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Condensation heat transfer and pressure drop characteristics of CO2 in a microchannel
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 7 e1 6 6 8
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Condensation heat transfer and pressure drop characteristics of CO2 in a microchannel Jaehyeok Heo a, Hanvit Park b, Rin Yun b,* a
Department of Mechanical Engineering, Korea University, Anam-Dong, Sungbuk-Ku, Seoul 136-713, Korea Department of Mechanical Engineering, Hanbat National University, Duckmyung-Dong, San 16-1, Daejeon 305-719, Korea b
article info
abstract
Article history:
The condensation heat transfer coefficient and pressure drop of CO2 in a multiport
Received 19 November 2012
microchannel with a hydraulic diameter of 1.5 mm was investigated with variation of the
Received in revised form
mass flux from 400 to 1000 kgm2s1 and of the condensation temperature from 5 to 5 C.
7 April 2013
The heat transfer coefficient and pressure drop increased with the decrease of conden-
Accepted 17 May 2013
sation temperature and the increase of mass flux. However, the rate of increase of the heat
Available online 28 May 2013
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
Keywords:
for heat transfer coefficient overpredicted the experimental data, and the deviation
Condensation
increased at high vapor quality and at high heat transfer coefficient. The smallest mean
Heat transfer
deviation of 51.8% was found by the Thome et al. model. For the pressure drop, the
Carbon dioxide
Mishima and Hibiki model showed mean deviation of 29.1%.
Heat transfer coefficient
ª 2013 Elsevier Ltd and IIR. All rights reserved.
Pressure drop Flow complexity Microchannel
Transfert de chaleur lors de la condensation et caracte´ristiques de chute de pression du CO2 a` l’inte´rieur d’un microcanal Mots cle´s : condensation ; transfert de chaleur ; dioxyde de carbone ; coefficient de transfert de chaleur ; chute de pression ; complexite´ de l’e´coulement ; microcanal
* Corresponding author. Tel.: þ82 42 821 1732; fax: þ82 42 821 1587. E-mail address: [email protected] (R. Yun). 0140-7007/$ e see front matter ª 2013 Elsevier Ltd and IIR. All rights reserved. http://dx.doi.org/10.1016/j.ijrefrig.2013.05.008
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Nomenclature A CC Cp c Dh fi G h ID j Nu Pr Re ReLo r U k T x
1.
2
area, m vena-contraction coefficient specific heat, kJkg1K1 convective film constant hydraulic diameters, mm interfacial roughness factor mass flux, kgm2s1 heat transfer coefficient, Wm2K1 inner diameter, mm superficial velocity, ms1 Nusselt number, hDk1 Prandtl number, Cpm k1 Reynolds number, GDm1 Reynolds number with only liquid internal radius of tube, m Overall heat transfer coefficient, Wm2 K1 thermal conductivity, Wm1K1 temperature, C vapor quality
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 temperature 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 relatively high liquid and vapor thermal conductivities, and low liquid viscosity and surface tension. In low-temperature applications, CO2 refrigeration systems with a condensation process in the cycle, such as cascade systems, have been widely 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 temperature was 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.
Greek symbols a void fraction D difference m viscosity, m2s1 q upper angle of the tube not wetted by stratified liquid, rad r density, kgm3 s surface tension, Nm1 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
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 coefficients 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 diameters between 0.96 and 1.289 mm. Zhang et al. (2012) utilized R22, R410A, and R407C as working fluids. They determined that the interface shear stress was the most 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 experimental 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 from 0.4 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.
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Table 1 e Existing studies on condensation in horizontal microchannels. References Haui and Koyama (2004) Park and Hrnjak (2009) Zhang et al. (2012) Del Col et al. (2010) Wang and Rose (2006) Shin and Kim (2005) Cavallini et al. (2005) Koyama et al. (2003) Wang et al. (2002) Garimella et al. (2005)
Fluids
Geometry of test tubes
CO2
Circular channels (Dh: 1.31 mm, 10 multiport)
CO2
Circular channels (Dh: 0.89 mm, 10 multiport)
R22, R410A, R407C R1234yf, R134a R134a
A single circular tube (Dh: 1.088, 1.289 mm)
R134a R134a, R410A R134a R134a R134a
A single circular tube (Dh: 0.96 mm) Single square/triangle/inverted triangle/rectangle channel (Dh: 0.577e1.2 mm) Single circular and rectangular tubes (Dh: 0.494e1.067 mm) Rectangular channels (Dh: 1.4 mm, 13 multiport) Rectangular channels (Dh: 1.11 mm, 8 multiport), (Dh: 0.80 mm, 19 multiport) Rectangular channels (Dh: 1.46 mm, 10 multiport) Circular channels (Dh: 0.5e4.91 mm, 1, 10, 17, 23 multiport)
(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 coefficient was underestimated by the existing models developed for macroscale and minichannels 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 flow patterns 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 in microchannels 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 microchannel. The effects of mass flux and condensation temperature 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
Test conditions
Measurements/ calculations
G: 123.2e315.2 kgm2s1 Tcond: 21.63e31.33 C G: 200e800 kgm2s1 Tcond: 15, 25 C G: 300e600 kgm2s1 Tcond: 30, 40 C G: 200e1000 kgm2s1 Tcond: 40 C G: 500 kgm2s1 Tcond: 50 C G: 100e600 kgm2s1 Tcond: 40 C G: 200e1400 kgme2s1 Tcond: 40 C G: 100e700 kgm2s1 Pcond: 1.7 MPa G: 75e750 kgm2s1 Tcond: 61e66.5 C G: 150e750 kgm2s1
Heat transfer coefficients Heat transfer coefficients, pressure drop Heat transfer coefficients, pressure drop Heat transfer coefficients, pressure drop Heat transfer coefficients Heat transfer coefficients Heat transfer coefficients, pressure drop Heat transfer coefficients, pressure drop Heat transfer coefficients Pressure drop
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 subcooler. 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
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Water bath Flow meter T P
Chiller Relief valve
Chiller
T
P
Test section
Heat exchanger
T
T
P
T
Flow meter
Receiver Differential tank Chiller Pressure transducer
Heat exchanger Chiller
P
Pressure
T
Temperature
T Heat exchanger
Needle valve Bypass valve CO2 tank Gear pump
T
T
P
P Mass flow meter
Sight glass
Fig. 1 e Schematics of experimental setup.
by the probe-type thermocouple. The condensation temperature 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
tests were calibrated simultaneously by using the constant temperature bath, and the differences among the thermocouples 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
9.52 Front view
1.8 2
4
1.2
10
18
Microchannel 20 Brine outlet
T
Side view
4
1.8
10 (ID=8)
T
10 (ID=8) 9.52 Brine inlet
CO outlet
40
Top view
40
Microchannel Brine inlet
Brine outlet
450 Auxiliary tubes DP
Fig. 2 e Details of test section.
P
100
T
20
T
18
P
CO inlet
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meter and the volumetric flow meter, respectively. The calibration 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 hydraulic 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 microchannel 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 channel. 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 Wm1K1. 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 thermophysical 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
100 80
Mean deviation (%)
60 40 20 0
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 conductivity of 0.3691 Wm1K1 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 expansional and contractional pressure drop from header to each port of microchannel 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 accelerational 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 1 1 þ ¼ UA ho Ao hi Ai
(1)
Q ¼ UADTlm
(2)
Nuo ¼ 0:0971Re0:5 Pr0:3
(3)
xtest;inlet ¼ itest;inlet il ivl
(4)
_ CO2 ivl xtest;outlet ¼ xtest;inlet Q_ test = m
(5)
DPtot ¼ DPfric þ DPace þ DPc DPe þ DPaux
(6)
-20 -40 -60 -80 -100 2.6
Table 2 e Test conditions.
2.8
3.0 2
Heat transfer coefficient (kW/m K) Fig. 3 e Deviations between measured and predicted single-phase heat transfer coefficient of CO2.
Test conditions Mass flux Condensation temperature Vapor quality
Ranges 400 to 1000 kgm2s1 5 to 5 C 0.0 to 1.0
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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.
Table 3 e Uncertainties in variables. Variables
Uncertainties
Fluid temperature Pressure of CO2
0.1 C 0.13% of the full scale 1.0% of the full range 0.2% of measurement 4.06%
Volume flow rate of cooling fluid Mass flow rate of CO2 Differential pressure drop of CO2 Condensation heat transfer coefficient of CO2 Vapor quality of CO2
Range 250e350 C 14.7e1000 psia 1.14e11.36 lpm 0e0.3 kgs1
9.26% 6.1%
# " #) x2i x2o ð1 xo Þ2 ð1 xi Þ2 þ þ rv ao rl ð1 ao Þ rv ai rl ð1 ai Þ
(7)
1 x2 x2 DPe ¼ G2 s s 1 þ rl ð1 aÞ rv a DPc ¼
G2 2rl
(8)
2
1 1 þ 1 s2 ð1 þ xðrl rv Þ=rl Þ Cc
(9)
Table 2 shows the present test conditions. The mass flux was varied to 400, 600, 800, and 1000 kgm2s1, and the condensation temperature changed from 5 to 5 C. Because the pressure 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
7
Heat transfer coefficient
7
Massflux: 800 kgm s
Massflux: 600 kgm s
Massflux: 1000 kgm s
6
6
6
6
5
5
5
5
4
4
4
4
3
3
3
3
-2
-1
Heat transfer coefficient (kW m K )
3.1.
7
7
Massflux: 400 kgm s
Results and discussion
Fig. 4 shows heat transfer coefficient with variation of the condensation temperature with mass fluxes of 400, 600, 800, and 1000 kgm2s1. It was found that the heat transfer coefficient increased with the decrease of condensation temperature for all mass flux conditions. For example, when the mass flux was 600 kgm2s1, 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 kgm2s1, 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 between the liquid and vapor, and its thermal resistance decreased. As noted, the condensation heat transfer coefficient was enhanced with decrease in the liquid film thickness, which acts as a thermal resistance in the annular flow condition. Fig. 4 also shows that the heat transfer coefficient dropped at a certain vapor quality at a mass flux of 1000 kgm2s1. 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 surface 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
(" DPacc ¼ G2
3.
2
2
2
o
-5 0 5
1
0 0.2
0.4
0.6
0.8
Vapor quality
1.0
Tcond ( C) -5 0 5
1
0 0.0
2
o
-5 0 5
1
0.4
0.6
Vapor quality
0.8
1.0
0.0
0.2
0.4
0.6
Vapor quality
0.8
-5 0 5
1
0 0.2
o
Tcond ( C)
Tcond ( C)
o
Tcond ( C)
1.0
0 0.0
0.2
0.4
0.6
0.8
1.0
Vapor quality
Fig. 4 e Variation of heat transfer coefficient with condensation temperature under different mass flux conditions.
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7
7
o
7 o
-2
6
6
6
5
5
5
4
4
4
3
3
3
Massflux -2 -1 400 kgm s -2 -1 600 kgm s -2 -1 800 kgm s -2 -1 1000 kgm s
2
1
1
0 0.2
0.4
0.6
0.8
Massflux -2 -1 400 kgm s -2 -1 600 kgm s -2 -1 800 kgm s -2 -1 1000 kgm s
2
0 0.0
1.0
o
Tcond : 5 C
Tcond : 0 C
-1
Heat transfer coefficient (kW m K )
Tcond : -5 C
Massflux -2 -1 400 kgm s -2 -1 600 kgm s -2 -1 800 kgm s -2 -1 1000 kgm s
2
1
0 0.2
Vapor quality
0.4
0.6
0.8
1.0
0.0
Vapor quality
0.2
0.4
0.6
0.8
1.0
Vapor quality
Fig. 5 e Variation of heat transfer coefficient with mass flux under the different condensation temperature.
reflected in the existing models, 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 increase of mass flux (Lu et al., 2004). Maldistribution is the noneven 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 quantitative effect of flow maldistribution on the heat transfer can be estimated by the Bielskus (2011). The cooling capacity of the heat exchanger with uniform distribution was increased by an average of 34% than that of the heat exchanger with maldistribution. The detail quantitative analyses of the maldistribution 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
200
180
-5 0 5
160
Pressure drop
Fig. 6 shows the variation of the pressure drop with condensation temperature under different mass flux conditions.
200
200 o
Tcond ( C)
-1
Pressure drop (kPam )
3.2.
200 o
180
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 increase of mass flux from 400 to 600, 800 and 1000 kgm2s1 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 increase of mass flux and the decrease of condensation temperature.
o
Tcond ( C)
180
-5 0 5
160
o
Tcond ( C)
180
-5 0 5
160
140
140
140
120
120
120
120
100
100
100
100
80
80
80
80
60
60
60
60
40
40
40
40
20
0 0.2
0.4
0.6
0.8
Vapor quality
1.0
20
20
Massflux 400: kgm s
Massflux: 600 kgm s 0 0.0
Massflux: 1000 kgm s
Massflux: 800 kgm s 0
0.2
0.4
0.6
0.8
Vapor quality
-5 0 5
160
140
20
Tcond ( C)
1.0
0.0
0.2
0.4
0.6
Vapor quality
0.8
1.0
0 0.0
0.2
0.4
0.6
0.8
1.0
Vapor quality
Fig. 6 e Variation of pressure drop with the condensation temperature under the different mass flux condition.
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-2
-1
Heat transfer coefficient (kWm K )
13
13 Data 5 C Data -5 C Data -15 C (Park and Hrnjak, 2009)
12 11
12 11
13 Data 5 C Data -5 C Data -15 C (Park and Hrnjak, 2009)
11
10
10
10
9
9
9
8
8
8
7
7
7
6
6
6
5
5
5
4
4
4
3
3
3
2
2
1
G: 400 kgm s
0 0.0
0.2
0.4
0.6
0.8
1.0
2
1 0 0.0
Data 5 C Data -5 C Data -15 C (Park and Hrnjak, 2009)
12
1
G: 600 kgm s
G: 800 kgm s
0 0.2
Vapor quality
0.4
0.6
0.8
1.0
Vapor quality
0.0
0.2
0.4
0.6
0.8
1.0
Vapor quality
Fig. 7 e Comparison of the present heat transfer coefficient with the existing data.
The pressure drop increased with the decrease of the condensation temperature. For instance, when the mass flux was 400 kgm2s1, 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 kgm2s1, 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 kPam1, 50.2 kPam1, 70.3 kPam1, and 91.3 kPam1 at 400, 600, 800, and 1000 kgm2s1, 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 drop with vapor quality increased with
20
+200%
15
+50%
-2
-1
Predicted h (kWm K )
+100%
0%
10
5
Bandhauer et al. (2005) Cavallini et al. (2002) Thome et al. (2003) 0 0
5
10
15 -2
20
-1
Measured h (kWm K ) 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 gradually 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 experimental 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 condensation temperature and mass flux on the heat transfer coefficient. However, the data of Park and Hrnjak (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 maldistribution is delineated with smaller hydraulic diameter. The comparatively larger flow resistance at header to microchannel with smaller hydraulic diameter forced the liquid flows toward other channels, thereby causing a more uniformed 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 each model. All the models overpredicted the present experimental data, and the deviation linearly increased with the increase of heat transfer coefficient. Among the existing models, the smallest mean deviation was 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
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Table 4 e Ranges of applicability of the existing models. References
Models
Fluids
Thome et al. (2003)
Heat transfer
Cavallini et al. (2002) Bandhauer et al. (2005) Garimella et al. (2005)
Heat transfer
Geometry of test tubes
Heat transfer
R11, R12, R22, R32, R113, R125, R134a, R236ea, R404A, R410A, Propane, n-butane, Iso-butane, Propylene R22, R134a, R125, R32, R236ea, R407C, R410A R134a
Pressure drop
R134a
Lee and Lee (2001)
Pressure drop
Water-air
Mishima and Hibiki (1996) Friedel (1979) McAdams et al. (1942)
Pressure drop
Water-air e e
Pressure drop Pressure drop
12
-2
-1
h (kWm K )
10
T
: -5 C
G: 400 kgm s
-2
-1
h (kWm K )
G: 150e750 kgm2s1
Horizontal microchannels (Dh: 0.78e6.67 mm) Vertical upward round tube (Di: 1e4 mm) (Dh > 1 mm) e
ReLo: 175e17,700 X: 0.303e79.4 jv : 0.0896e79.3 ms1 jl: 0.0116e1.67 ms1 ml m1 v < 1000 Similar vapor and liquid velocity
10
G: 400 kgm s
6 4
2
2
2
0.4
0.6
0.8
1.0
0 0.0
0.2
0.4
0.6
0.8
0 1.0 0.0
12 Thome et al. Measured data
T
: -5 C
G: 600 kgm s
10
T
:0C
8
8
6
6
4
4
4
2
2
2
0.6
0.8
0 1.0 0.0
0.2
0.4
0.6
0.8
0 1.0 0.0
12 Thome et al. Measured data
10
T
:0C
8
8
6
6
4
4
4
2
2
0 0.0
: -5 C
G: 800 kgm s
0.2
0.4
0.6
0.8
12 10
0 1.0 0.0
0.2
0.4
0.6
0.8
12 Thome et al. Measured data
10
1.0
0
8
8
6
6
4
4
2 0.2
0.4
0.6
Vapor quality
0.8
2 0 1.0 0.0
0.8
1.0
0.2
0.4
:5C
T
G: 600 kgm s
0.6
0.8
1.0
0.2
0.4
T
:5C
G: 800 kgm s
0.6
0.8
1.0
Thome et al. Measured data
4 T
:0C
G: 1000 kgm s
G: 1000 kgm s
0 0.0
0.0
10
6 : -5 C
0.6
12 Thome et al. Measured data
8
T
0.4
Thome et al. Measured data
10
G: 800 kgm s
6 T
0.2
:5C
T
12 Thome et al. Measured data
8
2
Annular Bubbly, wispy-annular
G: 400 kgm s
Thome et al. Measured data
10
G: 600 kgm s
6
0.4
e
12 Thome et al. Measured data
8
0.2
Thome et al. Measured data
:0C
4
0.2
G: 150e750 kgm2s1
12 T
Thome et al. Measured data
6
10 -1
Horizontal microchannels (Dh: 0.51e1.52 mm) Horizontal microchannels (Dh: 0.50e4.91 mm)
Annular, stratified, slug Annular, mist, and disperse wave Annular, disperse wave, mist, discrete wave, intermittent Laminar, Turbulent
4
12
-2
G: 100e750 kgm2s1
6
0 0.0
h (kWm K )
Plain tube (Di: 8 mm)
8
10
-1
Annular, intermittent, stratified-wavy, fully stratified, mist flow
8
12
-2
G: 24e1022 kgm s
8
0 0.0
h (kWm K )
10
Flow regimes
2 1
Horizontal plain tubes (Di: 3.1e21.4 mm)
12 Thome et al. Measured data
Applicable range
0.2
0.4
0.6
Vapor quality
0.8
T
2 0 1.0 0.0
:5C
G: 1000 kgm s
0.2
0.4
0.6
0.8
1.0
Vapor quality
Fig. 9 e Comparison of the measured data and predicted heat transfer coefficients by the Thome et al. model.
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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 flow pattern can be advanced to relatively lower vapor quality for the microchannel. Besides, the application ranges of mass flux of other models are limited to 750 kgm2s1 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 patterns 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 distributions 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 distributions to each port were categorized from case 1 to 5 as summarized 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 flow with 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 estimation of the heat transfer coefficient for the annular flow and the stratified flow as shown in Eqs. (10) and (11), respectively. 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 the mean deviation between the predicted and the experimental results was found by considering the flow maldistribution 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 conditions showed similar results with cases 4 and 5. The average
-1
Predicted pressure drop (kPam )
200
Garimella et al. (2005) Lee and Lee (2001) Mishima and Hibiki (1996)
180 160
+50%
0%
+200%
140 120 100 -50%
80 60 40 Fridel (1979) McAdams et al. (1942)
20 0 0
20
40
60
80
100
120
140
160
180
200
-1
Measured pressure drop (kPam ) Fig. 10 e Comparison of pressure drop between the present and the existing study.
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 microchannel was significantly affected by the non-even distribution 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. lL fi d
(10)
hf rq þ ð2p qÞrhc 2pr
(11)
hannular ¼ cRenL Prm L
hstratified ¼
hmist ¼ 0:023
!0:8 ktp GD Pr0:4 tp D mtp
(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
Table 5 e Simulated maldistribution inside microchannel and mean deviation of heat transfer coefficient between experiments and predictions.
Case Case Case Case Case
1 2 3 4 5
Maldistribution assumptions
Ch 1
Ch 2
Ch 3
Ch 4
Ch 5
Ch 6
Ch 7
Mean deviation (%)
BR: 7 LDR: 2, BR: 5 LDR: 2, BR: 4, VDR: 1 LDR: 2, BR: 2, VDR: 3 LDR: 4, BR: 2, VDR: 1
A S S S S
A A A A S
A A A M A
A A M M M
A A A M A
A A A A S
A S S S S
51.84 26.75 18.56 15.58 20.25
(BR: Balanced region, LDR: Liquid dominant region, VDR: Vapor dominant 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 7 e1 6 6 8
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 Lee model (2001), the McAdams 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 homogenous 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 difference between the vapor and liquid could become smaller with the decrease of tube diameter, which resulted in better predictions of pressure by a homogenous flow model than by a separated flow model. 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 better matched with 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 drop of CO2 in multiport rectangular microchannels were experimentally investigated with variation of the mass flux and the condensation temperature from 400 to 1000 kgm2s1 and from 5 to 5 C, respectively. The effect of condensation temperature and the mass 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 predictability of the existing models. Several pressure drop models 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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references
Abdelall, F.F., Hahn, G., Ghiaasiaan, S.M., Abdel-Khalik, S.I., Jeter, S.S., Yoda, M., Sadowki, D.L., 2005. Pressure drop caused by abrupt flow area changes in small channels. Exp. Therm. Fluid Sci. 29, 425e434. Ahmad, M., Berthoud, G., Mercier, P., 2009. General characteristics of two-phase flow distribution in a compact heat exchanger. Int. J. Heat Mass. Trans. 52, 442e450. Akers, W.W., Deans, H.A., Crosser, O.K., 1959. Condensing heat transfer within horizontal tubes. Chem. Eng. Prog. Symp. Ser. 55, 171e176. Bandhauer, T.M., Agarwal, A., Garimella, S., 2005. Measurement and modeling of condensation heat transfer coefficients in circular microchannels. In: Proceedings of the 3rd International Conference on Microchannels and Minichannels, Toronto, Ontario, Canada, ICMM2005e75248. Bansal, P., 2012. A review e status of CO2 as a low temperature refrigerant: fundamentals and R&D opportunities. Appl. Therm. Eng. 41, 18e29. Bielskus, A.V., 2011. Two Phase Flow Distribution in Parallel Flow Heat Exchangers e Experimentally Verified Model. Dissertation for the Degree of Master of Science. Graduate College of the University of Illinois at Urbana-Champaign. Carey, V.P., 2008. LiquideVapor Phase-Change Phenomena, second ed. Taylor & Francis Group, New York, NYS, pp. 559e653. Cavallini, A., Censi, G., Del Col, D., Doretti, L., Longo, G.A., Rossetto, L., 2002. Condensation of halogenated refrigerants inside smooth tubes. HVAC&R Res. 8, 429e451. Cavallini, A., Censi, G., Del Col, D., Doretti, L., Matkovic, M., Rosseto, L., Zilio, C., 2006. Condensation in horizontal smooth tubes: a new heat transfer model for heat exchanger design. Heat Transfer. Eng. 27, 31e38. Cavallini, A., Del Col, D., Doretti, L., Matkovic, M., Rosseto, L., Zilio, C., 2005. Condensation heat transfer and pressure gradient inside multiport minichannels. Heat Transfer Eng. 26, 45e55. Chen, S.L., Gerner, F.M., Tien, C.L., 1987. General film condensation correlations. Exp. Heat Transfer 1, 93e107. Del Col, D., Torresin, D., Cavallini, A., 2010. Heat transfer and pressure drop during condensation of the low GWP refrigerant R1234yf. Int. J. Refrig. 33, 1307e1318. Friedel, L., 1979. Improved friction pressure drop correlations for horizontal and vertical two-phase pipe flow. In: Proceedings of European Two-phase Flow Group Meeting, Ispra, Italy, Paper E2. Garimella, S., Agarwal, A., Killion, J.D., 2005. Condensation pressure drop in circular microchannels. Heat Trans. Eng. 26, 1e8. Haui, X., Koyama, S., 2004. An experimental study of carbon dioxide condensation in mini channels. J. Them Sci. 13, 358e365. Kang, P., Heo, J., Yun, R., 2012. Condensation heat transfer characteristics of CO2 in a horizontal smooth tube. Int. J. Refrig.. http://dx.doi.org/10.1016/j.bbr.2011.03.031. Koyama, S., Kuwahara, K., Nakashita, K., 2002. An experimental study on condensation of R134a in a multi-port extruded tube. In: Proc. Of the 9th International Refrigeration and Air Conditioning Conference at Purdue. USA, pp. 197e204. Koyama, S., Kuwahara, K., Nakashita, K., Yamamoto, K., 2003. An experimental study on condensation of refrigerant R134a in a multi-port extruded tube. Int. J. Refrig. 24, 425e432. Lee, H.J., Lee, S.Y., 2001. Pressure drop correlations for two-phase flow within horizontal rectangular channels with small heights. Int. J. Multiphase Flow 27, 783e796. Lu, M., Yang, B., Wang, C., 2004. Numerical study of flow maldistribution on the flow and heat transfer for multi-channel cold-plates. In: Proc. of the 20th IEEE SEMI-THERM Symposium, San Jose, CA, USA, pp. 205e212.
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journal homepage: www.elsevier.com/locate/ijrefrig
Condensation heat transfer and pressure drop characteristics of CO2 in a microchannel Jaehyeok Heo a, Hanvit Park b, Rin Yun b,* a
Department of Mechanical Engineering, Korea University, Anam-Dong, Sungbuk-Ku, Seoul 136-713, Korea Department of Mechanical Engineering, Hanbat National University, Duckmyung-Dong, San 16-1, Daejeon 305-719, Korea b
article info
abstract
Article history:
The condensation heat transfer coefficient and pressure drop of CO2 in a multiport
Received 19 November 2012
microchannel with a hydraulic diameter of 1.5 mm was investigated with variation of the
Received in revised form
mass flux from 400 to 1000 kgm2s1 and of the condensation temperature from 5 to 5 C.
7 April 2013
The heat transfer coefficient and pressure drop increased with the decrease of conden-
Accepted 17 May 2013
sation temperature and the increase of mass flux. However, the rate of increase of the heat
Available online 28 May 2013
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
Keywords:
for heat transfer coefficient overpredicted the experimental data, and the deviation
Condensation
increased at high vapor quality and at high heat transfer coefficient. The smallest mean
Heat transfer
deviation of 51.8% was found by the Thome et al. model. For the pressure drop, the
Carbon dioxide
Mishima and Hibiki model showed mean deviation of 29.1%.
Heat transfer coefficient
ª 2013 Elsevier Ltd and IIR. All rights reserved.
Pressure drop Flow complexity Microchannel
Transfert de chaleur lors de la condensation et caracte´ristiques de chute de pression du CO2 a` l’inte´rieur d’un microcanal Mots cle´s : condensation ; transfert de chaleur ; dioxyde de carbone ; coefficient de transfert de chaleur ; chute de pression ; complexite´ de l’e´coulement ; microcanal
* Corresponding author. Tel.: þ82 42 821 1732; fax: þ82 42 821 1587. E-mail address: [email protected] (R. Yun). 0140-7007/$ e see front matter ª 2013 Elsevier Ltd and IIR. All rights reserved. http://dx.doi.org/10.1016/j.ijrefrig.2013.05.008
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Nomenclature A CC Cp c Dh fi G h ID j Nu Pr Re ReLo r U k T x
1.
2
area, m vena-contraction coefficient specific heat, kJkg1K1 convective film constant hydraulic diameters, mm interfacial roughness factor mass flux, kgm2s1 heat transfer coefficient, Wm2K1 inner diameter, mm superficial velocity, ms1 Nusselt number, hDk1 Prandtl number, Cpm k1 Reynolds number, GDm1 Reynolds number with only liquid internal radius of tube, m Overall heat transfer coefficient, Wm2 K1 thermal conductivity, Wm1K1 temperature, C vapor quality
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 temperature 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 relatively high liquid and vapor thermal conductivities, and low liquid viscosity and surface tension. In low-temperature applications, CO2 refrigeration systems with a condensation process in the cycle, such as cascade systems, have been widely 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 temperature was 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.
Greek symbols a void fraction D difference m viscosity, m2s1 q upper angle of the tube not wetted by stratified liquid, rad r density, kgm3 s surface tension, Nm1 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
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 coefficients 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 diameters between 0.96 and 1.289 mm. Zhang et al. (2012) utilized R22, R410A, and R407C as working fluids. They determined that the interface shear stress was the most 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 experimental 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 from 0.4 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.
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Table 1 e Existing studies on condensation in horizontal microchannels. References Haui and Koyama (2004) Park and Hrnjak (2009) Zhang et al. (2012) Del Col et al. (2010) Wang and Rose (2006) Shin and Kim (2005) Cavallini et al. (2005) Koyama et al. (2003) Wang et al. (2002) Garimella et al. (2005)
Fluids
Geometry of test tubes
CO2
Circular channels (Dh: 1.31 mm, 10 multiport)
CO2
Circular channels (Dh: 0.89 mm, 10 multiport)
R22, R410A, R407C R1234yf, R134a R134a
A single circular tube (Dh: 1.088, 1.289 mm)
R134a R134a, R410A R134a R134a R134a
A single circular tube (Dh: 0.96 mm) Single square/triangle/inverted triangle/rectangle channel (Dh: 0.577e1.2 mm) Single circular and rectangular tubes (Dh: 0.494e1.067 mm) Rectangular channels (Dh: 1.4 mm, 13 multiport) Rectangular channels (Dh: 1.11 mm, 8 multiport), (Dh: 0.80 mm, 19 multiport) Rectangular channels (Dh: 1.46 mm, 10 multiport) Circular channels (Dh: 0.5e4.91 mm, 1, 10, 17, 23 multiport)
(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 coefficient was underestimated by the existing models developed for macroscale and minichannels 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 flow patterns 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 in microchannels 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 microchannel. The effects of mass flux and condensation temperature 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
Test conditions
Measurements/ calculations
G: 123.2e315.2 kgm2s1 Tcond: 21.63e31.33 C G: 200e800 kgm2s1 Tcond: 15, 25 C G: 300e600 kgm2s1 Tcond: 30, 40 C G: 200e1000 kgm2s1 Tcond: 40 C G: 500 kgm2s1 Tcond: 50 C G: 100e600 kgm2s1 Tcond: 40 C G: 200e1400 kgme2s1 Tcond: 40 C G: 100e700 kgm2s1 Pcond: 1.7 MPa G: 75e750 kgm2s1 Tcond: 61e66.5 C G: 150e750 kgm2s1
Heat transfer coefficients Heat transfer coefficients, pressure drop Heat transfer coefficients, pressure drop Heat transfer coefficients, pressure drop Heat transfer coefficients Heat transfer coefficients Heat transfer coefficients, pressure drop Heat transfer coefficients, pressure drop Heat transfer coefficients Pressure drop
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 subcooler. 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
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Water bath Flow meter T P
Chiller Relief valve
Chiller
T
P
Test section
Heat exchanger
T
T
P
T
Flow meter
Receiver Differential tank Chiller Pressure transducer
Heat exchanger Chiller
P
Pressure
T
Temperature
T Heat exchanger
Needle valve Bypass valve CO2 tank Gear pump
T
T
P
P Mass flow meter
Sight glass
Fig. 1 e Schematics of experimental setup.
by the probe-type thermocouple. The condensation temperature 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
tests were calibrated simultaneously by using the constant temperature bath, and the differences among the thermocouples 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
9.52 Front view
1.8 2
4
1.2
10
18
Microchannel 20 Brine outlet
T
Side view
4
1.8
10 (ID=8)
T
10 (ID=8) 9.52 Brine inlet
CO outlet
40
Top view
40
Microchannel Brine inlet
Brine outlet
450 Auxiliary tubes DP
Fig. 2 e Details of test section.
P
100
T
20
T
18
P
CO inlet
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meter and the volumetric flow meter, respectively. The calibration 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 hydraulic 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 microchannel 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 channel. 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 Wm1K1. 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 thermophysical 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
100 80
Mean deviation (%)
60 40 20 0
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 conductivity of 0.3691 Wm1K1 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 expansional and contractional pressure drop from header to each port of microchannel 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 accelerational 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 1 1 þ ¼ UA ho Ao hi Ai
(1)
Q ¼ UADTlm
(2)
Nuo ¼ 0:0971Re0:5 Pr0:3
(3)
xtest;inlet ¼ itest;inlet il ivl
(4)
_ CO2 ivl xtest;outlet ¼ xtest;inlet Q_ test = m
(5)
DPtot ¼ DPfric þ DPace þ DPc DPe þ DPaux
(6)
-20 -40 -60 -80 -100 2.6
Table 2 e Test conditions.
2.8
3.0 2
Heat transfer coefficient (kW/m K) Fig. 3 e Deviations between measured and predicted single-phase heat transfer coefficient of CO2.
Test conditions Mass flux Condensation temperature Vapor quality
Ranges 400 to 1000 kgm2s1 5 to 5 C 0.0 to 1.0
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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.
Table 3 e Uncertainties in variables. Variables
Uncertainties
Fluid temperature Pressure of CO2
0.1 C 0.13% of the full scale 1.0% of the full range 0.2% of measurement 4.06%
Volume flow rate of cooling fluid Mass flow rate of CO2 Differential pressure drop of CO2 Condensation heat transfer coefficient of CO2 Vapor quality of CO2
Range 250e350 C 14.7e1000 psia 1.14e11.36 lpm 0e0.3 kgs1
9.26% 6.1%
# " #) x2i x2o ð1 xo Þ2 ð1 xi Þ2 þ þ rv ao rl ð1 ao Þ rv ai rl ð1 ai Þ
(7)
1 x2 x2 DPe ¼ G2 s s 1 þ rl ð1 aÞ rv a DPc ¼
G2 2rl
(8)
2
1 1 þ 1 s2 ð1 þ xðrl rv Þ=rl Þ Cc
(9)
Table 2 shows the present test conditions. The mass flux was varied to 400, 600, 800, and 1000 kgm2s1, and the condensation temperature changed from 5 to 5 C. Because the pressure 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
7
Heat transfer coefficient
7
Massflux: 800 kgm s
Massflux: 600 kgm s
Massflux: 1000 kgm s
6
6
6
6
5
5
5
5
4
4
4
4
3
3
3
3
-2
-1
Heat transfer coefficient (kW m K )
3.1.
7
7
Massflux: 400 kgm s
Results and discussion
Fig. 4 shows heat transfer coefficient with variation of the condensation temperature with mass fluxes of 400, 600, 800, and 1000 kgm2s1. It was found that the heat transfer coefficient increased with the decrease of condensation temperature for all mass flux conditions. For example, when the mass flux was 600 kgm2s1, 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 kgm2s1, 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 between the liquid and vapor, and its thermal resistance decreased. As noted, the condensation heat transfer coefficient was enhanced with decrease in the liquid film thickness, which acts as a thermal resistance in the annular flow condition. Fig. 4 also shows that the heat transfer coefficient dropped at a certain vapor quality at a mass flux of 1000 kgm2s1. 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 surface 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
(" DPacc ¼ G2
3.
2
2
2
o
-5 0 5
1
0 0.2
0.4
0.6
0.8
Vapor quality
1.0
Tcond ( C) -5 0 5
1
0 0.0
2
o
-5 0 5
1
0.4
0.6
Vapor quality
0.8
1.0
0.0
0.2
0.4
0.6
Vapor quality
0.8
-5 0 5
1
0 0.2
o
Tcond ( C)
Tcond ( C)
o
Tcond ( C)
1.0
0 0.0
0.2
0.4
0.6
0.8
1.0
Vapor quality
Fig. 4 e Variation of heat transfer coefficient with condensation temperature under different mass flux conditions.
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7
7
o
7 o
-2
6
6
6
5
5
5
4
4
4
3
3
3
Massflux -2 -1 400 kgm s -2 -1 600 kgm s -2 -1 800 kgm s -2 -1 1000 kgm s
2
1
1
0 0.2
0.4
0.6
0.8
Massflux -2 -1 400 kgm s -2 -1 600 kgm s -2 -1 800 kgm s -2 -1 1000 kgm s
2
0 0.0
1.0
o
Tcond : 5 C
Tcond : 0 C
-1
Heat transfer coefficient (kW m K )
Tcond : -5 C
Massflux -2 -1 400 kgm s -2 -1 600 kgm s -2 -1 800 kgm s -2 -1 1000 kgm s
2
1
0 0.2
Vapor quality
0.4
0.6
0.8
1.0
0.0
Vapor quality
0.2
0.4
0.6
0.8
1.0
Vapor quality
Fig. 5 e Variation of heat transfer coefficient with mass flux under the different condensation temperature.
reflected in the existing models, 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 increase of mass flux (Lu et al., 2004). Maldistribution is the noneven 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 quantitative effect of flow maldistribution on the heat transfer can be estimated by the Bielskus (2011). The cooling capacity of the heat exchanger with uniform distribution was increased by an average of 34% than that of the heat exchanger with maldistribution. The detail quantitative analyses of the maldistribution 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
200
180
-5 0 5
160
Pressure drop
Fig. 6 shows the variation of the pressure drop with condensation temperature under different mass flux conditions.
200
200 o
Tcond ( C)
-1
Pressure drop (kPam )
3.2.
200 o
180
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 increase of mass flux from 400 to 600, 800 and 1000 kgm2s1 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 increase of mass flux and the decrease of condensation temperature.
o
Tcond ( C)
180
-5 0 5
160
o
Tcond ( C)
180
-5 0 5
160
140
140
140
120
120
120
120
100
100
100
100
80
80
80
80
60
60
60
60
40
40
40
40
20
0 0.2
0.4
0.6
0.8
Vapor quality
1.0
20
20
Massflux 400: kgm s
Massflux: 600 kgm s 0 0.0
Massflux: 1000 kgm s
Massflux: 800 kgm s 0
0.2
0.4
0.6
0.8
Vapor quality
-5 0 5
160
140
20
Tcond ( C)
1.0
0.0
0.2
0.4
0.6
Vapor quality
0.8
1.0
0 0.0
0.2
0.4
0.6
0.8
1.0
Vapor quality
Fig. 6 e Variation of pressure drop with the condensation temperature under the different mass flux condition.
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-2
-1
Heat transfer coefficient (kWm K )
13
13 Data 5 C Data -5 C Data -15 C (Park and Hrnjak, 2009)
12 11
12 11
13 Data 5 C Data -5 C Data -15 C (Park and Hrnjak, 2009)
11
10
10
10
9
9
9
8
8
8
7
7
7
6
6
6
5
5
5
4
4
4
3
3
3
2
2
1
G: 400 kgm s
0 0.0
0.2
0.4
0.6
0.8
1.0
2
1 0 0.0
Data 5 C Data -5 C Data -15 C (Park and Hrnjak, 2009)
12
1
G: 600 kgm s
G: 800 kgm s
0 0.2
Vapor quality
0.4
0.6
0.8
1.0
Vapor quality
0.0
0.2
0.4
0.6
0.8
1.0
Vapor quality
Fig. 7 e Comparison of the present heat transfer coefficient with the existing data.
The pressure drop increased with the decrease of the condensation temperature. For instance, when the mass flux was 400 kgm2s1, 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 kgm2s1, 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 kPam1, 50.2 kPam1, 70.3 kPam1, and 91.3 kPam1 at 400, 600, 800, and 1000 kgm2s1, 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 drop with vapor quality increased with
20
+200%
15
+50%
-2
-1
Predicted h (kWm K )
+100%
0%
10
5
Bandhauer et al. (2005) Cavallini et al. (2002) Thome et al. (2003) 0 0
5
10
15 -2
20
-1
Measured h (kWm K ) 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 gradually 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 experimental 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 condensation temperature and mass flux on the heat transfer coefficient. However, the data of Park and Hrnjak (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 maldistribution is delineated with smaller hydraulic diameter. The comparatively larger flow resistance at header to microchannel with smaller hydraulic diameter forced the liquid flows toward other channels, thereby causing a more uniformed 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 each model. All the models overpredicted the present experimental data, and the deviation linearly increased with the increase of heat transfer coefficient. Among the existing models, the smallest mean deviation was 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
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Table 4 e Ranges of applicability of the existing models. References
Models
Fluids
Thome et al. (2003)
Heat transfer
Cavallini et al. (2002) Bandhauer et al. (2005) Garimella et al. (2005)
Heat transfer
Geometry of test tubes
Heat transfer
R11, R12, R22, R32, R113, R125, R134a, R236ea, R404A, R410A, Propane, n-butane, Iso-butane, Propylene R22, R134a, R125, R32, R236ea, R407C, R410A R134a
Pressure drop
R134a
Lee and Lee (2001)
Pressure drop
Water-air
Mishima and Hibiki (1996) Friedel (1979) McAdams et al. (1942)
Pressure drop
Water-air e e
Pressure drop Pressure drop
12
-2
-1
h (kWm K )
10
T
: -5 C
G: 400 kgm s
-2
-1
h (kWm K )
G: 150e750 kgm2s1
Horizontal microchannels (Dh: 0.78e6.67 mm) Vertical upward round tube (Di: 1e4 mm) (Dh > 1 mm) e
ReLo: 175e17,700 X: 0.303e79.4 jv : 0.0896e79.3 ms1 jl: 0.0116e1.67 ms1 ml m1 v < 1000 Similar vapor and liquid velocity
10
G: 400 kgm s
6 4
2
2
2
0.4
0.6
0.8
1.0
0 0.0
0.2
0.4
0.6
0.8
0 1.0 0.0
12 Thome et al. Measured data
T
: -5 C
G: 600 kgm s
10
T
:0C
8
8
6
6
4
4
4
2
2
2
0.6
0.8
0 1.0 0.0
0.2
0.4
0.6
0.8
0 1.0 0.0
12 Thome et al. Measured data
10
T
:0C
8
8
6
6
4
4
4
2
2
0 0.0
: -5 C
G: 800 kgm s
0.2
0.4
0.6
0.8
12 10
0 1.0 0.0
0.2
0.4
0.6
0.8
12 Thome et al. Measured data
10
1.0
0
8
8
6
6
4
4
2 0.2
0.4
0.6
Vapor quality
0.8
2 0 1.0 0.0
0.8
1.0
0.2
0.4
:5C
T
G: 600 kgm s
0.6
0.8
1.0
0.2
0.4
T
:5C
G: 800 kgm s
0.6
0.8
1.0
Thome et al. Measured data
4 T
:0C
G: 1000 kgm s
G: 1000 kgm s
0 0.0
0.0
10
6 : -5 C
0.6
12 Thome et al. Measured data
8
T
0.4
Thome et al. Measured data
10
G: 800 kgm s
6 T
0.2
:5C
T
12 Thome et al. Measured data
8
2
Annular Bubbly, wispy-annular
G: 400 kgm s
Thome et al. Measured data
10
G: 600 kgm s
6
0.4
e
12 Thome et al. Measured data
8
0.2
Thome et al. Measured data
:0C
4
0.2
G: 150e750 kgm2s1
12 T
Thome et al. Measured data
6
10 -1
Horizontal microchannels (Dh: 0.51e1.52 mm) Horizontal microchannels (Dh: 0.50e4.91 mm)
Annular, stratified, slug Annular, mist, and disperse wave Annular, disperse wave, mist, discrete wave, intermittent Laminar, Turbulent
4
12
-2
G: 100e750 kgm2s1
6
0 0.0
h (kWm K )
Plain tube (Di: 8 mm)
8
10
-1
Annular, intermittent, stratified-wavy, fully stratified, mist flow
8
12
-2
G: 24e1022 kgm s
8
0 0.0
h (kWm K )
10
Flow regimes
2 1
Horizontal plain tubes (Di: 3.1e21.4 mm)
12 Thome et al. Measured data
Applicable range
0.2
0.4
0.6
Vapor quality
0.8
T
2 0 1.0 0.0
:5C
G: 1000 kgm s
0.2
0.4
0.6
0.8
1.0
Vapor quality
Fig. 9 e Comparison of the measured data and predicted heat transfer coefficients by the Thome et al. model.
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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 flow pattern can be advanced to relatively lower vapor quality for the microchannel. Besides, the application ranges of mass flux of other models are limited to 750 kgm2s1 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 patterns 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 distributions 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 distributions to each port were categorized from case 1 to 5 as summarized 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 flow with 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 estimation of the heat transfer coefficient for the annular flow and the stratified flow as shown in Eqs. (10) and (11), respectively. 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 the mean deviation between the predicted and the experimental results was found by considering the flow maldistribution 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 conditions showed similar results with cases 4 and 5. The average
-1
Predicted pressure drop (kPam )
200
Garimella et al. (2005) Lee and Lee (2001) Mishima and Hibiki (1996)
180 160
+50%
0%
+200%
140 120 100 -50%
80 60 40 Fridel (1979) McAdams et al. (1942)
20 0 0
20
40
60
80
100
120
140
160
180
200
-1
Measured pressure drop (kPam ) Fig. 10 e Comparison of pressure drop between the present and the existing study.
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 microchannel was significantly affected by the non-even distribution 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. lL fi d
(10)
hf rq þ ð2p qÞrhc 2pr
(11)
hannular ¼ cRenL Prm L
hstratified ¼
hmist ¼ 0:023
!0:8 ktp GD Pr0:4 tp D mtp
(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
Table 5 e Simulated maldistribution inside microchannel and mean deviation of heat transfer coefficient between experiments and predictions.
Case Case Case Case Case
1 2 3 4 5
Maldistribution assumptions
Ch 1
Ch 2
Ch 3
Ch 4
Ch 5
Ch 6
Ch 7
Mean deviation (%)
BR: 7 LDR: 2, BR: 5 LDR: 2, BR: 4, VDR: 1 LDR: 2, BR: 2, VDR: 3 LDR: 4, BR: 2, VDR: 1
A S S S S
A A A A S
A A A M A
A A M M M
A A A M A
A A A A S
A S S S S
51.84 26.75 18.56 15.58 20.25
(BR: Balanced region, LDR: Liquid dominant region, VDR: Vapor dominant 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 7 e1 6 6 8
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 Lee model (2001), the McAdams 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 homogenous 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 difference between the vapor and liquid could become smaller with the decrease of tube diameter, which resulted in better predictions of pressure by a homogenous flow model than by a separated flow model. 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 better matched with 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 drop of CO2 in multiport rectangular microchannels were experimentally investigated with variation of the mass flux and the condensation temperature from 400 to 1000 kgm2s1 and from 5 to 5 C, respectively. The effect of condensation temperature and the mass 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 predictability of the existing models. Several pressure drop models 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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