Heat transfer engineering an international journal, tập 31, số 12, 2010

Heat Transfer Engineering, 31(12):963–964, 2010
Copyright C Taylor and Francis Group, LLC
ISSN: 0145-7632 print / 1521-0537 online
DOI: 10.1080/01457631003638903

editorial

Advances in Heat Transfer
Engineering
PRADEEP BANSAL
Department of Mechanical Engineering, The University of Auckland, Auckland, New Zealand

It gives me a great pleasure to present this special issue on
“Advances in Heat Transfer Engineering” that contains some
selected papers that were presented at the 4th International Conference on Cooling and Heating Technologies, held in Jinhae,
Korea, during October 28–31, 2008. The conference was hosted
in the “green” and environmentally friendly Jinhae City of Korea, with Professor Hanshik Chung as the chairperson and local
host.
The conference provided an excellent platform for researchers from over 10 countries to present more than 80 papers covering a range of topics on “heat transfer engineering”
leading to sustainable environment. The conference specifically
emphasized the need for international cooperation on the global
warming issues leading to innovations in low carbon industry
and environmentally sustainable development.
This special issue of Heat Transfer Engineering includes
seven articles that cover a number of topics, ranging from uncovering the physics of frost formation on a flat plate to subcooled
flow boiling of CO2 at low temperatures.
The first article is by Shinhyuk Yoon, Gaku Hayase, and
Keumnam Cho. This article presents the details of an experimental apparatus that was used to collect novel data on the frost
formation on a flat plate, and correlations that were developed
for the local and average frost thickness, frost density, and frost

mass.
The second article is by Di Wu, Zhen Wang, Gui Lu, and
Xiaofeng Peng from the Department of Thermal Engineering,
Tsinghua University (China). The article introduced a new idea
to design high-performance air-cooling condensers to automatAddress correspondence to Professor Pradeep Bansal, Department of Mechanical Engineering, The University of Auckland, Private Bag–92019, Auckland, New Zealand. E-mail:

ically separate liquid from gas and to let condensation to occur
in the droplet and unsteady thin film mode, resulting in a high
average heat transfer coefficient.
The third contribution is by Xiaofeng Peng, Chen Fang, and
Fen Wang from the Department of Thermal Engineering, Tsinghua University (China). The article presents mathematical
treatment for better understanding of the vapor bubble transport
in two-phase flow in bead-packed structures.
Gyu-Jin Shim, M. M. A. Sarker, Choon-Geun Moon, HoSaeng Lee, and Jung-In Yoon, in the fourth article present experimental performance of a closed wet cooling tower (CWCT)
with multiple paths having a rated capacity of 9 kW. The study
concluded that a CWCT operating with two paths has higher
heat and mass transfer coefficients than that with single path.
The fifth article in this group is from Yifu Zhang, Weizhong
Li, and Shenglin Quan from the School of Energy and Power
Engineering, Dalian University of Technology (China). The
article presents a numerical method using a combination of
the level-set approach and finite-volume framework to simulate
two-dimensional laminar incompressible two-phase flows. The
method leads to the fluid properties (such as density, viscosity,
etc.) being smoothed as continuous properties.
Yong Yang, Shengqiang Shen, Taewoo Kong, and Kun
Zhang, also from the School of Energy and Power Engineering, Dalian University of Technology (China), in the sixth article describe a two-dimensional compressible numerical model
to evaluate steam properties by the Virial equation. The article studies the difference between condensation shock and
aerodynamic shock, and the influence of aerodynamic shock on
the nonequilibrium phase change is revealed.

The final article in this volume is from Xiumin Zhao and
Pradeep Bansal from the Department of Mechanical Engineering of the University of Auckland (New Zealand). This article

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presents an experimental investigation on the subcooled flow
boiling heat transfer characteristics of CO2 in a horizontal tube
below –30◦ C. The article develops a new empirical correlation
that agrees to within ±30% with the current CO2 experimental
data. It is expected that the data presented in this study would be
beneficial to industry and designers of compact heat exchangers
for CO2 at low temperatures.
I am extremely thankful to the conference organizers, specifically Professor Hanshik Chung for inviting me as a keynote
speaker and providing me the opportunity to be involved in
this conference, and to these authors, who worked diligently in
meeting the review schedule and responding to reviewers’ comments in a timely manner. My special thanks to all the reviewers,
who have done an excellent job in improving the quality of the
papers. Finally, I am also thankful to Professor Afshin Ghajar
for allowing me to publish this special volume of Heat Transfer
Engineering.

heat transfer engineering

Pradeep Bansal holds a personal chair in the Department of Mechanical Engineering at the University of
Auckland (New Zealand). Currently he is also serving

as the Postgraduate Associate Dean in the Faculty of
Engineering, and the Director of the Energy & Fuels
Research Unit at the University of Auckland. He is
a fellow of the American Society of Heating, Refrigerating, and Air-Conditioning Engineers (ASHRAE)
and of the Institute of Refrigerating, Heating, and AirConditioning Engineers (IRHRAE) of New Zealand.
He is also the chair of an ASHRAE Technical Committee (TC10.4) on “Ultra
low cryogenic temperatures,” as well as a member of its Handbook Committee, and a member of various other committees, including TC8.02, TC8.08,
TC8.09, TC10.6, TC10.7, and TC8.09. He serves on numerous national and
international committees, has collaborated with various international institutions, has supervised more than 50 graduate student theses, and has published
more than 200 technical papers, including 3 books. His research domain comprises fundamental heat transfer studies on natural refrigerants, development
of simulation models, and design and development of energy-efficient thermal
systems.

vol. 31 no. 12 2010


Heat Transfer Engineering, 31(12):965–972, 2010
Copyright C Taylor and Francis Group, LLC
ISSN: 0145-7632 print / 1521-0537 online
DOI: 10.1080/01457631003638911

Measurements of Frost Thickness
and Frost Mass on a Flat Plate
under Heat Pump Condition
SHINHYUK YOON,1 GAKU HAYASE,2 and KEUMNAM CHO3
1

Graduate School, Sungkyunkwan University, Suwon, Korea
System Appliances Division, Samsung Electronics Co., Ltd, Suwon, Korea
3

School of Mechanical Engineering, Sungkyunkwan University, Suwon, Korea
2

This study measured the frost thickness and frost mass on a flat plate to propose the correlation equations for the local
and average frost thickness, frost density, and frost mass. Key parameters were the cooling surface temperature of the flat
plate from 258.2 to 268.2 K, absolute humidity of air from 2.98 to 4.16 g/kgDA , air temperature from 273.5 to 280.2 K, and
air velocity from 1.0 to 2.5 m/s. A 50% ethylene glycol aqueous solution was used as a coolant. The sensitivity analysis of
the parameters such as air temperature, air humidity, air velocity, and surface temperature on the frost thickness and frost
mass were experimentally investigated under the heat pump condition. Correlation equations for the local and average frost
thickness and frost mass under the heat pump condition were proposed. The values predicted by the correlation equations
under the freezer condition were larger by a maximum of 30–50% than the values predicted by the present correlation
equations under the heat pump condition. The proposed correlation equations might be applied to the part of the freezer
condition.

INTRODUCTION
The use of air-source heat pumps for residential applications
has steadily increased. It has an advantage of using affluent heat
sources from the surrounding atmosphere. When the air temperature in winter is below the freezing temperature of water,
porous frost begins to form. The frost layer on the evaporator of
the heat pump acts as a resistance to heat transfer and reduces
air flow rate. Frost thickness, frost density, etc. are required to
be investigated to understand frost formation. Even though the
finned-tube evaporator for the heat pump mostly uses louvered
fins and slit fins instead of plate fins, the fin might be simplified
as a flat plate. There are lots of studies on frost on a flat plate in
the open literature. Most of them reported the frost pattern under
freezer conditions instead of heat pump conditions. Frost formations under heat pump conditions might be different from those
under freezer conditions due to different frost properties, even
This work was supported by SFARC at Sungkyunkwan University, and
Brain Korea 21 Project in Korea. The authors appreciate Samsung Electronics

Co. for providing test samples and advice.
Address correspondence to Professor Keumnam Cho, School of Mechanical Engineering, Sungkyunkwan University, 300 Chunchun-dong, Jangan-gu,
Suwon 440-746, Korea. E-mail:

though they show similar trends. Most of the following studies
are based on the freezer condition. Trammel et al. [1] studied
the frost layer on the flat plate. They found that the frost density
increases as the dew point and air velocity increase. Brian et al.
[2] provided measured frost density graphically. Frost density
was increased as the air temperature was increased. Sanders [3]
reported that the frost density was increased as wall and air
temperatures were increased. They also reported that higher air
humidity makes lower frost density. O’Neal and Tree [4] found
that the frost density increases as time passes, due to vapor diffusion. Na and Webb [5] investigated fundamental phenomena
related to frost deposition and growth. They found that water
vapor pressure at the frost surface is supersaturated, by applying laminar concentration boundary layer analysis. A couple of
other studies [6–8] modeled the frost formation process employing a semi-empirical transient model for a flat plate under forced
convection condition. Mao et al. [9, 10], Yang and Lee [11], and
Lee and Ro [12] proposed their own correlation equations for
the frost density and the frost thickness. None of them were
developed by considering heat pump conditions.
There are few studies under heat pump conditions so far.
Kwon et al. [13] investigated the frost formation on a flat
plate with local cooling. Our previous study, Shin et al. [14],

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reported that the pressure drop through the slit-fin-and-tube heat
exchanger under frosting condition at low velocity was higher
than that at high velocity, although the average frost thickness
at low velocity was less than that at high velocity. Most of frost
studies were performed by using the average frost properties,
even though they are locally different.
The objective of the present study is to suggest the correlation
equations for the local and average frost thickness and frost mass
on the flat plate under heat pump conditions.

EXPERIMENTAL APPARATUS
The schematic diagram of the experimental apparatus is
shown in Figure 1a. It consists of a psychrometric calorimeter, a refrigerant system, a wind tunnel, a data acquisition system, and a test section. The psychrometric calorimeter, which
had temperature control range of −10 to 50◦ C and humidity
control range of 30 to 95%, provided constant dry- and wetbulb temperature by using an air handling unit. The refrigerant
system used an ethylene glycol–water mixture for easy control
of the inlet temperature of refrigerant. Bypass solenoid valves
at both inlet and outlet of the test section made the refrigerant not flow into the test section prior to the test. The mass
flow rate of refrigerant was measured by a Coriolis mass flow
meter with an accuracy of ±0.1% of the full scale. The inlet
and outlet temperatures of refrigerant were measured by the
RTD with an accuracy of ±0.15◦ C. In the wind tunnel, air was
supplied by a 2.2-kW exhaust fan and four nozzles that had different diameters. To homogenize the air flow, honeycombs were
installed at the inlet and outlet of the test section. Insulating material was placed around the test apparatus to minimize the heat
loss. The wind-tunnel section designed was made of transparent
acryl with a width of 300 mm, height of 100 mm, and length of
1000 mm. The test section with a width of 200 mm and length of
150 mm was made of copper, and it was flush mounted at the

center of the bottom of the acryl wind tunnel.
Frost mass was measured by using aluminum tape and a
balance. A piece of thin aluminum tape was used to cover the
surface of the flat plate before each test, as shown in Figure 1b.
Frost mass was determined by measuring the change of the aluminum tape before and after the test. Frost mass was measured
every 30 min by repeating the frost mass measurements at the
same test condition, since it was very difficult to measure instantaneous frost mass. Four different frost masses were measured
every 30 min under the same condition since the test period
was 2 h. The frost surface temperatures on the flat plate were
measured by an infrared thermometer and T-type thermocouples. Figure 1b shows the positions measured the frost surface
temperature.
Frost thickness was measured by a digital CCD camera positioned properly by a stepping motor, as shown in Figure 2a.
The CCD camera took pictures of the frost every 10 min automatically. The flat plate was flush mounted at the bottom cenheat transfer engineering

Figure 1 Experimental apparatus.

ter of the acryl duct to avoid any edge effect of the duct. Most
commercial three-dimensional scanners are very expensive, and
they have a resolution of the order of 5 mm, which is not good
enough to measure the frost thickness. Since the frost profile
was almost symmetrical on the left- and right-hand sides along
the flow direction, a two-dimensional frost profile was utilized
instead of a three-dimensional profile. Frost thickness profiles
were measured at four different positions as shown in Figure 2b
to verify the two-dimensional frost profile. The frost thickness
was defined as shown in Figure 2c. Figure 3 showed the typical
frost thickness profiles at four different positions. Frost thickness profiles at the left and right sides showed similar patterns
with almost the same values, while the frost thicknesses at front
and rear sides were almost constant except for a small part of the
front and rear edges. This means that the frost thickness profile

might be determined by monitoring only the left-hand-side frost
thickness.
Two dry- and wet-bulb thermometers were installed before
and after the test section to measure the average temperature
and humidity of the moist air. The uncertainty of the air temperature measurement was 0.4◦ C, while the uncertainty of the
relative humidity was 1%. The refrigeration system consisted
of a refrigerator and a pump to circulate the refrigerant. A 50%
ethylene glycol aqueous solution was used as the refrigerant.
Flow rate of the refrigerant was set to 1 kg/min. The test data
were recorded every 2 s for 120 min.
Key experimental parameters were cooling surface temperature (Tw ), air humidity (wa ), air temperature (Ta ), and air velocity (Va ). They ranged from 258.2 to 268 K for the cooling
vol. 31 no. 12 2010


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967

Figure 4 Comparison of measured and estimated frost masses.

where δ f,m is the measured frost thickness, W is the length of
the flat plate, and ρ f,e is the estimated frost density by using Eq.
(2) suggested by Hayashi et al. [15]. In general, high values of
density are expected as the frost surface temperature approaches
the water triple point, and a curve like the one depicted as the exponential function by Hayashi et al. [15] is a good representation
of expected results. That’s why it is used for the comparison.
ρ f,e = 650 exp(0.277T f,m )
Figure 2 Frost thickness measurement.

surface temperature, from 2.98 to 4.16 g/kgDA for the air humidity, from 273.5 to 280.2 K for the air temperature, and from

1.0 to 2.5 m/s for the air velocity.
DATA REDUCTION
The measured frost mass was compared with the estimated
one calculated by using Eq. (1) to verify the validity of the
methodology of the frost thickness and frost temperature measurements.
m f,e =

ρ f,e · δ f,m · W d x

(1)

Figure 3 Typical frost profiles at four different positions (wa = 3.67 g/kgDA ,
Tw = 263.15 K, Ta = 275.15 K, Va = 1.5 m/s, t = 120 min).

heat transfer engineering

(2)

The measured average frost density might be determined by
using Eq. (3).
m f,m
(3)
ρ f,m =
Aal δ f,m
The uncertainty was ±1.2% for the frost thickness and
±8.2% for the frost mass through the uncertainty analysis suggested by Moffat [16].

RESULTS AND DISCUSSION
The measured frost mass was compared with the estimated
one to verify the methodology of the frost thickness and frost

temperature measurements as shown in Figure 4. The estimated
and measured frost masses agreed within 8%, which is within
the uncertainty range. This means that the methodology utilized
for the frost thickness measurement is appropriate.
The effect of the cooling surface temperature on the local
frost thickness at a position of 75 mm from the entrance and the
frost mass are shown in Figure 5. Both the local frost thickness
and frost mass were increased as the cooling surface temperature
was decreased. The local frost thicknesses for a cooling surface
temperature of 258.2 K were larger by 33.5% than those for a
cooling surface temperature of 263.2 K and by 63.3% than those
for a cooling surface temperature of 268.2 K. The frost masses
for a cooling surface temperature of 258.2 K were larger by 5.3%
than those for a cooling surface temperature of 263.2 K and by
13.6% than those for a cooling surface temperature of 268.2 K.
The cooling surface temperature affected more severely the local
frost thickness than the frost mass. The reason is as follows. As
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Figure 5 Effect of cooling surface temperature.

Figure 6 Effect of air humidity.

the cooling surface temperature decreases, heat from the phase
change process of a water molecule may be easily absorbed into

the frost layer, and then the surface temperature of the frost layer
is maintained at a low level. This reduces the humidity of the
boundary between the surface of the frost layer and the air, and
thus maintains a large concentration driving force. As a result, a
larger amount of frost is produced. However, it is supposed that
the lower temperature of the frost surface causes the formation
of small droplets or particles of water molecule, consequently
resulting in a coarse frost later, and then the structure made
during early crystal growth period affects the growth of the frost
layer.
Figure 6 shows the effect of the air humidity on the local frost
thickness and frost mass. The local frost thicknesses at a position of 75 mm from the entrance for a humidity of 4.16g/kgDA
were larger by 21.4% than those for a humidity of 3.67g/kgDA
and 52.3% than those for a humidity of 2.98g/kgDA . The frost
masses for a humidity of 4.16g/kgDA were larger by 16.4%
than those for a humidity of 3.67g/kgDA and 115.3% than those
for a humidity of 2.98g/kgDA . The effect of humidity on the
frost mass was almost the same order with the effect of cooling
surface temperature. This might be mainly because the high humidity causes a high concentration driving force that transports
a greater amount of water vapor from the air to the frost layer.

Figure 7 shows the effect of the air temperature on the local
frost thickness and frost mass. Even though air temperature was
ascertained to have a small effect compared to air humidity
and cooling surface temperature, an influence was nevertheless
found. The local frost thicknesses at the position of 75 mm from
the entrance for an air temperature of 273.5 K were larger by
1.2% than those for an air temperature of 275.2 K and by 8.5%
than those for an air temperature of 280.2 K. However, the frost
masses for an air temperature of 273.5 K were smaller by 5.2%

than those for an air temperature of 275.2 K and by 12.6%
than those for an air temperature of 280.2 K. The structure
of the frost layer constructed in the early crystal growth period
might play a role resulting in a large frost mass. During the early
crystal growth period, higher air-side surface temperatures of the
frost layer decrease the probability of small droplets or particles
formation from water vapor, and then cause a thinner and dense
frost layer. Increase of the frost density, which means a decrease
of the porosity, provides the larger specific surface area and then
causes the water vapor on the frost surface to diffuse easily into
the inner frost layer like a pumping effect.
Figure 8 shows the effects of air velocity on local frost thickness and frost mass. They are comparably smaller than the effects of the air humidity and the cooling surface temperature.
The local frost thicknesses at the position of 75 mm from the

heat transfer engineering

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969

Figure 7 Effect of air temperature.
Figure 8 Effect of air velocity.

entrance for an air velocity of 2.5 m/s were larger by 4.0% than
those for an air velocity of 1.5 m/s and by 6.1% than those for an
air velocity of 1.0 m/s. This might be because higher air velocity
results in a larger quantity of frost layer, slightly increased frost

layer thickness, and accelerated densification of the layer.
Most literature reports related to frost for the freezer stated
that the frost under the freezer condition is mainly due to the
high temperature of outside air. Based on the literatures for the
freezer, the freezer condition was set to 15 ≤ Ta (◦ C) ≤ 25,
6.94 ≤ wa (g/kgDA ) ≤ 12.50, and −25 ≤ Tw (◦ C) ≤ −15. A
few papers expanded the range to 5 ≤ T a (◦ C) ≤ 25, 3.58 ≤
wa (g/kgDA ) ≤ 12.50, and −35 ≤ Tw (◦ C) ≤ −15. Frost for the
heat pump is mainly caused by the cold air of outside in winter.
The air temperature and the absolute humidity of air for the heat
pump are lower than those for the freezer, while the cooling
surface temperature for the heat pump is higher than that for
the freezer. The heat pump condition was set to 0 ≤ Ta (◦ C) ≤
7, 2.98 ≤ wa (g/kgDA ) ≤ 4.16, and −15 ≤ Tw (◦ C) ≤ −5. Frost
characteristics for the heat pump might be different from those
for the freezer due to different operating conditions. Figure 9
shows the applicable range of the freezer condition as a dotted
circle and the heat pump condition as a solid circle.
Most literature reports suggested average values for the frost
thickness and frost mass instead of local values. Correlation
equations for the local frost thickness and frost density are proheat transfer engineering

posed as shown in Eqs. (4) and (5) by using measured local
data (ρ f,m and δ f,m ) and modifying the empirical equations
suggested by Yang and Lee [11]:
δ f, p = 3.782(L ∗ )−1.352 (wa )1.704 (Fo)0.6803 (T ∗ )2.035 (Re L )0.251
(4)

Figure 9 Applicable ranges of the proposed correlations.

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Figure 10 Local frost thickness and frost density.
Figure 11 Comparison of the measured average frost thickness and frost mass
with the values predicted by some correlations under heat pump conditions.

ρ f, p = 1.852 × 10−4 (ρice )(L ∗ )−0.976 (wa )2.312
×(Fo)

0.550

∗ 3.035

(T )

(Re L )

0.346

Figure 10 shows the local frost thicknesses and frost densities
measured and predicted. The local frost thickness and local frost
density get smaller along the air flow direction. Heat and mass
transfer at the leading edge is relatively brisk because of the
leading edge effect. The average values predicted by Yang and
Lee [11] were larger by 16 to 58% than the measured data,

since they predicted the values under the freezer condition. The
correlation equations for local frost thickness and frost density
under the heat pump condition in the present study might predict
much more accurately than the other correlation equations.
The average frost thickness and frost density might be expressed as Eqs. (6) and (7) by taking the average of local frost
thickness and frost density shown in Eqs. (4) and (5).
δ¯ f, p =

L
0

ρ¯ f, p =

(5)

Tt p − Tw
αa t
x
ρ Va L
, L∗ = , Fo = 2 , Re L = a
T∗ =
Ta − Tw
L
L
µa

δ f, p d x
(6)
L
heat transfer engineering

L
0

ρ f, p d x
L

(7)

The frost mass might be also estimated as shown in Eq. (8)
by using Eqs. (4) and (5) for the local values.
L

m¯ f, p =

ρ f, p δ f, p W d x

(8)

0

Figure 11 shows a comparison of the measured average frost
thicknesses and frost mass data with the values predicted by the
correlation equations developed under freezer condition. The
predicted values by the correlation equations under the freezer
condition suggested by Mao et al. [10], Yang and Lee [11],
and Lee and Ro [12], including data by Serker et al. [17], were
compared with the predicted values by the present correlation
equations (6) and (8). The comparison was done for condition
1 of Figure 9. The values predicted by the correlation equations under the freezer condition were larger by a maximum

of 30–50% than the values predicted by the present correlation
equations under the heat pump condition. This is mainly due
to the differences in the conditions such as air humidity, air
temperature, and surface temperature. The freezer condition is
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S. YOON ET AL.

971

on the frost thickness and frost mass were experimentally
investigated under the heat pump condition.
3. Correlation equations for the local and average frost thickness and frost mass under the heat pump condition were
proposed.
4. The values predicted by the correlation equations under the
freezer condition were larger by a maximum of 30–50%
than the values predicted by the present correlation equations
under the heat pump condition.
5. The proposed correlation equations might be applied to part
of the freezer condition.
NOMENCLATURE
area, m2
Fourier number ( = αa t/L2)
length of the flat plate, m
local frost mass, g
average frost mass, g
temperature, ◦ C
time, min
velocity, m/s

width of the flat plate, m
absolute humidity, kg/kg DA
position from the entrance of the flat plate, mm

A
Fo
L
m
m¯
T
t
V
W
w
x

Figure 12 Comparison of the measured average frost thickness and frost mass
with the values predicted by some correlations under freezer conditions.

usually expected to have more frost than the heat pump condition. Existing correlation equations under the freezer condition
including data overpredict by 30 to 50% the frost density and
frost mass under the heat pump condition.
Correlation equations suggested in the present study might
be extended to the freezer condition. This was examined by
comparing the predicted values by the same correlation equations utilized in Figure 11 and data by Serker et al. [17] with
the predicted values by the present correlation equations (6) and
(8) at condition 2 of Figure 9. Figure 12 shows the comparison.
The predicted values by the present correlation equations agreed
with the data by Serker et al. [17] within a maximum of 10%.
This means that the proposed correlation Eqs. (4) and (8) might

be applied to the part of the freezer condition.

CONCLUSIONS

Greek Symbols
α
δ
δ¯
µ
ρ
ρ¯

thermal diffusivity, m2/s
local frost thickness, mm
average frost thickness, mm
viscosity, N-s/m2
local frost density, kg/m3
average frost density, kg/m3
relative humidity,%

Subscripts
a
al
e
f
ice
m
p
tp
w

air
aluminum tape
estimated value
frost
ice condition at Tf
measured value
predicted value
triple point of water
cooling surface

The present study can be summarized as follows.
REFERENCES
1. The estimated and measured frost masses agreed within an
uncertainty range of 8%.
2. The sensitivity analysis of the parameters such as air temperature, air humidity, air velocity, and surface temperature
heat transfer engineering

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of Frost Formed on a Flat Plate Held at Sub-Freezing Temperature, ASHRAE Journal, vol. 7, no. 10, pp. 42–47, 2004.
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[13] Kwon, J. T., Lim, H. J., Kwon, Y. C., Koyama, S., Kim,
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Cooling, International Journal of Refrigeration, vol. 29,
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heat transfer engineering

Under Frosting Condition, Proc. Winter Annual Conference of SAREK (Seoul, Korea), pp. 252–257, 2007.
[15] Hayashi, Y., Aoki, K., and Yuhara, H., Study of Frost Formation Based on a Theoretical Model of the Frost Layer,
Heat Transfer-Japan Research, vol. 6, no. 3, pp. 79–94,
1977.
[16] Moffat, R. J., Using Uncertainty Analysis in the Planning
of an Experiment, Trans. ASME: Journal of Fluid Engineering, vol. 107, pp. 173–182, 1985.
[17] Serker, D., Karatas, H., and Egrican, N., Frost Formation
on Fin-And-Tube Heat Exchanger. Part I—Modeling of
Frost Formation on Fin-And-Tube Heat Exchangers, International Journal of Refrigeration, vol. 27, no. 4, pp.
367–374, 2004.
Shinhyuk Yoon is an M.S. degree student at
Sungkyunkwan University, Suwon, Korea, under the
supervision of Prof. Keumnam Cho. He received in
2008 a Diploma of Mechanical Engineering from
Sungkyunkwan University, Suwon, Korea. He is currently working on frost formation in compact heat
exchangers of heat pumps.

Gaku Hayase is a principal engineer at Samsung
Electronics Co., Suwon, Korea, and a Ph.D. degree
student at Kyushu University, Fukuoka, Korea, under
the supervision of Prof. Yasuyuki Takata. He received
his M.S. degree from the Tottori University in Japan.
He has been working at the Matsushita Refrigeration

Company since 1992 and at Mitsubishi Electronics
Co. since 1998. He is currently working on heat and
fluid dynamics, and compact heat exchangers of heat
pumps.
Keumnam Cho is a professor of mechanical engineering at Sungkyunkwan University, Suwon, Korea.
He received his M.S. and Ph.D. degrees from the
State University of New York at Stony Brook. He
has been teaching at Sungkyunkwan University since
1993. He has been the editor of the IJR (International
Journal of Refrigeration) since 2007 and has been a
vice-president of the IIR E1 Commission (International Institute of Refrigeration) and a vice-president
of SAREK (Society of Air-Conditioning and Refrigeration Engineers of Korea) since 2008. He is currently working on compression
and absorption refrigeration systems, and compact heat exchangers.

vol. 31 no. 12 2010


Heat Transfer Engineering, 31(12):973–980, 2010
Copyright C Taylor and Francis Group, LLC
ISSN: 0145-7632 print / 1521-0537 online
DOI: 10.1080/01457631003638952

High-Performance Air Cooling
Condenser With Liquid–Vapor
Separation
DI WU, ZHEN WANG, GUI LU, and XIAOFENG PENG
Department of Thermal Engineering, Tsinghua University, Beijing, China

In this investigation, an innovative idea was introduced to design a new kind of high-performance air cooling condensers.
This kind of condenser functions to automatically separate liquid from gas and makes condensation always occur in droplet

and unsteady thin film condensation mode everywhere in the whole condenser, which results in very high average heat transfer
coefficient. An introduction is presented to describe the basic principle and structure of the novel heat exchanger technology,
particularly the understanding from the fundamental experimental investigations. Furthermore, a series of experiments was
conducted to validate the high performance of an air-conditioning system with an innovative condenser. Though heat transfer
area of the condenser was about 37% less, the performance was as good as or even better than that of the original one.

INTRODUCTION
The air cooling condenser is an important kind of heat exchanger because of its applicability to a variety of engineering
equipment and processes, such as power engineering, chemical
processes, and air conditioning. Air cooling condensers have attracted wide attention, especially in terms of deficiency of water
resources and environmental deterioration. However, the disadvantages of an air cooling condenser, such as low overall heat
transfer coefficient, huge volume, and large pressure drop, limit
its extent of application. In air-conditioning systems, three major methods are usually employed to enhance the heat transfer
of air cooling condensers: enhancing air-side and tube side heat
transfer coefficient or overall heat transfer coefficient, increasing the heat transfer area, and augmenting mean temperature
difference [1].
Normally, enhanced tubes and high-performance fins or increasing heat transfer area is employed to significantly improve
the performance of condensers. However, in this way both manufacturing cost and power consumption are considerably increased. So far, this kind of heat transfer enhancement almost
reaches its upper limit due to very few benefits, or more material consumption and high manufacturing cost, while increasing

This research is currently supported by the National High Technological
Development Program (“863” Program) through contract 2007AA05Z200.
Address correspondence to Di Wu, Department of Thermal Engineering,
Tsinghua University, Beijing 100084, China. E-mail:

mean temperature difference is greatly restricted by the operating conditions [2].
A traditional condenser cooled by air is completely condensed in one or several flow routes with complicated two-phase
flow evolution. The condensate accumulates to form a thick film
on the tube surface and complicated two-phase flow [3], leading
to obvious decrease of the heat transfer coefficient and great

change of the tube wall temperature along the flow direction.
So, enhanced tubes [4–6], high-performance fins, and optimal
flow arrangements [7–9] are frequently used to improve the
condenser performance.
Recently, an innovative idea and technology were proposed
to design a new kind of high-performance condensers [10–14]
and heat exchangers [15]. In this paper, an attempt is made to describe and apply this idea and technology. This kind of condenser
would automatically separate liquid from the vapor–liquid twophase mixture and makes condensation always occur in droplet
and/or unsteady thin film condensation mode throughout the
whole condenser, resulting in very high average condensation
heat transfer coefficient. A series of experiments was conducted
to test an air conditioning system with an innovative condenser.
DESIGN PRINCIPLES
Condensation Inside a Tube
Normally, a condenser has relatively high heat transfer performance due to huge latent heat release, and it can maintain

973


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D. WU ET AL.

Figure 3 Flow regimes during condensation in horizontal tube.

Figure 1 Condensation modes: (a) thin film mode; (b) droplet mode.

relatively homogeneous wall temperature. When wall temperature is less than the saturation temperature of a vapor, condensation will occur on the wall. Generally, there are two modes
of condensation, droplet and thin film mode, as clearly shown
in Figure 1, due to different wall surface and condensed liquid conditions. For thin film condensation, a liquid thin film

forms between hot vapor and cool wall surface and the liquid
can successfully spread on the wall surface, which prevents vapor directly contacting wall surface and produces the main heat
transfer resistance. For droplet condensation, liquid is hard to
spread on the wall and there are large parts of the bare wall surface that hot vapor can directly contact. As a result, for specified
conditions, heat transfer performance of droplet condensation
is higher than that of thin film condensation. However, it is
more difficult to realize and maintain droplet condensation in
applications than people expect.
Figure 2 schematically shows a typical structure of a traditional air cooling condenser in domestic air-conditioning systems. The refrigerant flows into the condenser through two parallel routes, A and B. In each close serpentine tube route A or
B, the vapor phase condenses into liquid phase without phase
separation, experiencing an entire and complex two-phase flow
evolution, normally including single vapor phase flow, annular
flow, slug flow, plug flow, bubbly flow, and single liquid phase,
as shown in Figure 3.

Figure 2 Structure of a traditional air cooling condenser.

heat transfer engineering

During condensation, liquid film forms on the wall, which
prevents direct contact of refrigerant vapor with the cool wall
surface and produces main heat transfer resistance of the condensation. Consequently, heat transfer performance rapidly becomes worse as the liquid film becomes thicker and thicker, with
finally very complex two-phase flow with less and less vapor. For
film condensation, the local liquid film thickness, δx , and average
heat transfer coefficient of whole tube length, hx , can be approximately obtained from the Nusselt theoretical solution of condensation inside a horizontal tube for laminar liquid film flow as [1]
δx =

4ηl λl (Ts − Tw ) x
gρ2lr

h x = 0.943

gr ρ2l λ2l
µl x (Ts − Tw )

1/4

(1)

1/4

(2)

where, µl , λl , ρl , and r represent viscosity, thermal conductivity,
density of refrigerant liquid, and latent heat, respectively; T s
and T w are saturated temperature corresponding to operating
pressure and wall temperature, and x denotes tube length.
Figure 4 shows the predictions of Eqs. (1) and (2) for R22
at 1.95 MPa and T s –T w = 1.5 K. Obviously, the liquid film
thickness significantly increases during condensation, which
increases heat transfer resistance and decreases corresponding
heat transfer coefficient rapidly. Apparently, for a traditional
design in Figure 2 without liquid–vapor phase separation, the
whole condensation heat transfer is very weak in most regions
downstream of tube routes A and B where more and more condensed liquid is attached on the wall surface or accumulated into
a two-phase flow with much more liquid. Consequently, more
heat transfer area is required and the volume of the whole condenser is enlarged. Also, the pressure drop would be very high
and oscillate unsteadily due to the complicated two-phase flow,
which highly influences the stability and safety of the system.
Unlike that in the downstream region of tube route A or B,

the condensation mode is expected to be droplet condensation
or extremely thin film turbulent condensation (condensed liquid was fully disturbed by vapor and hard-to-maintain smooth
film) existing in the upstream zones of the tubes or zones very
close to the inlets [16], due to refrigerant vapor directly contacting the tube wall and keeping a relatively high velocity in this
region. As a result, condensation performance would be very
good. If condensation mode in the whole tube side can be maintained the same as that in the entrance region, such as the first
one or two straight tubes of route A or B (the marked entrance
region in Figure 4), the heat transfer will be enhanced enormously. One possible technology is to separate liquid phase
from liquid–vapor two-phase flow in time to reduce liquid
vol. 31 no. 12 2010


D. WU ET AL.

975

0.10

Entrance
Region

Liquid film thickness / mm

0.09
0.08
0.07
0.06
0.05
0.04
0.03

0.02
0.01
0.00
0.0

Figure 5 Test module.

0.5

1.0

1.5

2.0

For a condenser, Q, η, T f1 , and T f2 all are predetermined.
If the condensation heat transfer coefficient inside the tube, h1 ,
increases, (T f1 –Tw ) and/or inside tube area, A1 , will decrease
according to Eq. (3), and accordingly Tw and(Tw —T f2 ) increase.
As a result, the outside fin area, A2 , will decrease, which implies
that heat transfer performance of the condenser is enhanced. It is
concluded that if condensation heat transfer is greatly enhanced,
wall temperature will increase and temperature difference at two
sides will be redistributed, which enhances total performance
and reduce the heat transfer area.

x/m

4.5

Entrance
Region

-2

Heat transfer coefficient / kW m K

-1

(a) Liquid film thickness

4.0
3.5
3.0
2.5

Experimental Evidence

2.0
1.5
0.0

0.5

1.0

1.5

2.0

x/m
(b) Heat transfer coefficient
Figure 4 Condense film and heat transfer coefficient evolution: (a) liquid film
thickness; (b) heat transfer coefficient.

accumulation and keep unsteady thin liquid film or droplets
on the wall by innovative structure design. Meanwhile, vapor
velocity is nearly maintained at the same value as at the route
entrance in order to disturb the thin liquid film or even achieve
droplet condensation.

Heat Transfer Enhancement
Considering a typical convective heat transfer process in horizontal tubes with outside fins and neglecting resistance of thin
wall conduction, the total heat transfer for inside tube and air
side is expressed as
Q = h 1 A1 (Tf1 − Tω ) = h 2 A2 η(Tω − Tf2 )

(3)

where, h1 and h2 are the heat transfer coefficients for inside and
outside tube, respectively; A1 and A2 denote the areas of inside
tube and outside fins; η is fin efficiency; and T f1 and T f2 are
inside vapor and outside air temperature, respectively.
heat transfer engineering

To validate the condensation mode right in the entrance region of a tube, a series of fundamental experiments was conducted. The test module is schematically shown in Figure 5.
Vapor from the evaporator entered into a test channel with rectangle cross section, 10 mm in width and 12 mm in height.
The bottom of the test channel was a copper plate, 12 mm in
width, 72 mm in length, and 2 mm in thickness, and the other
three sides were covered by quartz plates. A cooling circle was

equipped under the copper plate to exhaust the heat transferred
from upper vapor condensation. In total, five T-type thermocouples embedded in copper plate were equipped to portray the
temperature evolution during vapor condensation on the copper
plate. The uncertainty of the temperature measurement was less
than 0.1 K. Meanwhile, a high-speed CCD camera was utilized
to capture the whole dynamic phenomena.
The inlet vapor Reynolds number is calculated as
Rev =

ρv u v de
µv

(4)

where de is the hydraulic diameter, defined as
4f
(5)
U
where U and f are wetted perimeter and cross-section area of
the test channel, respectively. The condensation heat transfer
coefficient is
qm r
(6)
hc =
Ac (Ts − Tw )
de =

vol. 31 no. 12 2010

976

D. WU ET AL.

Figure 6 Condensation behavior: (a) Rev = 1670, (b) Rev = 2600, (c) Rev =
5600, (d) Rev = 8960.

where qm represents mass flow rate of the condensed liquid, and
r is latent heat. Ac denotes the copper plate area, and T s and T w
are vapor saturated temperature and inner wall surface temperature (obtained using a one-dimensional [1-D] conduction model
from the averaged value of five thermocouples measuring outside wall temperatures), respectively.
Figure 6 schematically depicts the typical condensation
phenomena occurred on the tested copper plate. The arrows
represent flow direction of vapor. At the initial stage of condensation, droplet condensation always emerged. During condensation, droplets gradually grew and experienced coalescence, and
later large droplets formed and spread on the wall surface. At
a low Rev in Figure 6a, since the shear force induced by vapor
had little effect on the liquid film, a steady thin film formed
and covered the whole copper plate at t = 50 s without any
disturbance and rupture, and the film became thicker along the
channel as shown in Figure 3. At a relatively higher Rev , due
to the strong shear force of vapor, a thin film first formed very
close to the entrance at t = 17s in Figure 6b and t = 10s in Figure
6c, and then spread to downstream part of the plate at t = 22s in
Figure 6b and t = 14.6s in Figure 6c. However, at these moments, the liquid film was unstable and thickness was decreased
due to relatively high shear force and ruptured periodically.
Meanwhile, droplets emerged periodically in the downstream
zone. For this case, condensation of droplet and thin film turbulent mode coexisted, and the heat transfer was enhanced. When
vapor velocity increased further, as shown in Figure 6d, even
more unsteady brook-like condensation occurred, unlike the different droplet or film condensation mode. The liquid film was
highly disturbed by high-velocity vapor. The film was extremely

thin and ruptured periodically, which was expected to greatly
enhance heat transfer performance.
Temperature evolutions with different vapor inlet velocities
are depicted in Figure 7. At a low inlet velocity, droplets emerged
initially, and finally a thin film formed and was maintained, as
shown in Figure 6a. Consequently, wall temperature temporarheat transfer engineering

ily experienced a high value and decreased to a low value in
Figure 7a, as condensation varied from droplet to thin film.
As Rev increased from 3600 to 5600 in Figures 7b and 7c,
temperature experienced fluctuation at a relatively high value
and finally reached a steady value. This evolution characteristic corresponds to condensation behavior in Figures 6b and
6c. As vapor velocity increased, steady thin film condensation
gradually converted to periodical thin film turbulent and droplet
condensation.
The heat transfer coefficient is plotted as a function of Rev in
Figure 8, significantly increasing with Rev , which has the same
variation trend as the work of Annaiev et al. [17] and is expected to be induced by condensation-mode evolution. As Rev
increased, the stable thin film became instable and the thickness
decreased gradually due to the enhanced vapor shear force. As
a result, periodically fluctuating thin film and droplet condensation emerged and coexisted, which greatly enhanced heat transfer performance. Due to the increase of the condensation heat
transfer coefficient, the wall temperature accordingly increased
as shown in Figure 7, which is consistent with the analysis in
the previous section, and heat transfer performance is expected
to enlarge. As the value of hc changed from 10 kW/m2-K to 15.7
kW/m2-K, corresponding to changes in Rev from 2600 to 3600,
the increase gradient was relatively higher due to condensation
transition from steady film condensation to coexistence of unsteady thin turbulent film and droplet condensation, consistent
with that in Figures 6 and 7. Apparently, vapor velocity would
play an important role in enhancing the condensation heat transfer. It is expected that the condensation mode and corresponding

heat transfer characteristics observed in a rectangular channel
would be similar to that in a circular tube.
Novel Idea
The idea for enhancing the performance of an air cooling
condenser is illustrated in Figure 9. A short tube only containing the entrance effect of condensation is employed everywhere
in the whole condenser, and a liquid–vapor separator is introduced to drain the condensed liquid between two adjacent short
tubes. This idea can ensure that the condensation in the whole
condenser is in the expected modes, or droplet and unstable thin
film condensation mode.

STRUCTURE DESIGN
For traditional condensers, Rev is high enough to reach the
transition point from steady film condensation to coexistence
of unsteady thin turbulent and droplet condensation. However, since there is no liquid–vapor separation and more and
more condensed liquid accumulates in tubes into a two-phase
flow, the heat transfer will be greatly weakened. On the opposite, for an innovated condenser, condensed liquid is separated
by liquid–vapor separators in time and pure vapor enters the
vol. 31 no. 12 2010


D. WU ET AL.

977

Figure 8 Condensation heat transfer coefficient.

Figure 7 Wall surface temperature evolution: (a) Rev = 1670, (b) Rev = 2600,
(c) Rev = 5600.

succeeding tubes, or the condensation is kept as an unsteady

thin liquid film and/or droplet mode on the wall. If the reasonable structure is designed to remain with invariant high vapor
velocity everywhere in the condenser, the condensation heat
transfer would be further enhanced.
heat transfer engineering

An innovative condenser based on the proposed novel idea
was re-manufactured from the original condenser of an air
conditioner with refrigerating capacity of 2300 W, shown in
Figure 10. The coiled tube of the original one was composed
of 16 straight tubes and 15 return (or U) bends. One passage
tube length of the original one is about 690 mm. For the innovative design, the straight tube length was cut to less than
430 mm, as shown in Figure 9a, reduced 37.7% compared
to the original one, and other structure and sizes were not
changed. The flow passed was divided into two zones, condensation and condensate subcooled zone. In the condensation
zone the straight tubes were connected using two manifolds
at two ends. In these two manifolds several liquid–vapor separators were included to separate vapor from the two-phase
mixture there and form several flow passages in the condensation zone. This ensures that pure vapor enters the next flow
passage, so the high-performance condensation modes can be
reached in the tubes. Keeping an almost invariant vapor velocity
at all inlets of straight tubes is another important technology
in order to disturb the thin liquid film or even achieve droplet
condensation mode. This requires each flow passage to have different straight tube numbers, referring to Figure 10b. The downstream serpentine subcooled zone was almost similar to original
condenser.
During operation, refrigerant vapor would first enter the condensation zone. Different from the original one, the condensation zone has left and right manifolds rather than return bends at
the two sides of straight tubes. After condensation in one tube
pass, condensed liquid separates from the liquid–vapor mixture
automatically and flows into the downstream serpentine tube.
Consequently, a single vapor phase enters without condensed
liquid, and droplet condensation or thin liquid condensation is
expected to occur in the succeeding flow passage, similar to

Figure 9 Novel idea for condenser design.

vol. 31 no. 12 2010


978

D. WU ET AL.

Figure 11 Experimental system.

Figure 10 Improvement of a practical condenser: (a) structure of original
condenser, (b) structure of innovative design, and (c) original and improved
condenser.

the phenomena in Figures 6 and 9. The heat transfer is greatly
enhanced and high condensation performance is achieved in the
entire condensation zone. After condensation, condensed liquid
goes into the downstream serpentine zone, consisting of several
straight tubes and is subcooled to a required value.

TEST EXPERIMENTS
A series of experiments was conducted to validate the high
performance of an air-conditioning system with an innovative
condenser. The experimental system is schematically illustrated
heat transfer engineering

in Figure 11. Utilizing electrical heaters and humidifiers, two
separated rooms provided the standard indoor (dry-bulb temperature 27◦ C, wet-bulb temperature 19◦ C) and outdoor conditions

(dry-bulb temperature 35◦ C, wet-bulb temperature 24◦ C) for
indoor and outdoor unit, respectively. Measuring dry-bulb and
wet bulb-temperature of the air, before and after flowing through
the evaporator, the enthalpy difference was calculated. Together
with measurement of air flow rate, refrigerating capacity was determined from the air side. After monitoring compressor work,
the value of the energy efficiency ratio (EER), equal to the ratio
of refrigerating capacity to compressor work, was obtained.
Both the refrigerant charge and the length of capillary are
very important parameters that influence the performance of an
air conditioner. In each experiment, both capillary length and
refrigerant charge were optimized to obtain the highest EER and
refrigerating capacity. Table 1 lists the experimental results of an
air conditioning system with original and innovative condenser,
respectively.
Utilizing liquid–vapor separators to separate condensed liquid without any leak of refrigerant vapor and appropriate tube
distribution to keep inlet velocity high and nearly invariant for
each vapor flow passage, the vapor phase is expected to condense in the mode of droplet or unsteady thin liquid condensation occurring in short straight tubes. Apparently, heat transfer
capacity of the condenser was greatly enhanced, about 165 W,
or from 3127 W to 3292 W in Table 1, though the heat transfer
area had about 37% decrease. Meanwhile, refrigerating capacity also had a significant increase from 2300 W to 2400 W.
On the other hand, this innovative design got rid of complicated
Table 1

Experimental results of two systems

Item
Refrigerant charge (g)
Air flow rate from evaporator
(m3/h)
Refrigerating capacity (W)

EER
Condenser cooling capacity
(W)

vol. 31 no. 12 2010

Innovative
condenser

Original
condenser

472
366

700
400

2400
2.69
3292

2300
2.78
3127


D. WU ET AL.

two-phase flow patterns and consequently had low pressure drop

and relatively steady operation conditions, though the pressure
drop was not measured in this investigation. Less refrigerant,
about 32% decrease, was filled. However, EER had a small decrease from 2.78 to 2.69. This might be mainly due to several
reasons. The air flow rate from the evaporator was 366 m3/h,
8.5% less than the rating value of 400 m3/h. It is expected that
when air flow rate reaches 400 m3/h, not only EER will reach
2.78 or more, but the capacity of condenser and corresponding
refrigerating capacity of system will be enlarged much more.
Also, the original system, which was designed for a traditional
condenser, might not match with this new condenser very well.
Very clearly, both heat transfer performance of the condenser
and corresponding refrigerating capacity of the system were significantly improved by means of liquid–vapor separation technology, together with reasonable design of tube route distribution in condensation zone. In the present case, heat transfer area
can be significantly reduced, almost 37% or even more, compared to original one. This reduction is expected for any other
air cooling condensers.

979

tical applications, and particularly can be extended from airconditioning systems to other kinds of air cooling condensers,
such as that utilized in power plants, chemical engineering, and
so on.

NOMENCLATURE
A
de
f
g
hc
hx
qm
Q

r
Re
T
u
U
x

area, m2
hydraulic diameter, m
cross section area of test channel, m2
gravity, m s−2
heat transfer coefficient of condensation, W m−2 K−1
average heat transfer coefficient, W m−2 K−1
mass flow rate of condensed liquid, kg s−1
heat transfer rate, W
latent heat, J kg−1
Reynolds number, dimensionless
temperature, K
velocity, m.s−1
wetted perimeter, m
tube length, m

CONCLUSIONS
In the present investigation, an innovative idea and technology were introduced to design a new kind of high-performance
air cooling condensers. This kind of condenser would automatically separate liquid from the vapor–water mixture or two-phase
flow in time by the innovative structure design. The key technology of structure design includes liquid–vapor separator structure and the tube number distribution of each flow passage in
condensation zone. Consequently, the single vapor phase enters without any condensed liquid, and droplet condensation or
thin liquid condensation is expected to occur. The heat transfer is greatly enhanced, and high condensation performance
is achieved in the entire condensation zone. Furthermore, heat
transfer enhancement was analyzed and a series of fundamental

experiments was conducted to validate the condensation mode
right in the entrance region of a tube. It is concluded that heat
transfer coefficient increased significantly with increase of Rev ,
which is expected to be induced by condensation mode transition from steady film condensation to coexistence of unsteady
thin turbulent and droplet condensation.
A series of experiments was conducted to validate the high
performance of an air-conditioning system with an innovative
condenser. Very clearly, both heat transfer performance of the
condenser and corresponding refrigerating capacity of the system were significantly improved by means of liquid–vapor separation technology, together with reasonable design of the tube
route distribution in the condensation zone. In the present case,
heat transfer area can be significantly reduced, almost 37% or
even more, compared to the original one. It is expected that
the enhanced heat transfer principle and corresponding innovative design idea in present work have great potential for pracheat transfer engineering

Greek Symbols
δx
η
λ
µ
ρ

local liquid film thickness, m
fin efficiency, dimensionless
thermal conductivity, W m−1 K−1
viscosity, Pa s
density, kg m−3

Subscripts
1
2

c
f1
f2
l
s
v
w

inside tube
fin side
copper plate
fluid 1 (vapor)
fluid 2 (air)
refrigerant liquid
saturated state
vapor
Wall

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heat transfer engineering

[17] Annaiev, E. P., Boyko, L. D., and Kruzhilin, G. N., Heat
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Di Wu received a B.E. degree in mechanical engineering from Tsinghua University, Beijing, China in
2006. He is a Ph.D. student in the Department of Thermal Engineering, Tsinghua University. His current
interests include transport phenomena in porous media and microscale heat transfer (with/without phase
change).

Zhen Wang received a B.E. degree in mechanical engineering from Tsinghua University, Beijing, China
in 2005. He is a Ph.D. student in the Department
of Thermal Engineering, Tsinghua University. His
current interests include heat transfer and bubble behavior in the nucleate boiling of binary mixtures and
evaporation of binary mixtures.

Gui Lu received a B.E. degree in mechanical engineering from Tsinghua University, Beijing, China in
2006. He is a master’s candidate in the Department
of Thermal Engineering, University of Science and

Technology. His current interests include transport
phenomena in thin liquid films and proton exchange
membrane fuel cells.

Xiaofeng Peng received B.E. (1983) and D.E. (1987)
degrees in thermal engineering from Tsinghua University, Beijing, China. He is a professor in the Department of Thermal Engineering, Tsinghua University. His research interests include micro heat transfer, heat transfer in porous media, and phase-change
transport phenomena.

vol. 31 no. 12 2010


Heat Transfer Engineering, 31(12):992–997, 2010
Copyright C Taylor and Francis Group, LLC
ISSN: 0145-7632 print / 1521-0537 online
DOI: 10.1080/01457631003638994

Performance Characteristics of a
Closed-Circuit Cooling Tower With
Multiple Paths
GYU-JIN SHIM,1 M. M. A. SARKER,2 CHOON-GEUN MOON,3
HO-SAENG LEE,4 and JUNG-IN YOON5
1

Department of Refrigeration and Air-Conditioning Engineering, Pukyong National University, Pusan, South Korea
Department of Mathematics, Bangladesh University of Engineering and Technology, Dhaka, Bangladesh
3
Daeil Co., Ltd., Pusan, South Korea
4
Pukyong National University, Pusan, South Korea
5

School of Mechanical Engineering, Pukyong National University, Pusan, South Korea
2

The performance of a closed-circuit wet cooling tower (CWCT) with multiple paths having a rated capacity of 9 kW has been
studied experimentally. When the CWCT has to operate with a partial load, the required quantity of cooling water reduces
and thereby the velocity of the process fluid inside the tubes decreases. The velocity of the process fluid can be increased by
installing blocking tubes in the heat exchanger. The test section in this experiment has multiple paths that have been used as
the inlet for cooling water that flows from the top part of the heat exchanger. The heat exchanger consists of eight rows and
12 columns and the tubes are in a staggered arrangement. Heat and mass transfer coefficients and temperature drops were
calculated with several variations including multiple paths. The results obtained from this study were compared with those
reported and found to conform well. The investigation indicates that a CWCT operating with two paths has higher heat and
mass transfer coefficients than with one path.

INTRODUCTION
Cooling towers are used frequently to reject heat from an
industrial system or process without thermally polluting surface
water. In general, cooling towers are classified into open and
closed type. Open cooling towers expose the water directly to
the atmosphere and transfer source heat load directly to the air,
causing the air pollution. The other type, called closed-circuit
cooling towers, which maintain an indirect contact between the
fluid and the atmosphere, are being used increasingly due to the
nonpollution of cooling water or air [1–3].
Several authors conducted experimental studies of closedcircuit cooling towers and proposed correlations of heat and
mass transfer coefficients as a function of tube diameter and
designed conditions [4–7].
In a closed-circuit wet cooling tower (CWCT), cooling water
and spray water circulation pumps and fans are the prime factors responsible for power consumption. A cooling water pump
Address correspondence to Professor Jung-In Yoon, School of Mechanical Engineering, Pukyong National University, Pusan, South Korea. E-mail:

consumes 60–70% of total power consumption [8, 9]. Due to energy consumption concerns, there is a need to design the CWCT
in such a way that the requirement of the quantity of cooling
water can be reduced. In an efficient design, one can curtail expenditure incurred by tubes by reducing the quantity of cooling
water. In that case, the cooling capacity could be a bit low if it
is applied to typical CWCT with one path, because the velocity
of process fluid in the tubes decreases. To increase the velocity
of the process fluid in the tube, blocking tubes can be installed
in the heat exchanger in a multiple-path system. Figure 1 shows
the concept of multiple paths.
In the relevant literature, no results have been reported so
far involving the CWCT with multiple paths. In this scenario,
the objective of this article is to obtain basic data from this
experimental study on a small-size CWCT with multiple paths
and analyze the performance characteristics.
EXPERIMENTAL APPARATUS AND METHOD
A schematic diagram of the experimental apparatus used in
this study is shown in Figure 2. In the experiment, the prototype

992


G.-J. SHIM ET AL.

993

Table 1 Specifications of heat exchanger and experimental conditions
Tube diameter
Transverse pitch
Row pitch

Dimension
Cooling water
Spray water
Air

19.05 [mm]
38.1 [mm]
46 [mm]
W 0.304, L 0.6, H 0.525 [m]
Flow rate
Inlet temperature
Flow rate
Velocity
Inlet wet-bulb temperature

600–3120 [kg/h]
32–50 [◦ C]
720–2160 [kg/h]
1.0–3.5 [m/s]
23–29 [◦ C]

Figure 1 Concept of multiple paths.

CWCT is used where the tube section is located at the upper
part, and fans are installed at the lower part. In the tube section,
the tubes, spray system, eliminator, and the other peripherals
connecting the parts are sequentially organized and are kept in a
casing. The copper tube with an outer diameter of 19.05 mm is
used in the W(0.304 m) × L(0.6 m) × H(0.525 m) dimensional
tower and in a staggered arrangement. Water pressure nozzles

are used to distribute the spray water over the tube bundles,
and air is circulated counterflow by a sirocco fan. The fan’s
motor is equipped with a variable speed control to change air
velocity. The T-type thermocouple temperature sensor with a
diameter of 0.3 mm is used while measuring the temperature of
cooling water and air at the inlet and the outlet of the CWCT.
The humidity sensor is used to measure the inlet and outlet air
humidity of the CWCT. The humidity and temperature at five
points in the air inlet and the outlet are measured at every 5 s
and the averages of these values are applied.
Design conditions are a cooling capacity of 9 kW, for a cooling water inlet temperature of 37◦ C, a cooling water flow rate of
1560 kg/h, air wet-bulb temperature of 27◦ C, and air velocity of
3 m/s. The cooling water is supplied by pipes and the pipes are
connected to the distribution head through eight horizontal cooling tubes. The cooling water flows downward from the top. The
cooling water after coming out through the outlet of the CWCT
is sent to the constant-temperature tank. The cooling water gains

heat and gets stabilized to a certain temperature while passing
though the constant-temperature tank. Then it recirculates to the
CWCT. The spray water is uniformly distributed at the upper
part of the coils by pump and circulates in the tower. The lower
water tank section consists of spray-water collecting tank and
ambient-air forcing fans. Ambient air constraints were maintained to the required state with the help of cooler, air heater,
and humidifier. Table 1 gives the experimental conditions and
the tower geometry. Under the experimental conditions given
in Table 1, the experiment was conducted with changing the
flow rate and inlet temperature of cooling water, flow rate of
spray water, and wet-bulb temperature and velocity of inlet air.
After running the experiment for a while, the temperature of
spray water and cooling water get stabilized and experimental

data are read after all conditions are normalized. The data were
recorded with the help of an automatic data logger (MX100),
and all the readings at all inlets and outlets were collected after
the experiment stabilized in a steady state. The duration of the
test run can be no less than 1 h. The uncertainties of the measured and calculated parameters are estimated by following the
procedures described in ASME PTC-23 [10] (with a level of
confidence of 95%). The experimental uncertainties are associated with measurement devices and sensors. The specifications
of measuring devices are shown in Table 2. The method is based
on a combining of all uncertainties primary experimental measurements. The uncertainty values are 1.14% and 2.45% for the
water flow rate and total systematic and random uncertainty,
respectively.

THEORETICAL BACKGROUND
Heat transfers from a hot process fluid inside tubes to spray
water and to air through a water film. Heat transfers from spray
Table 2 Measuring devices specifications
Parameter

Sensor

Water temperature
Water flow rate
Relative humidity
Air temperature
Air velocity

T type
Dwyer, series VFB
Capacitive
T type

Anemometer

Figure 2 Schematic diagram of the experimental apparatus.

heat transfer engineering

vol. 31 no. 12 2010

Range

Accuracy

–200 to 400 [◦ C]
0–70 [L/m]
0–100 [%]
–200 to 400 [◦ C]
0–20 [m/s]

±0.1[◦ C]
±1[%]
±1[%]
±0.1[◦ C]
±0.03 [m/s]


994

G.-J. SHIM ET AL.

water to air in latent and sensible forms. The rate of heat lost by

cooling water is given by:
dqc = m c c pc dtc = Uo (tc − ts )d A

Do
Di

+

Do
ln
2λ

Do
Di

+

1
ho

(2)

To calculate heat transfer coefficient for water inside tubes, the
correlation of Nu that was proposed by Gnielinski [11] for fully
developed turbulent flow was utilized:
Nu =

( f /8)(Re − 1000) Pr[1 + (Di /l)0.67 ]
1 + 12.7( f /8)0.5 (Pr0.67 −1)

Heat capacity of the air [kW]

where Uo is the overall heat transfer coefficient based on the
outer area of the tube. To calculate Uo , the following equation
was used:
1
1
=
Uo
hi

14

(1)
12
+15%
10
-15%
8

6

(3)

The rate of heat gain by air is:
dqa = m a di a = k(i − i a )d A

4

(4)

4

The mass transfer coefficient can be obtained using mass balance:

8

10

12

14

Heat capacity of the cooling water [kW]
Figure 3 Heat balance of the experimental apparatus.

(5)

Here, dxLM is the logarithmic mean humidity difference, defined
as:
xa,2 − xa,1
d xL M =
(6)
x −x
ln x3 −xa,1
a,2
3

More details about the underlying theory can be found
[4–6].

RESULTS AND DISCUSSIONS
To check the reliability of the experimental apparatus using
the heat and mass transfer balance, Eqs. (1) and (4) are used.
The results have been shown in Figure 3 where the heat balance
data that have fallen within ±15% were used. The heat balance
of the apparatus could be claimed to be satisfactory.
Figures 4 and 5 show the mass transfer coefficient, k, as a
function of air velocity and flow rate of spray water per unit
breadth, , in the CWCT. In the Figure 4, mass transfer coefficients are compared to the values of the correlations by Parker
and Treybal [4] and Nitsu et al. [5]. In the case of the CWCT
using one path, mass transfer coefficients are similar to the
correlation of Nitsu et al. [5]. This means that there is a high reliability for the experimental apparatus. It is observed that mass
transfer coefficients that were calculated for the CWCT having
one path and two paths increased with the increase of the air
velocity. This is mainly because the measured temperatures of
spray water at the surface of tubes in the outlet of the CWCT
using two paths are higher than the other, causing an increase in
the absolute humidity at the outlet of CWCT using two paths.
Mass transfer coefficients having two paths are approximately
heat transfer engineering

43% and 17% higher than those having one path when air velocities were 1 m/s and 3.5 m/s, respectively. In Figure 5, in the
case of the CWCT using two paths, mass transfer coefficients
are also higher than the other with regard to the flow rate of
spray water per unit breadth.
Figures 6 and 7 show the heat transfer coefficient, ho , versus
flow rate of spray water per unit breadth and air velocity. In both
figures, heat transfer coefficients found for the CWCT using one
path can be claimed to be highly similar to the correlation given

by Nitsu et al. [5]. Furthermore, heat transfer coefficients for two
paths are similar to the correlation by Parker and Treybal [4].
This indicates that heat transfer coefficients increase with the
increase of flow rate of spray water per unit breadth but show
hardly any increase with respect to air velocity. Heat transfer
1.0

Mass transfer coefficient [kg/m 2s]

m a (xa,2 − xa,1 ) = k Ad x L M

6

0.8

mc : 1,560 kg/h

19.05mm, One path
19.05mm, Two paths
Parker and Treybal [4]
Nitsu et al. [5]

ms : 1,080 kg/h

0.6

Nitsu et al
0.4
Parker and Treybal
0.2

0.0
1

2

3

Air velocity [m/s]
Figure 4 Mass transfer coefficient k as a function of air velocity.

vol. 31 no. 12 2010

4


G.-J. SHIM ET AL.
14

1.2
mc : 1,560 kg/h

19.05mm, One path
19.05mm, Two paths
Parker and Treybal [4]
Nitsu et al. [5]

1.0

12

v : 3 m/s

19.05mm, One path
19.05mm, Two paths

10

0.8
o

Range [ C]

Mass transfer coefficient [kg/m 2s]

995

0.6
Nitsu et al.
0.4
Parker and Treybal

8

6

mc : 1,560 kg/h
ms : 1,080 kg/h

4

0.2

v : 3 m/s

2
0.0
0.008

0.010

0.012

0.014

0.016

0.018

0.020

0.022

0.024

0
30

Flow rate of spray water per unit breadth [kg/ms]

40

45

50
o

Figure 5 Mass transfer coefficient k as a function of flow rate of spray water
per unit breadth.

Heat transfer coefficient [W/m 2K]

35

Cooling water temperature [ C]
Figure 8 Temperature range as a function of cooling-water temperature.

coefficients in the CWCT using two paths are higher than those
in a CWCT with one path in both figures.

3000

mc : 1,560 kg/h
v : 3 m/s

19.05mm, One path
19.05mm, Two paths
Parker and Treybal [4]
Nitsu et al. [5]

2500

2000

1500
Parker and Treybal
Nitsu et al.
1000

500

0
0.008

0.010

0.012

0.014

0.016

0.018

0.020

0.022

0.024

Flow rate of spray water per unit breadth [kg/ms]
Figure 6 Heat transfer coefficient ho as a function of flow rate of spray water
per unit breadth.

Temperature range (drop) with respect to a variable coolingwater inlet temperature (CWIT) and wet-bulb temperature
(WBT) are shown in Figures 8 and 9. It is evident that range
increases almost linearly with the increasing temperature of
cooling water. At the standard design condition, ranges that
were measured in the CWCT using one and two paths are 4.2◦ C
and 5.1◦ C, respectively. The range in the CWIT using two paths
is approximately 20% higher than that with one path. From the
temperature range against variable inlet wet-bulb temperature, it
is clear that the range decreases with the increase of the wet-bulb
temperature. This is because when the wet-bulb temperature at
the inlet increases, the temperature difference between the inlet
cooling water and air decreases. Thus, the vaporization of the
spray water outside the pipes decreases so that the falling of
the temperature of the cooling water flowing inside the tubes

7
2500

19.05mm, One path
19.05mm, Two paths

mc : 1,560 kg/h

19.05mm, One path
19.05mm, Two paths
Parker and Treybal [4]

Nitsu et al. [5]

ms : 1,080 kg/h

6

2000

o

Range [ C]

Heat transfer coefficient [W/m 2K]

3000

1500
Parker and Treybal
Nitsu et al.

1000

5

4

mc : 1,560 kg/h
ms : 1,080 kg/h
v : 3 m/s

500

3
0
1

2

3

4

Air velocity [m/s]
Figure 7 Heat transfer coefficient ho as a function of air velocity.

heat transfer engineering

23

24

25

26

27

28

29

30

o

Wet-bulb Temperature [ C]
Figure 9 Temperature range as a function of inlet wet-bulb temperature.

vol. 31 no. 12 2010


996

G.-J. SHIM ET AL.

CONCLUSIONS

Cooling capacity [kW]

12
19.05mm, One path
19.05mm, Two paths

10

8

6

mc : 1,560 kg/h

ms : 1,080 kg/h

4

2
1

2

3

4

Air velocity [m/s]

The fundamental study of the performance characteristics
of the closed-circuit wet cooling tower with multiple paths has
been done experimentally with a rated capacity of 9 kW. The
results can be summarized as follows:
Heat and mass transfer coefficient of the CWCT using one
path was found to conform well to the already reported results
for almost all cases considered.
In the optimum level, mass transfer coefficients for variable
air velocity and spray water flow rate of the CWCT having two
paths are respectively about 43% and 28% higher than those
having one path.
The temperature drop of the cooling water for the CWIT
having two paths is nearly 20% higher than that with one path.

Figure 10 0 Cooling capacity as a function of air velocity.

NOMENCLATURE

decreases. Temperature drops of the CWIT using two paths are
higher than for those with one path at all cases. This is mainly
because the heat transfer coefficient for water inside tubes of
the CWCT using two paths is almost two times higher than the
other.
Figures 10 and 11 show cooling capacity as a function of
air velocity and cooling-water flow rate. The cooling capacity
of the CWCT having two paths is remarkably higher than that
with one path in both cases, due to the significant enhancement
of heat transfer coefficient for cooling water inside tubes in
the germane cases. In Figure 11, it is indicated that cooling
capacities of the CWCT having one and two paths are increasing
until cooling water flow rate is 32 L/min and converging to
one point, respectively. This is mainly because the rate of heat
transfer between the inside and outside of the tubes no longer
increases after cooling-water flow rate is 32 L/min.

14
19.05mm, One path
19.05mm, Two paths

Cooling capacity [kW]

12

A
cp

D
f
v
H
h
hi
ho
i
k
L
l
m
Nu
Pr
q
Re
t
Uo
W
x

10

area, m2
specific heat, J kg−1 K−1
tube diameter, m
friction factor, dimensionless
air velocity, m s−1
height, m
convective heat transfer coefficient, W m−2 K−1

heat transfer coefficient for water inside the tubes, W m−2
K−1
heat transfer coefficient between tube external surface and
spray water film, W m−2 K−1
enthalpy, J kg−1
mass transfer coefficient, kg m−2 s−1
length of heat exchanger, m
tube length, m
mass flow rate, kg s−1
Nusselt number, dimensionless
Prandtl number, dimensionless
rate of heat transfer, W
Reynolds number, dimensionless
temperature, ◦ C
overall heat transfer coefficient based on the outer area of
the tube, W m−2 K−1
width, m
absolute humidity, kg kg DA −1

Greek Symbols

8

6

λ

ms : 1,080 kg/h
v : 3 m/s

thermal conductivity,W m−1 k−1
flow rate of spray water per unit breadth, kg m−1 s−1

4

Subscripts
2
0

10

20

30

40

50

60

Cooling water flow rate [l/min]
Figure 11 Cooling capacity as a function of cooling-water flow rate.

heat transfer engineering

a
c
i

air
cooling water
inside
vol. 31 no. 12 2010


G.-J. SHIM ET AL.

o
s
1
2
3

outside
spray water
at tower inlet
at tower outlet
Interface (spray water film/air)

997
Gyu-Jin Shim is an M.S student in the Department
of Refrigeration and Air-Conditioning at Pukyong
National University, Pusan, South Korea. He received
his bachelor’s degree in 2006 from Pukyong National
University. He received a best award paper from the
Korean Society of Heat & Cold Energy Engineers in
2007. He is currently working on the characteristics
of a closed circuit cooling tower with multiple paths.

Superscripts
saturated condition
REFERENCES
[1] Bedekar, S. V., Nithiarasu, P., and Seetharamuz, K. N., Experimental Investigation of the Performance of a CounterFlow, Packed-Bed Mechanical Cooling Tower, Energy,
vol. 23 pp. 943–947, 1998.
[2] Hasan, A., and Siren, K., Theoretical Analysis of Closed
Wet Cooling Towers and Its Applications in Cooling of
Buildings, Energy and Buildings, vol. 34, pp. 477–486,
2002.
[3] Stabat, P., and Marchio, D., Simplified for IndirectedContact Evaporative Cooling-Tower Behaviour, Applied
Energy, vol. 78, pp. 433–451, 2004.
[4] Parker, R. O., and Treybal, R. E., The Heat, Mass Transfer Characteristics of Evaporative Coolers, Chemical Engineering Progress Symposium Series, Buffalo, NY, pp.
138–149, 1961.
[5] Nitsu, Y., Naito, K., and Anzai, T., Studies of the Characteristics and Design Procedure of Evaporative Coolers,
Journal of the Society of Heating, Air-Conditioning, Sanitary Engineers of Japan, vol. 41, no. 12, vol. 43, no. 7, pp.
10–15, 1969.
[6] Mizushina, T., Ito, R., and Miyashita, H., Characteristics
and Methods of Thermal Design of Evaporative Cooler,
International Chemical Engineering, vol. 8, pp. 532–538,
1968.
[7] Facao, J., and Oliveira, A., Heat and Mass Transfer Correlations for the Design of Small Indirect Contact Cooling
Towers, Applied Thermal Engineering, vol. 24, no. 14–15,
pp. 1969–1978, 2004.
[8] Sarker, M. M. A., Kim, E., Moon, C. G., and Yoon, J. I.,
Performance Characteristics of the Hybrid Closed Circuit
Cooling Tower, Energy and Building, vol. 40, no. 8, pp.
1529–1535, 2008.
[9] Shim, G. J., Sarker, M. M. A., Baek, S. M., Lee, H. S.,
Kim, E. P., and Yoon, J. I., Experimental Study of Closed
Wet Cooling Tower With Multi Path, Proc. 4th BSMEASME International Conference on Thermal Engineering,

Dhaka, pp. 482–487, 2008.
[10] ASME PTC-23. Atmospheric Water Cooling Equipment,
American Society of Mechanical Engineers, New York,
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[11] Gnielinski, V., Zur W¨arm¨ubertragung bei laminarer
Rohrstr¨omung und konstanter Wandtemperatur, Chem.
Ing. Tech, vol. 61, no. 2, pp. 161–163, 1989.
heat transfer engineering

M. M. A. Sarker is an associate professor of mathematics at Bangladesh University of Engineering and
Technology, Dhaka, Bangladesh. He received his
M.Sc. (applied mathematics) degree from the University of Dhaka, and an advanced studies in master of statistics degree from Katholieke Uviversiteit
Leuven, Belgium. He completed a Ph.D. from Pukyong National University, Pusan, South Korea. He has
been teaching at BUET since 1994. His research contributions were in the field of refrigeration and airconditioning engineering. He is currently working on the numerical aspect of
enhancement of cooling capacity in hybrid closed-circuit cooling towers.

Choon-Geun Moon is a research engineer at DAEIL
Co., Ltd., Pusan, South Korea. He received his M.Sc.
degree from Pukyong National University and his
Ph.D. in refrigeration and air-conditioning in 2004
from Pukyong National University. He was previously in charge of a laboratory involved in characteristics of heat and mass transfer on absorption refrigerator and desiccant dehumidifier. He was a researcher
on liquid desiccant dehumidifiers at the University
of Auckland for about 2 years (2006–2007). He is
currently working on the performance of inverter cooled chillers.

Ho-Saeng Lee is a research engineer at Pukyong National University, Pusan, South Korea. He received
his M.Sc. degree from Pukyong National University,
and his Ph.D. in refrigeration and air-conditioning
in 2006 from Pukyong National University. He was
previously in charge of a laboratory involved in performance characteristics of refrigeration systems using hydrocarbon refrigerants. He was a researcher on

characteristics of oil properties from discharge refrigerants in compressors at the University of Illinois at
Urbana–Champaign in 2007. He is currently working on the development of the
LNG refrigeration cycle.

Jung-In Yoon is a professor at the School of Mechanical Engineering at Pukyong National University, Pusan, South Korea. He received his M.Sc degree from Pukyong National University and his Ph.D.
degree in 1995 from Tokyo University of Agriculture & Technology, Japan. He has been teaching at
Pukyong National University since 1995 except for
one year spent at the Thermodynamic Laboratory of
the University of Auckland in New Zealand. His research contributions are in the field of refrigeration
and air-conditioning engineering. He is currently working on the development
of the LNG refrigeration cycle, performance of inverter industrial cooled chillers
on high-accuracy temperature control, life-cycle cost evaluation of absorption
chillers, and design of a shell-an-tube heat exchanger program.

vol. 31 no. 12 2010