condensation heat transfer and pressure drop characteristics of co2 in a microchannel
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Condensation heat transfer and pressure dropcharacteristics of CO2 in a microchannel
Jaehyeok Heo a, Hanvit Park b, Rin Yun b,*aDepartment of Mechanical Engineering, Korea University, Anam-Dong, Sungbuk-Ku, Seoul 136-713, KoreabDepartment of Mechanical Engineering, Hanbat National University, Duckmyung-Dong, San 16-1,
Daejeon 305-719, Korea
a r t i c l e i n f o
Article history:
Received 19 November 2012
Received in revised form
7 April 2013
Accepted 17 May 2013
Available online 28 May 2013
Keywords:
Condensation
Heat transfer
Carbon dioxide
Heat transfer coefficient
Pressure drop
Flow complexity
Microchannel
* Corresponding author. Tel.: þ82 42 821 173E-mail address: [email protected] (R.
0140-7007/$ e see front matter ª 2013 Elsevhttp://dx.doi.org/10.1016/j.ijrefrig.2013.05.008
a b s t r a c t
The condensation heat transfer coefficient and pressure drop of CO2 in a multiport
microchannel with a hydraulic diameter of 1.5 mm was investigated with variation of the
mass flux from 400 to 1000 kgm�2s�1 and of the condensation temperature from �5 to 5 �C.
The heat transfer coefficient and pressure drop increased with the decrease of conden-
sation temperature and the increase of mass flux. However, the rate of increase of the heat
transfer coefficient was retarded by these changes. The gradient of the pressure drop with
respect to vapor quality is significant with the increase of mass flux. The existing models
for heat transfer coefficient overpredicted the experimental data, and the deviation
increased at high vapor quality and at high heat transfer coefficient. The smallest mean
deviation of �51.8% was found by the Thome et al. model. For the pressure drop, the
Mishima and Hibiki model showed mean deviation of 29.1%.
ª 2013 Elsevier Ltd and IIR. All rights reserved.
Transfert de chaleur lors de la condensation etcaracteristiques de chute de pression du CO2 a l’interieur d’unmicrocanal
Mots cles : condensation ; transfert de chaleur ; dioxyde de carbone ; coefficient de transfert de chaleur ; chute de pression ; complexite de
l’ecoulement ; microcanal
2; fax: þ82 42 821 1587.Yun).ier Ltd and IIR. All rights reserved.
Nomenclature
A area, m2
CC vena-contraction coefficient
Cp specific heat, kJkg�1K�1
c convective film constant
Dh hydraulic diameters, mm
fi interfacial roughness factor
G mass flux, kgm�2s�1
h heat transfer coefficient, Wm�2K�1
ID inner diameter, mm
j superficial velocity, ms�1
Nu Nusselt number, hDk�1
Pr Prandtl number, Cpm k�1
Re Reynolds number, GDm�1
ReLo Reynolds number with only liquid
r internal radius of tube, m
U Overall heat transfer coefficient, Wm�2 K�1
k thermal conductivity, Wm�1K�1
T temperature, �Cx vapor quality
Greek symbols
a void fraction
D difference
m viscosity, m2s�1
q upper angle of the tube not wetted by stratified
liquid, rad
r density, kgm�3
s surface tension, Nm�1
Subscripts
acc accelerational
aux auxiliary
c contractional, convective
cond condensation
e expansional
fric fricational
i internal
l liquid
lm log mean temperature difference
o external
tot total
tp two-phase
v vapor
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 3 6 ( 2 0 1 3 ) 1 6 5 7e1 6 6 81658
1. Introduction
CO2 has been utilized as a representative natural refrigerant in
various refrigeration systems, from domestic to commercial
applications. When CO2 is applied to a water heater, heat
pump, or vending machine working under ambient tempera-
ture conditions, the refrigeration cycle undergoes a transcritical
process due to its relatively low critical temperature. Recently,
CO2 refrigeration systems have been extended in application
to food storage facilities and industrial food processing, in
which the evaporation temperatures are less than �25 �C.Bansal (2012) showed that CO2 has favorable thermophysical
properties as a refrigerant at low temperature, such as rela-
tively high liquid and vapor thermal conductivities, and low
liquid viscosity and surface tension. In low-temperature ap-
plications, CO2 refrigeration systems with a condensation
process in the cycle, such as cascade systems, have beenwidely
applied. In the case of CO2/NH3 cascade system, the maximum
COP was found at condensation temperature of CO2 from �10
to 10 �C, which are dependent on the evaporation temperature
of CO2. In the present study, the condensation temperaturewas
determined from �5 to 5 �C. The cascade system has shown
benefits regarding coefficient of performance (COP) compared
to when the transcritical cycle of a multi-stage compression
system is used (Sawalha, 2008). Because most studies on CO2
systems have focused on the design of the gas cooler, the
research on CO2 condensation has been relatively limited, and
proper design for CO2 condensers is needed. Considering the
relatively low pressure drop of CO2 in a tube compared to that
of conventional refrigerants, the application of microchannels
to the CO2 condensers is promising as a compact heat
exchanger. It should be noted that microchannel condenser
showed high heat transfer coefficient too.
Many studies on the convective condensation in mini and
microchannels have been conducted, and the latest studies
with channels less than 3 mm are summarized in Table 1.
Wang and Rose (2006) theoretically analyzed the effect of
channel shape on condensation in horizontal microchannels.
The existence of a thin condensate film region around channel
determines the effect of channel shapes on the condensation
heat transfer. The heat transfer coefficient was the highest at
square channels, followed in order by rectangle, triangle, and
circle channels in the high vapor quality region. Shin and Kim
(2005) studied the condensation heat transfer inside circular
and rectangular mini-channels. In low mass flux conditions,
rectangular channels showed slightly higher heat transfer co-
efficients than circular channels due to the effect of the gutter
flow of the liquid at the corners induced by surface tension in
the rectangular channels. As the mass flux increased, the heat
transfer coefficients of the circular channels were higher than
those of the rectangular channels. Zhang et al. (2012) and Del
Col et al. (2010) investigated condensation heat transfer with
alternative refrigerants in single circular tubes with inner di-
ameters between 0.96 and 1.289mm. Zhang et al. (2012) utilized
R22, R410A, and R407C as working fluids. They determined that
the interface shear stresswas themost dominant factor at high
vapor quality, and conduction in the liquid layer was dominant
at low vapor quality in condensation heat transfer. Existing
correlations showed substantial discrepancies with experi-
mental data. Del Col et al. (2010) compared the heat transfer
coefficient of R1234yf with that of R134a. They showed that the
heat transfer coefficient of R1234yf was lower than that of
R134a by 15e30% in the vapor quality region ranging from0.4 to
0.8. This was explained by the high thermal resistance of
R1234yf by relatively low thermal conductivity. The prediction
of the heat transfer coefficient by the model of Cavallini et al.
Table 1 e Existing studies on condensation in horizontal microchannels.
References Fluids Geometry of test tubes Test conditions Measurements/calculations
Haui and Koyama
(2004)
CO2 Circular channels (Dh: 1.31 mm, 10 multiport) G: 123.2e315.2 kgm�2s�1
Tcond: 21.63e31.33 �CHeat transfer coefficients
Park and Hrnjak
(2009)
CO2 Circular channels (Dh: 0.89 mm, 10 multiport) G: 200e800 kgm�2s�1
Tcond: �15, �25 �CHeat transfer coefficients,
pressure drop
Zhang et al. (2012) R22, R410A,
R407C
A single circular tube (Dh: 1.088, 1.289 mm) G: 300e600 kgm�2s�1
Tcond: 30, 40 �CHeat transfer coefficients,
pressure drop
Del Col et al. (2010) R1234yf,
R134a
A single circular tube (Dh: 0.96 mm) G: 200e1000 kgm�2s�1
Tcond: 40 �CHeat transfer coefficients,
pressure drop
Wang and Rose
(2006)
R134a Single square/triangle/inverted triangle/rectangle
channel (Dh: 0.577e1.2 mm)
G: 500 kgm�2s�1
Tcond: 50 �CHeat transfer coefficients
Shin and Kim
(2005)
R134a Single circular and rectangular tubes
(Dh: 0.494e1.067 mm)
G: 100e600 kgm�2s�1
Tcond: 40 �CHeat transfer coefficients
Cavallini et al.
(2005)
R134a,
R410A
Rectangular channels (Dh: 1.4 mm, 13 multiport) G: 200e1400 kgme2s�1
Tcond: 40 �CHeat transfer coefficients,
pressure drop
Koyama et al.
(2003)
R134a Rectangular channels (Dh: 1.11 mm, 8 multiport),
(Dh: 0.80 mm, 19 multiport)
G: 100e700 kgm�2s�1
Pcond: 1.7 MPa
Heat transfer coefficients,
pressure drop
Wang et al. (2002) R134a Rectangular channels (Dh: 1.46 mm, 10 multiport) G: 75e750 kgm�2s�1
Tcond: 61e66.5 �CHeat transfer coefficients
Garimella et al.
(2005)
R134a Circular channels (Dh: 0.5e4.91 mm, 1, 10, 17,
23 multiport)
G: 150e750 kgm�2s�1 Pressure drop
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 3 6 ( 2 0 1 3 ) 1 6 5 7e1 6 6 8 1659
(2006) was well matched within �15%. Cavallini et al. (2005),
Koyama et al. (2003), and Wang et al. (2002) studied the
condensation heat transfer of R134a in multiport rectangular
channels with hydraulic diameters between 0.8 and 1.5 mm.
The condensation heat transfer coefficientwasunderestimated
by the existing models developed for macroscale and mini-
channels in the study by Cavallini et al. (2005). Prediction by the
models of Cavallini et al. (2002) and Moser et al. (1998) had
relatively small deviation between þ20% and �30%. Koyama
et al. (2003) compared their experimental data with the model
of Moser et al. (1998), and the deviations ranged from þ300
to �20%. Wang et al. (2002) observed the transition between
flowpatterns of slug,wavy, and annular flow. They developed a
heat transfer model based on the observed flow patterns. Haui
and Koyama (2004) and Park and Hrnjak (2009) experimentally
studied the condensation of CO2 in circular multi-channels.
The pressure drop in Haui and Koyama’s study was estimated
within a deviation of �65% by the Koyama et al. (2002) model.
Park and Hrnjak (2009) showed that the Thome et al. (2003)
model and the Akers et al. (1959) model estimated their
experimental results with 18.3% and 17.5% absolute average
deviations in the heat transfer coefficient, respectively. The
absolute average deviation of the pressure drop estimated by
the models of McAdams et al. (1942), Friedel (1979), and
Mishima and Hibiki (1996) were 13.0, 49.4, and 21.9%,
respectively.
As summarized in the literature review, studies on the flow
condensation of CO2 inmicrochannels have been very limited,
especially with regard to experimental study in multiport
rectangular microchannels. The objective of this study is to
investigate the condensation heat transfer and pressure drop
characteristics of CO2 in a multiport rectangular micro-
channel. The effects of mass flux and condensation temper-
ature on the heat transfer and the pressure drop of CO2 are
explained by the thermophysical properties of CO2 and by the
two-phase flow conditions. The existing prediction models of
the heat transfer coefficient and pressure drop are compared
with present experimental data for utilization in the design of
CO2 condensers.
2. Experiments
2.1. Test facilities and experiment methods
Fig. 1 shows the schematic of the test setup. It was composed
of a magnetic gear pump, preheater, test section, and sub-
cooler. The working fluid was pure CO2, and the secondary
fluid for the preheater and the test section was a mixture of
Ethylene Glycol (EG) and water. The magnetic gear pump,
which can work without oil, circulates the CO2 through the
test section. The preheater was utilized to adjust the inlet
vapor quality of the test section, and a plate heat exchanger
was used as the preheater. Two plate heat exchangers were
utilized for the subcooler, which liquefied the CO2 from the
outlet of the test section. The constant temperature bath was
connected to the preheater, to provide heat to the liquefied
CO2. The chiller, which was connected to the test section,
removes heat from the CO2 to condense it throughout the
test section. Other chillers shown in Fig. 1 were used to
subcool the CO2, and to control the system pressure after
finishing the test for safety. The mass flow rate of the CO2
was controlled by changing the gear speed of the magnetic
gear pump, and the condensation temperature was adjusted
by controlling both temperatures of the chillers, which were
connected to the subcooler, and the charge amount of the
CO2 in the system. The inlet vapor quality of the test section
was set by changing the heat input to the preheater. The
mass flow rate was measured by a Coriolis type mass flow
meter, and the thermodynamic status of the subcooled CO2
at the preheater inlet was calculated from the measured
temperature and pressure. The temperature was measured
T
Heat
exchanger
Relief valve
Bypass valve
Chiller
ChillerChiller
Gear pump
Differential
Pressure transducer
Water
bath
Sight glass
CO2 tank
Test section
P
Receiver
tankChiller
T
T
T
T
Heat
exchanger
Mass flow meter
P
PP
Flow meter
T P
Flow
meter
Heat
exchanger
T T
T
Pressure
Temperature
Needle valve
P
Fig. 1 e Schematics of experimental setup.
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 3 6 ( 2 0 1 3 ) 1 6 5 7e1 6 6 81660
by the probe-type thermocouple. The condensation temper-
ature was calculated by measuring the saturation pressure,
and the pressure drop across the test tube was measured by
the differential pressure transducer. The mass flow rate of
the EG and water mixture was calculated by measuring the
volume flow rate and the temperatures at the preheater and
the test section. All thermocouples utilized in the present
18
1
2
1.2
TP
CO outlet Brine inlet
40
1.8
Brine inlet
T
450
18
10 (ID=8)
Front vi
Side vie
Top vie
Microchannel
Microcha
9.52
DP
Auxiliary t
Fig. 2 e Details of
tests were calibrated simultaneously by using the constant
temperature bath, and the differences among the thermo-
couples were less than 0.1 �C. The mass flow meter and the
volumetric flow meters were calibrated by measuring the
total weight of the fluid within the specific time interval by
changing the flow rate. Water and the EG and the water
mixture were utilized for the calibration of the mass flow
.8 4
CO inlet
T P
Brine outlet
40
4
Brine outletT
20
10 (ID=8)
100
ew
w
w
nnel
20
10
9.52
ubes
test section.
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 3 6 ( 2 0 1 3 ) 1 6 5 7e1 6 6 8 1661
meter and the volumetric flow meter, respectively. The cali-
bration data for the pressure transducers were provided by
the manufacturer, and the differential pressure transducer
was calibrated by using the commercial pressure calibrator.
Fig. 2 shows the details of the test section. An aluminum
microchannel with seven rectangular channels with a hy-
draulic diameter of 1.5 mm was utilized. The length, width,
and height of the microchannel were 450 mm, 18 mm and
1.8 mm, respectively. Both sides of the microchannel were
inserted into horizontally located headers with an outside
diameter of 10 mm. The header was also made of aluminum,
machined in the longitudinal direction to be fitted to the
microchannel, and blazed to the microchannel. Two inlets
were made for even distribution of CO2 into the micro-
channel by maintaining similarities between them. The
channel for the EG and water mixture was formed by bonding
two acrylic blocks with rectangular shape of the flow chan-
nel. Two 9.52 mm circular holes for the inlet and outlet of the
secondary fluid were machined in the acrylic block. The flow
direction between the CO2 and the secondary fluid was
counter flow. Every part of the test setup, including the test
section, was heavily insulated with the insulator, the thermal
conductivity of which was less than 0.04 Wm�1K�1. After
checking no change in temperature and pressure for 10 min
after reaching steady-state, data were collected for 90 s. Data
from the data logger was averaged over the collected time,
and they were modified to the Engineering Equation Solver
(EES) input format. The equations were programmed to
determine the heat transfer coefficient. All needed thermo-
physical properties of CO2 at saturation or superheat status,
and the EG and water mixture at subcooled status were
determined by the equation of state included in EES.
2.2. Data reduction
The condensation heat transfer coefficient hi was calculated
using Eq. (1). The thermal resistance of the microchannel wall
was neglected in Eq. (1), which was estimated by less than
0.2% compared to that of fluids. The UA value in Eq. (1) was
obtained by using the heat transfer rate to CO2 from brine, and
2.6 2.8 3.0-100
-80
-60
-40
-20
0
20
40
60
80
100
Mea
n de
viat
ion
(%)
Heat transfer coefficient (kW/m2K)
Fig. 3 e Deviations between measured and predicted
single-phase heat transfer coefficient of CO2.
the temperatures of the inlet and outlet CO2 and brine, as
shown in Eq. (2). The brine side heat transfer coefficient was
determined by utilizing the Wilson plot method. Eq. (3) shows
the brine side heat transfer coefficient ho. The thermal con-
ductivity of 0.3691 Wm�1K�1 brine was utilized for Eq. (3). The
single-phase heat transfer coefficient for CO2 deviates from
the Gnielinski model with an average mean deviation of 24%
with the variation of the secondary fluid’s mass flow rate as
shown in Fig. 3. The average percentage of thermal resistance
of the annulus-side over the total thermal resistance was
15.3%, and the average uncertainty of the annulus-side heat
transfer coefficient was �3.9%. The energy balance in the test
section between the secondary fluid and CO2 was 5.83%. The
inlet vapor quality of the test section was calculated by Eq. (4).
The enthalpy at the test section inlet can be obtained by using
the thermodynamic status of CO2 at the preheater inlet, and
the heat input to the preheater. The saturation properties in
Eq. (4) were calculated based on the measured saturation
pressure. Eq. (5) was utilized to get the outlet vapor quality of
the test section. The average change of quality across the
subsection was 0.031. Auxiliary tubes were used as inlet and
outlet ports, and were connected to the microchannel as
shown in Fig. 2. The net frictional pressure drop across the
microchannel needed to be recalculated by subtracting the
pressure drop in the auxiliary tubes from the measured
pressure drop. Besides, the accelerational, sudden expan-
sional and contractional pressure drop from header to each
port ofmicrochannel should be considered. The pressure drop
in the auxiliary tubes was estimated by the Cavallini et al.
(2002) models, which showed the best predictability in a
smooth tube with CO2 (Kang et al., 2012). The net frictional
pressure drop was calculated by using Eq. (6). The accelera-
tional pressure drop was predicted by using the Eq. (7). The
pressure drop from the sudden contraction at inlet and from
the sudden expansion at outlet was estimated by using Eqs. (8)
and (9), respectively (Abdelall et al., 2005).
1
UA¼ 1
hoAoþ 1
hiAi(1)
Q ¼ UADTlm (2)
Nuo ¼ 0:0971Re0:5Pr0:3 (3)
xtest;inlet ¼�itest;inlet � il
��ivl (4)
xtest;outlet ¼ xtest;inlet � _Qtest=�_mCO2
� ivl�
(5)
DPtot ¼ DPfric þ DPace þ DPc � DPe þ DPaux (6)
Table 2 e Test conditions.
Test conditions Ranges
Mass flux 400 to 1000 kgm�2s�1
Condensation temperature �5 to 5 �CVapor quality 0.0 to 1.0
Table 3 e Uncertainties in variables.
Variables Uncertainties Range
Fluid temperature 0.1 �C �250e350 �CPressure of CO2 0.13% of the
full scale
�14.7e1000 psia
Volume flow rate of
cooling fluid
1.0% of the
full range
1.14e11.36 lpm
Mass flow rate of CO2 0.2% of
measurement
0e0.3 kgs�1
Differential pressure drop
of CO2
�4.06%
Condensation heat transfer
coefficient of CO2
�9.26%
Vapor quality of CO2 �6.1%
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 3 6 ( 2 0 1 3 ) 1 6 5 7e1 6 6 81662
DPacc ¼ G2
("x2o
rvaoþ ð1� xoÞ2rlð1� aoÞ
#�"x2i
rvaiþ ð1� xiÞ2rlð1� aiÞ
#)(7)
DPe ¼ G2s
�s� 1
��1� x2
rlð1� aÞ þx2
rva
�(8)
DPc ¼ G2
2rl
�1Cc
� 1
�2
þ �1� s2�ð1þ xðrl � rvÞ=rlÞ (9)
Table 2 shows the present test conditions. Themass fluxwas
varied to 400, 600, 800, and 1000 kgm�2s�1, and the condensa-
tion temperature changed from �5 to 5 �C. Because the pres-
sure drop of CO2 is much smaller compared to that of the
conventional refrigerants, the present mass flux is considered
as the operation condition of the CO2 condenser. Besides, the
present mass flux condition provided the different ranges of
mass flux from the previous studies. The vapor quality ranged
from 0.0 to 1.0. Table 3 shows the measurement variables, the
measuring ranges, and the uncertainties of the measurement.
It also provides the uncertainty propagations of pressure drop,
heat transfer coefficient, and inlet vapor quality of the test
section. These were calculated by the method of which
0.2 0.4 0.6 0.8 1.00
1
2
3
4
5
6
7
0.0 0.2 0.4 0.6 0.8 1.00
1
2
3
4
5
6
7Massflux: 400 kgm s
Tcond (oC) -5 0 5
mWk(tneiciffeocrefsnarttae
H-2K-1
)
Vapor quality
Tcond (oC) -5 0 5
Massflux: 600 kgm s
Vapor quality
Fig. 4 e Variation of heat transfer coefficient with condensa
is described in the NIST Technical Note (Taylor and Kuyatt,
1994). As shown in Table 3, the average uncertainties of the
pressure drop, the heat transfer coefficient, and the inlet vapor
quality of the test section were �4.06%, �9.26% and �6.1%,
respectively.
3. Results and discussion
3.1. Heat transfer coefficient
Fig. 4 shows heat transfer coefficient with variation of the
condensation temperature with mass fluxes of 400, 600, 800,
and 1000 kgm�2s�1. It was found that the heat transfer coef-
ficient increased with the decrease of condensation temper-
ature for all mass flux conditions. For example, when the
mass flux was 600 kgm�2s�1, the heat transfer coefficients at
�5 �C and 0 �C were larger than that at 5 �C by 29.2 and 25.2%,
respectively. For the mass flux of 800 kgm�2s�1, the heat
transfer coefficients at �5 �C and 0 �C were larger than that at
5 �C, by 29.6 and 26.1%, respectively. As the condensation
temperature decreased, the liquid film on the tube wall
became thinner due to the variation of the density ratio be-
tween the liquid and vapor, and its thermal resistance
decreased. As noted, the condensation heat transfer coeffi-
cient was enhanced with decrease in the liquid film thickness,
which acts as a thermal resistance in the annular flow con-
dition. Fig. 4 also shows that the heat transfer coefficient
dropped at a certain vapor quality at a mass flux of
1000 kgm�2s�1. This phenomenon can be explained by the
flow complexity. As explained in Chen et al. (1987), flow
complexity can be specified as the two-phase flow status,
which is different from the annular flow having smooth sur-
face and continuous liquid film. The flow changed from
annular to mist-annular or mist flow by the high vapor shear.
Flow complexity comes from the high liquid entrainment at
high vapor quality, and significantly deteriorates the
condensation heat transfer coefficient. This decreasing trend
of heat transfer coefficient at high vapor quality was well
0.0 0.2 0.4 0.6 0.8 1.00
1
2
3
4
5
6
7
0.0 0.2 0.4 0.6 0.8 1.00
1
2
3
4
5
6
7
Tcond (oC) -5 0 5
Massflux: 800 kgm s
Vapor quality
Tcond (oC) -5 0 5
Massflux: 1000 kgm s
Vapor quality
tion temperature under different mass flux conditions.
0.2 0.4 0.6 0.8 1.00
1
2
3
4
5
6
7
0.0 0.2 0.4 0.6 0.8 1.00
1
2
3
4
5
6
7
0.0 0.2 0.4 0.6 0.8 1.00
1
2
3
4
5
6
7
Massflux 400 kgm-2s-1
600 kgm-2s-1
800 kgm-2s-1
1000 kgm-2s-1
Tcond : -5oC
Heat tran
sfer co
efficien
t
(kW
m
-2
K-1
)
Vapor quality
Massflux 400 kgm-2s-1
600 kgm-2s-1
800 kgm-2s-1
1000 kgm-2s-1
Tcond : 0oC
Vapor quality
Massflux 400 kgm-2s-1
600 kgm-2s-1
800 kgm-2s-1
1000 kgm-2s-1
Tcond : 5oC
Vapor quality
Fig. 5 e Variation of heat transfer coefficient with mass flux under the different condensation temperature.
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 3 6 ( 2 0 1 3 ) 1 6 5 7e1 6 6 8 1663
reflected in the existingmodels, such as the Chen et al. model,
the Soliman et al. model, and the Traviss et al. model (Carey,
2008). It was found that the maldistribution of CO2 from
header to each port of microchannel increased with the in-
crease of mass flux (Lu et al., 2004). Maldistribution is the non-
even distribution of liquid and vapor phase amount from
header to each port of microchannel. The flow pattern of
annular flow at the header can be easily changed to mist and
annular flow having a non-continuous condensation film,
which worsened the heat transfer coefficient. The quantita-
tive effect of flow maldistribution on the heat transfer can be
estimated by the Bielskus (2011). The cooling capacity of the
heat exchangerwith uniform distributionwas increased by an
average of 34% than that of the heat exchanger with maldis-
tribution. The detail quantitative analyses of the maldistri-
bution were provided in the section 3.3.
The data in Fig. 4 was rearranged with variation of the
mass flux under different condensation temperatures to
0.2 0.4 0.6 0.8 1.00
20
40
60
80
100
120
140
160
180
200
0.0 0.2 0.4 0.6 0.8 1.00
20
40
60
80
100
120
140
160
180
200
Massflux 400: kgm s
Tcond (oC) -5 0 5
Pre
ss
ure
d
ro
p (k
Pa
m-1
)
Vapor quality
Tcond (oC) -5 0 5
Massflux: 600 kgm s
Vapor quality
Fig. 6 e Variation of pressure drop with the condensation
observe the effect of mass flux on the heat transfer coefficient,
as shown in Fig. 5. The condensation heat transfer coefficient
increased with the increase of mass flux for all condensation
temperatures. When the condensation temperature was 0 �C,the increasing rates of heat transfer coefficient with the in-
crease of mass flux from 400 to 600, 800 and 1000 kgm�2s�1
were 17.0%, 21.7%, and 40.4%, respectively. It was also found
that the increasing the rate of the heat transfer coefficients
began to slope downwards with the increase of mass flux,
especially at �5 �C. As noted, flow transition can held back the
estimated increase of heat transfer coefficient with the in-
crease of mass flux and the decrease of condensation
temperature.
3.2. Pressure drop
Fig. 6 shows the variation of the pressure drop with conden-
sation temperature under different mass flux conditions.
0.0 0.2 0.4 0.6 0.8 1.00
20
40
60
80
100
120
140
160
180
200
0.0 0.2 0.4 0.6 0.8 1.00
20
40
60
80
100
120
140
160
180
200
Tcond (oC) -5 0 5
Massflux: 800 kgm s
Vapor quality
Tcond (oC) -5 0 5
Massflux: 1000 kgm s
Vapor quality
temperature under the different mass flux condition.
0.0 0.2 0.4 0.6 0.8 1.00
1
2
3
4
5
6
7
8
9
10
11
12
13
0.0 0.2 0.4 0.6 0.8 1.00
1
2
3
4
5
6
7
8
9
10
11
12
13
0.0 0.2 0.4 0.6 0.8 1.00
1
2
3
4
5
6
7
8
9
10
11
12
13 Data 5 C Data -5 C Data -15 C (Park and Hrnjak, 2009)
G: 400 kgm sHe
at tra
ns
fe
r c
oe
ffic
ie
nt
(k
Wm
-2
K-1
)
Vapor quality
G: 600 kgm s
Data 5 C Data -5 C Data -15 C (Park and Hrnjak, 2009)
Vapor quality
G: 800 kgm s
Data 5 C Data -5 C Data -15 C (Park and Hrnjak, 2009)
Vapor quality
Fig. 7 e Comparison of the present heat transfer coefficient with the existing data.
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 3 6 ( 2 0 1 3 ) 1 6 5 7e1 6 6 81664
The pressure drop increased with the decrease of the conden-
sation temperature. For instance, when the mass flux was
400 kgm�2s�1, the pressure drops at �5 �C and 0 �C were larger
than that at 5 �C by 115.4% and 43.1%, respectively. For a mass
flux of 600 kgm�2s�1, the pressure drops at�5 �C and 0 �C were
larger than that at 5 �C by 84.6 and 51.4%, respectively. The
effects of mass flux on the pressure drop were more apparent
compared to the effects of condensation temperature on the
pressure drop. The average pressure drops per unit length of
the tube were 35.3 kPam�1, 50.2 kPam�1, 70.3 kPam�1, and
91.3 kPam�1 at 400, 600, 800, and 1000 kgm�2s�1, respectively. It
is evident that the vapor velocity and liquid viscosity increased
with the decrease of condensation temperature. Increasing
vapor velocity increased interface shear stress between liquid
and vapor under annular flow conditions. It was observed that
the gradient of pressure dropwith vapor quality increased with
0 5 10 15 20
0
5
10
15
20
+50%
+100%
+200%
Bandhauer et al. (2005) Cavallini et al. (2002) Thome et al. (2003)
Pred
icted
h
(kW
m K
)
-2
-1
Measured h (kWm K )-2 -1
0%
Fig. 8 e Comparison of the measured and predicted heat
transfer coefficients.
the increase of mass flux. As noted, the flow complexity is
dominant with the increase of mass flux, which significantly
increased the pressure drop. It was also found that the effects
of the condensation temperature on the pressure drop gradu-
ally decreased with the increase of mass flux.
3.3. Discussion
The present heat transfer coefficient was compared with the
studies of Park and Hrnjak (2009) as shown in Fig. 7. The data
under the same mass flux was compared, and the experi-
mental conditions of Park and Hrnjak (2009) are summarized
in Table 1. The present results showed a similar trend to those
found by Park and Hrnjak in terms of the effects of conden-
sation temperature and mass flux on the heat transfer coef-
ficient. However, the data of Park andHrnjak (2009) was higher
than that of ours by 17e89%with the increase of vapor quality.
Generally, the condensation heat transfer coefficient
increased with the decrease of hydraulic diameter. The mal-
distribution is delineated with smaller hydraulic diameter.
The comparatively larger flow resistance at header to micro-
channel with smaller hydraulic diameter forced the liquid
flows toward other channels, thereby causing a more uni-
formed flow distribution. Higher heat transfer coefficient and
the constantly increasing trend of the heat transfer coefficient
at high vapor quality were expected with Park’s and Hanjak’s
experiment.
Fig. 8 shows a comparison of the measured heat transfer
coefficients and those predicted by the Thome et al. model
(2003), the Cavallini et al. model (2002), and the Bandhauer
et al. model (2005). Table 4 summarized the application
range of eachmodel. All themodels overpredicted the present
experimental data, and the deviation linearly increased with
the increase of heat transfer coefficient. Among the existing
models, the smallestmean deviationwas found by the Thome
et al. (2003) model, which was �51.8%. Even the Bandhauer
et al. model, which was developed for microchannel
condensation heat transfer coefficients, showed high
0.0 0.2 0.4 0.6 0.8 1.00
2
4
6
8
10
12
0.0 0.2 0.4 0.6 0.8 1.00
2
4
6
8
10
12
0.0 0.2 0.4 0.6 0.8 1.00
2
4
6
8
10
12
0.0 0.2 0.4 0.6 0.8 1.00
2
4
6
8
10
12
0.0 0.2 0.4 0.6 0.8 1.00
2
4
6
8
10
12
0.0 0.2 0.4 0.6 0.8 1.00
2
4
6
8
10
12
0.0 0.2 0.4 0.6 0.8 1.00
2
4
6
8
10
12
0.0 0.2 0.4 0.6 0.8 1.00
2
4
6
8
10
12
0.0 0.2 0.4 0.6 0.8 1.00
2
4
6
8
10
12
0.0 0.2 0.4 0.6 0.8 1.00
2
4
6
8
10
12
0.0 0.2 0.4 0.6 0.8 1.00
2
4
6
8
10
12
0.0 0.2 0.4 0.6 0.8 1.00
2
4
6
8
10
12
h (k
Wm
-2
K-1
)
Thome et al. Measured data
T : -5 C
G: 400 kgm s
Thome et al. Measured data
T : -5 C
G: 600 kgm s
h (kW
m-2
K-1
)
Thome et al. Measured data
T : -5 C
G: 800 kgm s
h (k
Wm
-2
K-1
)
Thome et al. Measured data
T : -5 C
G: 1000 kgm s
h (kW
m-2
K-1
)
Vapor quality
Thome et al. Measured data
T : 0 C
G: 400 kgm s
Thome et al. Measured data
T : 0 C
G: 600 kgm s
Thome et al. Measured data
T : 5 C
G: 600 kgm s
Thome et al. Measured data
T : 0 C
G: 800 kgm s
Thome et al. Measured data
T : 5 C
G: 800 kgm s
Thome et al. Measured data
T : 0 C
G: 1000 kgm s
Vapor quality
Thome et al. Measured data
T : 5 C
G: 1000 kgm s
Vapor quality
Thome et al. Measured data
T : 5 C
G: 400 kgm s
Fig. 9 e Comparison of the measured data and predicted heat transfer coefficients by the Thome et al. model.
Table 4 e Ranges of applicability of the existing models.
References Models Fluids Geometry of test tubes Applicable range Flow regimes
Thome et al. (2003) Heat transfer R11, R12, R22, R32, R113,
R125, R134a, R236ea, R404A,
R410A, Propane, n-butane,
Iso-butane, Propylene
Horizontal plain tubes
(Di: 3.1e21.4 mm)
G: 24e1022 kgm�2s�1 Annular, intermittent,
stratified-wavy, fully
stratified, mist flow
Cavallini et al.
(2002)
Heat transfer R22, R134a, R125, R32,
R236ea, R407C, R410A
Plain tube (Di: 8 mm) G: 100e750 kgm�2s�1 Annular, stratified,
slug
Bandhauer et al.
(2005)
Heat transfer R134a Horizontal microchannels
(Dh: 0.51e1.52 mm)
G: 150e750 kgm�2s�1 Annular, mist, and
disperse wave
Garimella et al.
(2005)
Pressure drop R134a Horizontal microchannels
(Dh: 0.50e4.91 mm)
G: 150e750 kgm�2s�1 Annular, disperse
wave, mist, discrete
wave, intermittent
Lee and Lee (2001) Pressure drop Water-air Horizontal microchannels
(Dh: 0.78e6.67 mm)
ReLo: 175e17,700
X: 0.303e79.4
Laminar, Turbulent
Mishima and Hibiki
(1996)
Pressure drop Water-air Vertical upward round
tube (Di: 1e4 mm)
jv : 0.0896e79.3 ms�1
jl: 0.0116e1.67 ms�1
e
Friedel (1979) Pressure drop e (Dh > 1 mm) ml mv�1 < 1000 Annular
McAdams et al.
(1942)
Pressure drop e e Similar vapor and
liquid velocity
Bubbly, wispy-annular
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 3 6 ( 2 0 1 3 ) 1 6 5 7e1 6 6 8 1665
0 20 40 60 80 100 120 140 160 180 200
0
20
40
60
80
100
120
140
160
180
200
+50%
-50%
+200%
Pre
dic
te
d p
re
ss
ure
d
ro
p
(k
Pa
m-1
)
Measured pressure drop (kPam-1
)
0%
Fridel (1979) McAdams et al. (1942)
Garimella et al. (2005) Lee and Lee (2001) Mishima and Hibiki (1996)
Fig. 10 e Comparison of pressure drop between the present
and the existing study.
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 3 6 ( 2 0 1 3 ) 1 6 5 7e1 6 6 81666
overprediction of the present data. Fig. 9 showed the detail
prediction results by the Thome et al. model. The deviation
increased with increase of mass flux and at high vapor quality
region, which was found to be similar for all comparisons by
using other existing models. As explained, two phase flow
patterns can be easily changed from the annular flow to the
mist-annular or mist flow with increase mass flux at high
vapor quality, which significantly decreased the condensation
heat transfer coefficient. The transition of the flowpattern can
be advanced to relatively lower vapor quality for the micro-
channel. Besides, the application ranges of mass flux of other
models are limited to 750 kgm�2s�1 as shown in Table 4 except
the Thome et al. model. Above reasons can explain the low
predictability of the existing models. In Park and Hrnjak’s
study (2009), the over-prediction of the existing model was
also observed at a high coefficient range, and modifications of
flow pattern were suggested as a possible reason.
The effects of the two-phase flow maldistribution in
microchannel on the condensation heat transfer coefficient
of CO2 were quantitatively analyzed by simulating flow pat-
terns in each port of microchannel. According to Ahmad
et al. (2009)’s study on the two-phase distribution, liquid is
dominant in the channels near the header, and vapor portion
increased farther from the header. Accordingly, the flow dis-
tributions at each port of the microchannel were divided into
liquid dominant, balanced, and vapor dominant regions. Each
region of the liquid dominant, the balanced, and the vapor
dominant was matched to the flow patterns of the stratified
flow, the annular flow, and the mist flow. The flow distribu-
tions to each port were categorized from case 1 to 5 as sum-
marized in Table 5. Case 1 simulated the uniform distribution,
and this balanced region was being diminished from case 2 to
case 5. As noted, the flow pattern at ports near header was
simulated as the stratified flowwith liquid dominant, and that
at ports near center was regarded as the mist flow with vapor
dominant. The Thome et al. model (2003) was used for esti-
mation of the heat transfer coefficient for the annular flow
and the stratified flow as shown in Eqs. (10) and (11), respec-
tively. The modified DittuseBoelter equation of Eq. (12) was
applied to the calculation of the heat transfer coefficient in
mist flow (Carey, 2008). As shown in Table 5, significant
decrease of themean deviation between the predicted and the
experimental results was found by considering the flow mal-
distribution to each port. The results of cases 3 and 4 were the
closest to the measured results, and the measured data in the
high condensing temperature and the high mass flux condi-
tions showed similar results with cases 4 and 5. The average
Table 5 e Simulated maldistribution inside microchannel andexperiments and predictions.
Maldistribution assumptions Ch 1 Ch 2 Ch
Case 1 BR: 7 A A A
Case 2 LDR: 2, BR: 5 S A A
Case 3 LDR: 2, BR: 4, VDR: 1 S A A
Case 4 LDR: 2, BR: 2, VDR: 3 S A M
Case 5 LDR: 4, BR: 2, VDR: 1 S S A
(BR: Balanced region, LDR: Liquid dominant region, VDR: Vapor dominan
heat transfer coefficient under the flow condition in case 5
decreased by 44.2% when compared with that of the flow
condition in case 1. Based on the present simulation results,
the condensation heat transfer coefficient in the micro-
channel was significantly affected by the non-even distribu-
tion of the liquid and the vapor phase, which determined the
two phase flow pattern at each port. The model considering
the possible flow patterns at each port was proved to provide
much precise prediction results with the experimental data.
hannular ¼ cRenLPr
mL
lL
dfi (10)
hstratified ¼ hf rqþ ð2p� qÞrhc
2pr(11)
hmist ¼ 0:023ktp
D
GDmtp
!0:8
Pr0:4tp (12)
Fig. 10 shows the comparison of pressure drop between
the present and previous studies. Among the models for a
macro-scale tube, the Friedel (1979) model and the McAdams
et al. (1942) model were validated. Among the pressure drop
models for tubes or channels with a small diameter, the
Garimella et al. (2005) model, the Lee and Lee model (2001),
and the Mishima and Hibiki model (1996) were compared with
mean deviation of heat transfer coefficient between
3 Ch 4 Ch 5 Ch 6 Ch 7 Mean deviation (%)
A A A A 51.84
A A A S 26.75
M A A S 18.56
M M A S 15.58
M A S S 20.25
t region, A: Annular, S: Stratified, M: Mist).
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 3 6 ( 2 0 1 3 ) 1 6 5 7e1 6 6 8 1667
the tested data. The smallest deviation was found by the
Mishima and Hibiki (1996) model. The mean deviations of the
Mishima and Hibiki model (1996), the Garimella et al. (2005),
the Lee and Leemodel (2001), theMcAdams et al. model (1942),
and the Friedel (1979) model were 29.1, 36.6, 47.0, 64.4, and
74.3%, respectively. Park and Hrnjak (2009) suggested a ho-
mogenous flow model instead of a separated flow model for
the prediction of pressure drop of CO2 in microchannels by
considering the relatively low velocity difference between
vapor and liquid in a small-diameter tube. The velocity dif-
ference between the vapor and liquid could become smaller
with the decrease of tube diameter, which resulted in better
predictions of pressure by a homogenous flowmodel than by a
separated flowmodel. In this study, the McAdams et al. (1942)
model, which was developed based on a homogenous flow
model, showed better prediction than the Friedel (1979)
model, which was developed based on a separated flow
model. However, the estimation by recent pressure drop
models developed for tubes or channels with a small diameter
were bettermatchedwith the experimental data than those of
classical models. Considering the values of mean deviation,
the models of Mishima and Hibiki (1996) and Garimella
et al. (2005) are recommendable to predict the condensation
pressure drop in a microchannel for CO2.
4. Conclusions
The condensation heat transfer coefficient and pressure dropof
CO2 in multiport rectangular microchannels were experimen-
tally investigated with variation of the mass flux and the
condensation temperature from 400 to 1000 kgm�2s�1 and from
�5 to 5 �C, respectively. The effect of condensation temperature
and themass flux on the heat transfer coefficients were similar
to those of existing studies. However, the effects of flow
complexity and flow pattern transition on heat transfer and
pressure drop were dominant at high mass flux, at low
condensation temperature, and at high vapor quality. The
existing models overpredicted the present experimental data
by 0%e200%, and the deviation was significant at high heat
transfer coefficient. Both the thermophysical properties of CO2,
which are different from conventional refrigerants, and the
effect on the thickness and shape of the liquid film and flow
complexity from the microchannel, can explain the low pre-
dictability of the existingmodels. Several pressure dropmodels
for macro- and microscale tubes and channels were compared
with the measured data, and the mean deviations of the
models by Mishima and Hibiki (1996) and Garimella et al. (2005)
were 29.1 and 36.6%, respectively.
Acknowledgment
This research was supported by the Basic Science Research
Program through the National Research Foundation of Korea
(NRF) funded by the Ministry of Education, Science and
Technology (2011-0025728).
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