown in Figs. (4e7). For

mass velocity G ¼ 100 kg m�2s�1, dry out occurs at vapor

quality about 34%e38% of the evaporator tube with tube in-

clinations 0� to 45� for constant heat flux of 3 kWm�2. Dry-out

occurs at x ¼ 40e44% for the tube when tube inclination in-

creases to 60 �e90�. When mass velocity increases to

200 kgm�2s�1, the dry-out phenomena occurs at x¼ 0.41e0.43

at 30 �e45� tube inclination and it varies to about x¼ 0.45e0.48

to increase in inclination to 60 �e90� for a heat flux of

6 kW m�2. The cause of early dry-out is the termination of

nucleate boiling because of the local void disseminates as a

vapor drape on the heating wall for the lower vapor velocity

attained by the lesser volatile components of the refrigerant

mixture R407C during the evaporation at the liquidevapor

interface; and the bubble crowding impairs the surface cooling

by clogging the incoming liquid layer.

The variation of vapor quality for the maximum heat

transfer coefficient with the tube inclination at different mass

velocities of R407C is depicted in Fig. 8. From Fig. 9 it is obvious

that the tube inclination influences the average heat transfer

coefficient significantly. The figure shows the effect of tube

inclination angle on the average heat transfer coefficient for

fixed heat flux of q ¼ 6 kW m�2 with mass velocity varying

from 100 to 300 kg m�2s�1. It reveals that for all mass veloc-

ities, the heat transfer coefficient is the maximum at 90� of

tube inclination. Average heat transfer coefficient increases to

only 5% and 8% for tube inclination of 30� and 45�, respec-tively, with respect to the horizontal position of the evapo-

rator tube at a low mass velocity of 100 kg m�2s�1 with heat

flux of 6 kW m�2 which lies within the uncertainty of 11% in

experimental value. But the average heat transfer coefficient

increases to 12% for tube inclination of 60� and 15% for 90�

tube inclination with respect to the horizontal position of the

evaporator tube. The effect of tube inclination goes downwith

increment in mass velocity after the dry-out occurs. Before

dry-out, the average heat transfer coefficient increases with

increase in mass velocity and the difference also increases. In

comparison to the horizontal position of the evaporator tube,

average heat transfer coefficient in inclined tube increases

2%e11% for mass velocity of 300 kg m�2s�1for varying tube

inclination from 30� to 90�. When mass velocity increases to

200 kg m�2s�1, the increment in average heat transfer coeffi-

cient is only up to 3%e14% of the horizontal tube. Before dry

out, the average heat transfer coefficient increases about

8e31% of tube inclination increases from 30� to 90� with

Pf mf kf ANa jb

5.8 0.00020393 0.09494 1.2699e�06 133.14

1.3 0.00020029 0.09423 0.0980e�06 133.70

8.6 0.00019568 0.09330 0.0912e�06 134.43

Fig. 8 e Variation of vapor quality for the maximum heat

transfer coefficient with tube inclination at different mass

velocities.

Fig. 10 e Effect of mass velocity on heat transfer coefficient

at tube inclination of 60� and 30�.

i n t e r n a t i o n a l j o u r n a l o f r e f r i g e r a t i o n 4 5 ( 2 0 1 4 ) 1e1 2 7

respect to horizontal tube at a particular mass velocity. Ver-

tical flow of R407C has the worst heat transfer coefficient for

all mass velocities irrespective of heat flux after dry out.

Average heat transfer coefficient, in general, is defined as the

mean of local heat transfer coefficients at different vapor

qualities for a given imposed heat flux and constant mass

velocity.

The comparison of the flow boiling performance inside the

tube with 30� inclination with the evaporator tube inclined at

60� for mass velocity varies from 100 kg m�2 s�1 to

300 kg m�2 s�1 at a fixed heat flux of 4.5 kW m�2 has been

shown in Fig. 10. Indeed, the local heat transfer coefficients

increase 17e26% for increase in mass velocity from

100 kg m�2s�1 to 300 kg m�2 s�1 for a ¼ 30�; but those vary to

20e33% to increase the inclination to 60�. However, the

average heat transfer coefficient is 21% higher at the tube

inclination of 30� and 25% higher for a ¼ 60� with the same

mass velocity variation inside the tube. The vapor phase

Fig. 9 e Variation of average heat transfer coefficient with

tube inclinations at q ¼ 6 kW m¡2.

velocity has been reduced for low vapor quality resulting in a

corresponding decrease in inertia force. On low vapor quality,

thus, the gravitational force affects the two-phase flow in a

considerable manner and this causes the variation of heat

transfer coefficient for different tube inclinations. Also the

variation of heat transfer coefficient in vapor quality for a

particular tube inclination has uneven behavior due to the

transition of different flow pattern through the growth of the

boiling along the tube length. From Fig. 10, it is evident that for

the same heat flux, as mass velocity increases, inclination in

evaporator tube facilitates the boiling performance of refrig-

erant mixture R407C. The cause may be the buoyancy effect,

which acts in the direction of the fluid flow accelerating the

upstream flow as the tube inclination increases and enhances

the convective transport and the ratio of gravitational and

frictional pressure drop plays an important role as the evap-

oration proceeds. This phenomenon also has been demon-

strated by Kattan et al. (1998) by describing the lower heat

transfer coefficient for downward flowwith respect to upward

and horizontal flow of pure refrigerant. Also the adverse effect

of flow stratification in horizontal tube has been decreased by

the tube inclination in upward direction (Akhavan-Behabadi

et al., 2011).

However, the interfacial shear stress has been increased by

the turbulence created by the very high mass velocity and

thus, the bubbles formed on the heated surface of the tube

could not attain at a large size as seen in nucleate boiling. By

increasing themass velocity, the influence of nucleate boiling,

which is the dominant part of the low vapor quality region

became less effective to contribute in the flow boiling of

refrigerant upstream. And for this, in the current study for

R407C, it is revealed that an eminent decrease in the local heat

transfer coefficient occurs from inlet to outlet of the evapo-

rator tube by 14e27% and 11e24% with the mass velocity

varies for 100 kg m�2s�1 to 300 kg m�2s�1, respectively,

depending on different tube inclinations.

Two-phase pressure drop gradients of R407C as a function

of vapor quality for two different refrigerant mass velocities

and tube inclinations are reported in Fig. 11. The pressure drop

Fig. 11 e Pressure gradient versus vapor quality for

G ¼ 100 kg m¡2 s¡1 and G ¼ 300 kg m¡2 s¡1.

Fig. 12 e Comparison of present R407C data with

established correlations.

i n t e rn a t i o n a l j o u rn a l o f r e f r i g e r a t i o n 4 5 ( 2 0 1 4 ) 1e1 28

is evaluated as the pressure drop per unit length between the

inlet and the exit from the test evaporator. For both the mass

velocities, it can be observed that the pressure drop increases

with both, the vapor quality and the tube inclination before

the dry-out takes place. Pressure drop increases with mass

velocity, as can be seen from the figure. The effect of heat flux

on pressure drop is very much negligible. In general, the total

pressure drop ismainly caused by the frictional pressure drop.

The static pressure drop can not be neglected for any incli-

nation of the evaporator tube other than that for horizontal

one. Also, the momentum pressure drop has an effect when

discussing the heat transfer performances in a evaporator. But

as the tube length is not so long, the major contribution to the

two-phase pressure loss would be the frictional loss. The

frictional pressure drop is a change of pressure ensuing from

the energy dissipated in the flow by friction, eddying etc.

(Aprea et al., 2008). As the flow goes downstream and vapor-

ization takes place, the density of the liquidevapor mixture

decreases. As a result, the flow acceleratesmuchmore and the

pressure drops increases. As soon as dry-out prevails, the

liquid layer leaves the tube wall; and with the increase in

vapor quality, the decrease in mean density because of the

lower density of the vapor, the pressure drop starts decreasing

which is seen from Fig. 11.

Present test data of R407C for horizontal test evaporator are

compared with established correlations proposed by

Kandlikar (1990), Liu and Winterton (1991) and Wojtan et al.

(2005) as shown in Fig. 12. The experimental results are

shown with the comparison of �50% to þ300% errors with a

mean deviation ranging from �37% to þ176%. Among them,

the mean deviation is the smallest by using Wojtan et al.

(2005) correlation. As can be seen, these established correla-

tions cannot predict the heat transfer coefficients for R407C

under the present test conditions mainly because of these

models do not account for the additional resistance due to the

mass transfer. Under this circumstances, new correlations to

predict the heat transfer coefficients of R407C for flow boiling

in smooth tube inclined at different inclinations have been

developed and discussed next.

5. Developing new correlation

Kattan et al. (1995) has used the dimensional reduced pressure

correlation of Cooper (1984) with some modifications as

shown in Equation (12):

hNB ¼ 55P0:12r

��log10Pr

��0:55M�0:5q0:67FC (12)

Analytical mass transfer resistance factor FC is a function

of boiling temperature range DTbp which is the difference

between bubble point temperature and dew point tempera-

ture for R407C. Jung et al. (2003) also showed that Cooper's poolboiling correlation proved the best prediction in their study.

By using the experimental data in this study, modified

nucleate boiling correlation is proposed as:

hNB ¼ 45P0:12r

��log10Pr

��0:55M�0:5q0:67FC (13)

Where FC is proposed by Thome (1989) as:

FC ¼(1þ

�hideal

q

�$DTbp$

"1� exp

�q

rf hfgbf

!#)�1

(14)

The mass transfer coefficient bf has a fixed value of

0.0003 ms�1 based on experimental studies on various hydro-

carbons and aqueous mixture and hideal is expressed by Equa-

tion (13)with FC¼ 1. Zurcher et al. (1998) also has compared the

above method successfully to their R407C flow boiling data.

Nucleate boiling of Zeotropic mixture is suppressed much

more than pure fluid as boiling progresses due to loss of wall

super heat at mixing of different volatile constituents. Gunjor

and Winterton (1989) expressed their flow boiling data for

vertical up-flows and down-flows of water, ethylene glycol

and several refrigerants with introducing nucleate boiling

suppression factor which was the function of Martinelli

parameter Xtt, Boiling number Bo, and liquid Reynolds num-

ber Ref. The flow pattern map presented by Kattan et al. (1998)

has described that for G ¼ 200 kg m�2s�1, intermittent flow

occurswhen x� 0.4 and annular flow startswith an increasing

rise for x > 0.4 where partial dry-out prevails after x � 0.93. A

distinct change in flow pattern and corresponding rise in heat

i n t e r n a t i o n a l j o u r n a l o f r e f r i g e r a t i o n 4 5 ( 2 0 1 4 ) 1e1 2 9

transfer coefficient also has been noticed in the present study.

To relate the experimental data in the present study, incor-

porating laminar effects on flow boiling, a new suppression

factor as a function of Prandtl number, Reynolds number and

vapor quality is, thus, proposed as:

for G � 200 ; S ¼h1þ x�1:3ð1� xÞ�1:8Pr�2:5

f Re0:009f

i�1

(15)

for G>200 ; S ¼h1þ x�1:2ð1� xÞ�2Pr�2:5

f Re0:009f

i�1(16)

The convective heat transfer coefficient correlation has

been presented by Dittus and Boelter (1930) and is written in

Equation (17):

hCB ¼ 0:023 Re0:8f Pr0:4f Kf

.Di (17)

The extensive study of Dittus and Boelter (1930) for flowing

of liquid alone using the mass velocity of G(1�x) in a tube of

qstrat ¼ 2p� 2

266664pð1� εÞ þ

�3p2

�1 =

3 n1� 2ð1� εÞ þ ð1� εÞ1=3 � ε

1=3o� 1200

εð1� εÞf1� 2ð1� εÞgn1þ 4

hε2 þ ð1� εÞ2

io377775 (23)

internal diameter Di gives much higher value due to higher

mass transfer resistance in boiling for refrigerant mixtures

than pure fluid. Kattan et al. (1998) had visualized the annular

ring of liquid inside tubemore realistically as a filmflow rather

than as tubular flow and confronted convective coefficient

with respect to liquid film thickness, d in place of tube internal

diameter and represented Equation (18) as:

hCB ¼ 0:0133 Reo:69L Pr0:4f Kf

.d (18)

Ref ¼ 4Gð1� xÞdð1� εÞmf

(19)

Fig. 13 e Two-phase flow cross-section of tu

d ¼ Di �"�

Di�2

� 2Af

#0:520)

2 2 2p� qdry

Af ¼ Að1� εÞ (21)

The cross sectional area occupied by liquid-phase,Af is the

function of cross-sectional area of the tube, A and vapor void

fraction, εwhich has been presented by Steiner (1993) as given

in Equation (22):

ε¼ xrg

8<:½1þ0:12ð1�xÞ�

xrgþ1�x

rf

!þ1:18

G

24gs

�rf�rg

�r2f

35

1=4

·ð1�xÞ9=;

�1

(22)

Wojtan et al. (2005) has presented the non-iterative geo-

metric expression for stratified angle qStrat which is shown in

Equation (23) to find dry angle qdry with the help of liquid

height Hf as shown in Fig. 13 and in Equations (24e26):

qdry ¼ 2p� 2 cos�1

�Di � 2Hf

D

�(24)

i

Hf ¼ 0:5Di

�1� cos

�2p� qstrat

2

�(25)

From the data acquired in the present study, similar

convective heat transfer coefficient correlations except with

minute modification in Equation (18) has been proposed and

that is written in Equation (26) where the value of liquid film

thickness (d) is given by Equation (14) for G � 200 and for

G > 200, d ¼ 0 for annular flow:

be; (a) stratified flow, (b) annular flow.

i n t e rn a t i o n a l j o u rn a l o f r e f r i g e r a t i o n 4 5 ( 2 0 1 4 ) 1e1 210

hCB ¼ 0:0023 RenLPr0:3f Kf

.d

(26)

Fig. 14 e Comparison of the entire experimental results to

the predicted results by new proposed correlation.

for G � 200; n ¼ 0:65 and for G> 200; n ¼ 0:6

Due to co-current flow of liquid and vapor in flow boiling of

refrigerant mixture through evaporator tube, an enhance-

ment factor should be introduced with proposed modified

Dittus and Boelter (1930) convective heat transfer coefficient

correlation Equation (26). A new enhancement factor E is

developed using regression methods. Factor E can be

expressed as written in Equations (27) and (28):

for G � 200 E ¼"1þ xð1� xÞPr0:01f

rf

rg� 1

!#0:1(27)

for G> 200 E ¼"1þ xð1� xÞ2 Pr0:01f

rf

rg� 1

!#0:01(28)

The current study includes a higher range of mass velocity

from 100 kg m�2s�1 to 300 kg m�2s�1 than that had been

considered in few past researches and also demarcates

annular flow and stratified-wavy flow for evaporation of

R407C through the inclined small diameter tube. Jung et al.

(2003) had employed correction factor as a function of Marti-

nelli parameter, Xtt, liquid and vapormole fraction and boiling

number, Bo. Present study proposes two separate correction

functions to distinguish stratified-wavy regime and slug

regime as:

for G � 200; fðxÞ ¼241þ 0:1x

6 =

5ð1� xÞ17=1035

0:5

(29)

fðx;KÞ ¼24x3 =

2ð1� xÞ7

=

5·K4 =

5

35

0:001

(30)

for G> 200; fðxÞ ¼241þ 0:1x

3 =

2ð1� xÞ235

0:5

(31)

fðx;KÞ ¼24xð1� xÞ3

=

2·K3 =

5

35

0:001

(32)

where K ¼ Bo�0.4

The two-phase heat transfer coefficient, htp is obtained

from an asymptotic expression that combines the nucleate

boiling and convective boiling contributions using an expo-

nent, n as suggested by Wattelet et al. (1994) shown in Equa-

tion (33):

htp ¼hnNB þ hn

CB

�1=n (33)

The value of n ¼ 2, has been found out by the regression

method with present test data points. Thus, it is proposed in

the present study that the two-phase flow heat transfer co-

efficient correlation is given by Equation (34):

htp ¼ 1:45fðxÞ·fðx;KÞ� ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi

ðE·hCBÞ2 þ ðS·hNBÞ2q �

·Fa (34)

The factor Fa has been incorporated to specify the exten-

sive effect of tube inclinations on flow boiling heat transfer

coefficient of R407C. It is expressed as a function of tube

inclination, a and vapor quality, xwhich is sturdily dependent

on the variation of mass velocity and heat flux as shown in

Equation (35e36):

for G � 200; Fa ¼ ½1þ 0:5xð1� xÞsin a· cosðaþ 25Þ� (35)

for G>200; Fa ¼ ½1þ 0:5xð1� xÞsin a· cosðaþ 50Þ� (36)

The above correlation given in Equation (34) predicts the

heat transfer coefficients for flow boiling of zeotropic refrig-

erant blend R407C through the tube inclined at any angles

from þ0� to þ90�; it showed good agreement of experimental

data with the calculated one for an error range of ±20% as

shown in Fig. 14. Mean percentage error (MPE), mean absolute

error (MAE), and root mean square error (RMSE) of the corre-

lation are defined by Equations (37e39) as discussed by Hall

and Mudawar (2000).

MPE ¼ 1P

XP

1

hevPRED � hevEXP

hevEXP

� 100% (37)

MAE ¼ 1

P

XP

1

����hevPRED � hevEXP

hevEXP

����� 100% (38)

RMSE ¼ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi1P

XP

1

�hevPRED � hevEXP

hevEXP

�2s

� 100% (39)

Where P is the number of heat transfer coefficient data points.

The mean percentage error furnishes an indication whether a

correlation is more likely to under-predict or over-predict the

heat transfer coefficient and the mean absolute error treats

the absolute error from each data point equally (Hall and

Mudawar, 2000). On the other hand, RMSE attributes better

weight to those predictions exhibiting larger deviations from

the experimental value (Hall and Mudawar, 2000). The new

correlation agrees closely compared to that of the experi-

mental values with an MPE of �3.68%, an MAE of 10.44% and

an RMSE of 12.88%. The summary of newpredicted correlation

is shown in Table 5.

Table 5 e Summary of new proposed correlation.

htp ¼ 1:45fðxÞ$fðx;KÞ½ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiðE$hCBÞ2 þ ðS$hNBÞ2

q�$Fa ; x ¼ mg

mfþmg; K ¼ Bo�0:4

hCB ¼ 0:0023 RenLPr0:3f Kf=d ; ReL ¼ 4Gð1�xÞd

ð1�εÞmf

hNB ¼ 45P0:12r ð�log10PrÞ�0:55M�0:5q0:67FC

FC ¼(1þ

�hidealq

�$DTbp$

"1� exp

�q

rf hfgbf

!#)�1

ε ¼ xrg

8<:241þ 0:12ð1� xÞ�

xrgþ 1�x

rf

!þ 1:18

G

24gsðrf�rgÞ

r2f

351=4

·ð1� xÞ9=;

�1

d ¼ Di2 �

"�Di2

�2

� 2Af

2p�qdry

#0:5where Af ¼ Að1� εÞ

For G � 200

n ¼ 0.65

qstrat ¼ 2p� 2

264pð1� εÞ þ

�3p2

�1 =

3

f1� 2ð1� εÞ þ ð1� εÞ1=3 � ε1=3g � 1

200 εð1� εÞf1� 2ð1� εÞgf1þ 4½ε2 þ ð1� εÞ2�g

375

qdry ¼ 2p� 2 cos�1

�Di�2Hf

Di

�; Hf ¼ 0:5Di

�1� cos

�2p�qstrat

2

�Fa ¼ [1 þ 0.5x(1 � x)sina cos(a þ 25)]

fðxÞ ¼

ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi241þ 0:1x

6 =

5

vuuut ð1� xÞ17=1035 ; fðx;KÞ ¼

24x3 =

2ð1� xÞ7

=

5 ·K4 =

5

350:001

E ¼"1þ xð1� xÞPr0:01f

rfrg� 1

!#0:1; S ¼ ½1þ x�1:3ð1� xÞ�1:8Pr�2:5

f Re0:009f ��1;

For G > 200

n ¼ 0.6; qdry ¼ 0

Fa ¼ [1 þ 0.5x(1 � x)sina cos(a þ 50)]

fðxÞ ¼

ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi241þ 0:1x

3 =2

vuuut ð1� xÞ235; fðx;KÞ ¼

24xð1� xÞ3

=2$K

3 =5

350:001

E ¼"1þ xð1� xÞ2Pr0:01f

rfrg� 1

!#0:01; S ¼ ½1þ x�1:2ð1� xÞ�2Pr�2:5

f Re0:009f ��1

i n t e r n a t i o n a l j o u r n a l o f r e f r i g e r a t i o n 4 5 ( 2 0 1 4 ) 1e1 2 11

6. Conclusion

1. Local heat transfer coefficients of pure R407C, in flow

boiling through an inclined tube with five inclination an-

gles from 0� to 90� in the direction of refrigerant flow (up-

ward flow) were measured at the evaporating temperature

in the range of 6 �Ce9 �C, the heat flux from 3 to 6 kW m�2

and themass velocity from 100 to 300 kgm�2s�1. The effect

of mass velocity, heat flux and tube inclinations on evap-

orative heat transfer coefficient had been investigated.

2. Heat transfer coefficient increases with the increase in

mass velocity and heat flux. Tube inclination angle affects

flow boiling heat transfer coefficient of R407C in a signifi-

cant manner. The effect of tube inclination angle is

prominent at highmass velocities for all inclinations of the

tube.

3. Heat transfer coefficient increases with increase in vapor

quality up to 34%e38% of the tube for stratified-wavy flow

region irrespective of tube inclinations at lowmass velocity

and then decreases slowly till the refrigerant exits from the

evaporator tube. The same occurs in 40%e48% of the tube

for annular flow at the highermass velocity of R407C inside

evaporator at any inclination of the tube.

4. The highest heat transfer coefficient is attained for tube

inclination of 90� in the direction of fluid flow for all mass

velocities. Pressure drop increases with the increase in

mass velocity and vapor quality before the dry-out occurs

for all inclinations of the evaporator tube. For the same

mass velocity, pressure drop increases with tube

inclination.

5. A correlation which predicts the flow boiling heat transfer

coefficients of R407C inside a plain tube inclined at

different angles within the error range of ±20% has been

proposed. The new boiling heat transfer correlation for

R407C in small diameter inclined tubes had�3.68% of MPE,

an MAE of 10.44% and an RMSE of 12.88%.

Acknowledgment

The financial support for this research by the Institute Fund is

gratefully acknowledged.

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