E-mail: p.bansal@auckland.ac.nz ically separate liquid from gas and to let condensation to occur in the droplet and unsteady thin film mode, resulting in a highaverage heat transfer coef
Trang 2CopyrightC Taylor and Francis Group, LLC
ISSN: 0145-7632 print / 1521-0537 online
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
Con-ference 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
Ko-rea, with Professor Hanshik Chung as the chairperson and local
host
The conference provided an excellent platform for
re-searchers from over 10 countries to present more than 80
pa-pers 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
uncov-ering the physics of frost formation on a flat plate to subcooled
flow boiling of CO2at low temperatures
The first article is by Shinhyuk Yoon, Gaku Hayase, and
Keumnam Cho This article presents the details of an
experi-mental 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
automat-Address correspondence to Professor Pradeep Bansal, Department of
Me-chanical Engineering, The University of Auckland, Private Bag–92019,
Auck-land, New Zealand E-mail: p.bansal@auckland.ac.nz
ically separate liquid from gas and to let condensation to occur
in the droplet and unsteady thin film mode, resulting in a highaverage heat transfer coefficient
The third contribution is by Xiaofeng Peng, Chen Fang, andFen Wang from the Department of Thermal Engineering, Ts-inghua University (China) The article presents mathematicaltreatment 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, Saeng Lee, and Jung-In Yoon, in the fourth article present ex-perimental performance of a closed wet cooling tower (CWCT)with multiple paths having a rated capacity of 9 kW The studyconcluded that a CWCT operating with two paths has higherheat and mass transfer coefficients than that with single path.The fifth article in this group is from Yifu Zhang, Weizhong
Ho-Li, and Shenglin Quan from the School of Energy and PowerEngineering, Dalian University of Technology (China) Thearticle presents a numerical method using a combination ofthe level-set approach and finite-volume framework to simulatetwo-dimensional laminar incompressible two-phase flows Themethod leads to the fluid properties (such as density, viscosity,etc.) being smoothed as continuous properties
Yong Yang, Shengqiang Shen, Taewoo Kong, and KunZhang, also from the School of Energy and Power Engineer-ing, Dalian University of Technology (China), in the sixth arti-cle describe a two-dimensional compressible numerical model
to evaluate steam properties by the Virial equation The ticle studies the difference between condensation shock andaerodynamic shock, and the influence of aerodynamic shock onthe nonequilibrium phase change is revealed
ar-The final article in this volume is from Xiumin Zhao andPradeep Bansal from the Department of Mechanical Engineer-ing of the University of Auckland (New Zealand) This article
963
Trang 3presents an experimental investigation on the subcooled flow
boiling heat transfer characteristics of CO2in a horizontal tube
below –30◦C The article develops a new empirical correlation
that agrees to within±30% with the current CO2experimental
data It is expected that the data presented in this study would be
beneficial to industry and designers of compact heat exchangers
for CO2at low temperatures
I am extremely thankful to the conference organizers,
specif-ically 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’
com-ments 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.
Pradeep Bansal holds a personal chair in the
Depart-ment 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, erating, and Air-Conditioning Engineers (ASHRAE) and of the Institute of Refrigerating, Heating, and Air- Conditioning Engineers (IRHRAE) of New Zealand.
Refrig-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 Commit- tee, 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 institu- tions, has supervised more than 50 graduate student theses, and has published more than 200 technical papers, including 3 books His research domain com- prises fundamental heat transfer studies on natural refrigerants, development
of simulation models, and design and development of energy-efficient thermal systems.
Trang 4CopyrightC 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,1GAKU HAYASE,2and KEUMNAM CHO3
1Graduate School, Sungkyunkwan University, Suwon, Korea
2System Appliances Division, Samsung Electronics Co., Ltd, Suwon, Korea
3School of Mechanical Engineering, Sungkyunkwan University, Suwon, Korea
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/kg DA , 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
tem-perature in winter is below the freezing temtem-perature 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
forma-tions under heat pump condiforma-tions 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
Mechan-ical Engineering, Sungkyunkwan University, 300 Chunchun-dong, Jangan-gu,
Suwon 440-746, Korea E-mail: keumnamcho@skku.edu
though they show similar trends Most of the following studiesare based on the freezer condition Trammel et al [1] studiedthe frost layer on the flat plate They found that the frost densityincreases as the dew point and air velocity increase Brian et al.[2] provided measured frost density graphically Frost densitywas increased as the air temperature was increased Sanders [3]reported that the frost density was increased as wall and airtemperatures were increased They also reported that higher airhumidity makes lower frost density O’Neal and Tree [4] foundthat the frost density increases as time passes, due to vapor dif-fusion Na and Webb [5] investigated fundamental phenomenarelated to frost deposition and growth They found that watervapor pressure at the frost surface is supersaturated, by apply-ing laminar concentration boundary layer analysis A couple ofother studies [6–8] modeled the frost formation process employ-ing a semi-empirical transient model for a flat plate under forcedconvection condition Mao et al [9, 10], Yang and Lee [11], andLee and Ro [12] proposed their own correlation equations forthe frost density and the frost thickness None of them weredeveloped 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 flatplate with local cooling Our previous study, Shin et al [14],
965
Trang 5reported 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
calorime-ter, a refrigerant system, a wind tunnel, a data acquisition
sys-tem, and a test section The psychrometric calorimeter, which
control range of 30 to 95%, provided constant dry- and
wet-bulb 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
refriger-ant 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
dif-ferent diameters To homogenize the air flow, honeycombs were
installed at the inlet and outlet of the test section Insulating
ma-terial 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
alu-minum 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
instan-taneous 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
thermocou-ples Figure 1b shows the positions measured the frost surface
temperature
Frost thickness was measured by a digital CCD camera
po-sitioned properly by a stepping motor, as shown in Figure 2a
The CCD camera took pictures of the frost every 10 min
auto-matically The flat plate was flush mounted at the bottom
cen-Figure 1 Experimental apparatus.
ter of the acryl duct to avoid any edge effect of the duct Mostcommercial three-dimensional scanners are very expensive, andthey have a resolution of the order of 5 mm, which is not goodenough to measure the frost thickness Since the frost profilewas almost symmetrical on the left- and right-hand sides alongthe flow direction, a two-dimensional frost profile was utilizedinstead of a three-dimensional profile Frost thickness profileswere measured at four different positions as shown in Figure 2b
to verify the two-dimensional frost profile The frost thicknesswas defined as shown in Figure 2c Figure 3 showed the typicalfrost thickness profiles at four different positions Frost thick-ness profiles at the left and right sides showed similar patternswith almost the same values, while the frost thicknesses at frontand rear sides were almost constant except for a small part of thefront and rear edges This means that the frost thickness profilemight be determined by monitoring only the left-hand-side frostthickness
Two dry- and wet-bulb thermometers were installed beforeand after the test section to measure the average temperatureand humidity of the moist air The uncertainty of the air tem-perature measurement was 0.4◦C, while the uncertainty of therelative 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 datawere recorded every 2 s for 120 min
Key experimental parameters were cooling surface ature (Tw), air humidity (wa), air temperature (Ta), and air ve-locity (Va) They ranged from 258.2 to 268 K for the cooling
Trang 6S YOON ET AL 967
Figure 2 Frost thickness measurement.
surface temperature, from 2.98 to 4.16 g/kgDAfor the air
hu-midity, 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
Figure 4 Comparison of measured and estimated frost masses.
where δf ,mis the measured frost thickness, W is the length of
the flat plate, and ρf ,eis the estimated frost density by using Eq.
(2) suggested by Hayashi et al [15] In general, high values ofdensity are expected as the frost surface temperature approachesthe water triple point, and a curve like the one depicted as the ex-ponential function by Hayashi et al [15] is a good representation
of expected results That’s why it is used for the comparison
±8.2% for the frost mass through the uncertainty analysis gested by Moffat [16]
sug-RESULTS AND DISCUSSION
The measured frost mass was compared with the estimatedone to verify the methodology of the frost thickness and frosttemperature measurements as shown in Figure 4 The estimatedand measured frost masses agreed within 8%, which is withinthe uncertainty range This means that the methodology utilizedfor the frost thickness measurement is appropriate
The effect of the cooling surface temperature on the localfrost thickness at a position of 75 mm from the entrance and thefrost mass are shown in Figure 5 Both the local frost thicknessand frost mass were increased as the cooling surface temperaturewas decreased The local frost thicknesses for a cooling surfacetemperature of 258.2 K were larger by 33.5% than those for acooling surface temperature of 263.2 K and by 63.3% than thosefor a cooling surface temperature of 268.2 K The frost massesfor 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 by13.6% than those for a cooling surface temperature of 268.2 K.The cooling surface temperature affected more severely the localfrost thickness than the frost mass The reason is as follows As
Trang 7Figure 5 Effect of cooling surface temperature.
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
posi-tion 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
than those for a humidity of 3.67g/kgDAand 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
hu-midity causes a high concentration driving force that transports
a greater amount of water vapor from the air to the frost layer
Figure 6 Effect of air humidity.
Figure 7 shows the effect of the air temperature on the localfrost thickness and frost mass Even though air temperature wasascertained to have a small effect compared to air humidityand cooling surface temperature, an influence was neverthelessfound The local frost thicknesses at the position of 75 mm fromthe entrance for an air temperature of 273.5 K were larger by1.2% 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 frostmasses 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 periodmight play a role resulting in a large frost mass During the earlycrystal growth period, higher air-side surface temperatures of thefrost layer decrease the probability of small droplets or particlesformation from water vapor, and then cause a thinner and densefrost layer Increase of the frost density, which means a decrease
of the porosity, provides the larger specific surface area and thencauses the water vapor on the frost surface to diffuse easily intothe inner frost layer like a pumping effect
Figure 8 shows the effects of air velocity on local frost ness and frost mass They are comparably smaller than the ef-fects of the air humidity and the cooling surface temperature.The local frost thicknesses at the position of 75 mm from the
Trang 8S YOON ET AL 969
Figure 7 Effect of air temperature.
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,
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
pro-Figure 8 Effect of air velocity.
posed as shown in Eqs (4) and (5) by using measured localdata (ρf ,m and δf ,m) and modifying the empirical equations
suggested by Yang and Lee [11]:
δf ,p = 3.782(L∗ −1.352(w a)1.704 (F o)0.6803 (T∗ 2.035(ReL)0.251
(4)
Figure 9 Applicable ranges of the proposed correlations.
Trang 9Figure 10 Local frost thickness and frost density.
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
ex-pressed as Eqs (6) and (7) by taking the average of local frost
thickness and frost density shown in Eqs (4) and (5)
The frost mass might be also estimated as shown in Eq (8)
by using Eqs (4) and (5) for the local values
1 of Figure 9 The values predicted by the correlation tions under the freezer condition were larger by a maximum
equa-of 30–50% than the values predicted by the present correlationequations under the heat pump condition This is mainly due
to the differences in the conditions such as air humidity, airtemperature, and surface temperature The freezer condition is
Trang 10S YOON ET AL 971
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
condi-tion 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
equa-tions 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
The present study can be summarized as follows
1 The estimated and measured frost masses agreed within an
uncertainty range of 8%
2 The sensitivity analysis of the parameters such as air
tem-perature, air humidity, air velocity, and surface temperature
on the frost thickness and frost mass were experimentallyinvestigated under the heat pump condition
3 Correlation equations for the local and average frost ness and frost mass under the heat pump condition wereproposed
thick-4 The values predicted by the correlation equations under thefreezer condition were larger by a maximum of 30–50%than the values predicted by the present correlation equationsunder the heat pump condition
5 The proposed correlation equations might be applied to part
of the freezer condition
NOMENCLATURE
Fo Fourier number (= αat/L2)
L length of the flat plate, m
ρ local frost density, kg/m3
¯ρ average frost density, kg/m3
[1] Trammel, G J., Little D C., and Lillgore, E M., A Study
of Frost Formed on a Flat Plate Held at Sub-Freezing
Tem-perature, ASHRAE Journal, vol 7, no 10, pp 42–47, 2004.
Trang 11[2] Brian, P L T., Reid R C., and Shah Y T., Frost
Deposi-tion on Cold Surfaces, Industrial & Engineering Chemistry
Fundamentals, vol 9, no 3, pp 375–380, 1970.
[3] Sanders, C T., The Influence of Frost Formation and
De-frosting on the Performance of Air Coolers, Ph.D Thesis,
Delft Technical University (The Netherlands), 1974
[4] O’Neal, D L., and Tree, D R., Measurement of Frost
Growth and Density in a Parallel Plate Geometry, ASHRAE
Trans., vol 90, part 2, no 2843, pp 278–290, 1984.
[5] Na, B C., and Webb, R L., Mass Transfer on and Within
a Frost Layer, International Heat and Mass Transfer, vol.
47, pp 899–911, 2004
[6] Mago, P J., and Sherif, S A., Frost Formation and Heat
Transfer on a Cold Surface in Ice Fog, International
Jour-nal of Refrigeration, vol 28, pp 538–546, 2005.
[7] Le Gall, R., Grillot, J M., and Jallut, C., Modelling of
Frost Growth and Densification, International Journal of
Heat and Mass Transfer, vol 40, pp 3177–3187, 1997.
[8] L¨uuer, A., and Beer, A., Frost Deposition in a Parallel Plate
Channel Under Laminar Flow Conditions, International
Journal of Thermal Sciences, vol 39, pp 85–95, 2000.
[9] Mao, Y., Besant, R W., and Rezkallah, K S., Measurement
and Correlations of Frost Properties With Airflow over a
Flat Plate, ASHRAE Trans Research, vol 91, pp 267–281,
1992
[10] Mao, Y., Besant, R W., and Chen, H., Frost
Character-istics and Heat Transfer on a Flat Plate Under Freezer
Operating Conditions: Part 1, Experimentation and
Corre-lations, ASHRAE Trans Research, vol 105, pp 231–251,
1999
[11] Yang D K., and Lee K S., Modeling of Frosting Behavior
on a Cold Plate, International Journal of Refrigeration,
vol 28, no 3, pp 396–402, 2005
[12] Lee, Y B., and Ro, S T., Frost Formation on a Vertical Plate
Simultaneously Developing Flow, Experimental Thermal
and Fluid Science, vol 26, pp 939–945, 2002.
[13] Kwon, J T., Lim, H J., Kwon, Y C., Koyama, S., Kim,
D H., and Kondou, C., An Experimental Study on
Frost-ing of Laminar Air Flow on a Cold Surface With Local
Cooling, International Journal of Refrigeration, vol 29,
pp 754–760, 2006
[14] Shin, S H., Cho, K., and Hayase, G., Effect of Air Velocity
on Frost Formation of Slit Fin-And-Tube Heat Exchanger
Under Frosting Condition, Proc Winter Annual ence of SAREK (Seoul, Korea), pp 252–257, 2007.
Confer-[15] Hayashi, Y., Aoki, K., and Yuhara, H., Study of Frost mation Based on a Theoretical Model of the Frost Layer,
For-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 neering, vol 107, pp 173–182, 1985.
Engi-[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, ternational Journal of Refrigeration, vol 27, no 4, pp.
In-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 cur- rently 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
engi-neering 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 tional Institute of Refrigeration) and a vice-president
(Interna-of SAREK (Society (Interna-of Air-Conditioning and eration Engineers of Korea) since 2008 He is currently working on compression and absorption refrigeration systems, and compact heat exchangers.
Trang 12CopyrightC 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
ex-changer 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
at-tracted wide attention, especially in terms of deficiency of water
resources and environmental deterioration However, the
disad-vantages 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
ma-jor 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,
increas-ing the heat transfer area, and augmentincreas-ing mean temperature
difference [1]
Normally, enhanced tubes and high-performance fins or
in-creasing heat transfer area is employed to significantly improve
the performance of condensers However, in this way both
man-ufacturing cost and power consumption are considerably
in-creased So far, this kind of heat transfer enhancement almost
reaches its upper limit due to very few benefits, or more
mate-rial 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: wudi02@mails.thu.edu.cn
mean temperature difference is greatly restricted by the ing conditions [2]
operat-A traditional condenser cooled by air is completely densed in one or several flow routes with complicated two-phaseflow evolution The condensate accumulates to form a thick film
con-on the tube surface and complicated two-phase flow [3], leading
to obvious decrease of the heat transfer coefficient and greatchange of the tube wall temperature along the flow direction
So, enhanced tubes [4–6], high-performance fins, and optimalflow arrangements [7–9] are frequently used to improve thecondenser 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 de-scribe and apply this idea and technology This kind of condenserwould automatically separate liquid from the vapor–liquid two-phase mixture and makes condensation always occur in dropletand/or unsteady thin film condensation mode throughout thewhole condenser, resulting in very high average condensationheat 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 formance due to huge latent heat release, and it can maintain
per-973
Trang 13Figure 1 Condensation modes: (a) thin film mode; (b) droplet mode.
relatively homogeneous wall temperature When wall
tempera-ture is less than the saturation temperatempera-ture of a vapor,
conden-sation 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
liq-uid conditions For thin film condensation, a liqliq-uid thin film
forms between hot vapor and cool wall surface and the liquid
can successfully spread on the wall surface, which prevents
va-por 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
sur-face 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
tradi-tional air cooling condenser in domestic air-conditioning
sys-tems The refrigerant flows into the condenser through two
par-allel 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.
Figure 3 Flow regimes during condensation in horizontal tube.
During condensation, liquid film forms on the wall, whichprevents direct contact of refrigerant vapor with the cool wallsurface and produces main heat transfer resistance of the con-densation Consequently, heat transfer performance rapidly be-comes worse as the liquid film becomes thicker and thicker, withfinally very complex two-phase flow with less and less vapor Forfilm condensation, the local liquid film thickness,δx, and average
heat transfer coefficient of whole tube length, h x, can be mately obtained from the Nusselt theoretical solution of conden-sation inside a horizontal tube for laminar liquid film flow as [1]
approxi-δx =
4ηlλl(Ts− Tw) x
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 Ts–Tw = 1.5 K Obviously, the liquid filmthickness significantly increases during condensation, whichincreases heat transfer resistance and decreases correspondingheat transfer coefficient rapidly Apparently, for a traditionaldesign in Figure 2 without liquid–vapor phase separation, thewhole condensation heat transfer is very weak in most regionsdownstream of tube routes A and B where more and more con-densed liquid is attached on the wall surface or accumulated into
a two-phase flow with much more liquid Consequently, moreheat transfer area is required and the volume of the whole con-denser is enlarged Also, the pressure drop would be very highand 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 uid was fully disturbed by vapor and hard-to-maintain smoothfilm) existing in the upstream zones of the tubes or zones veryclose to the inlets [16], due to refrigerant vapor directly contact-ing the tube wall and keeping a relatively high velocity in thisregion As a result, condensation performance would be verygood If condensation mode in the whole tube side can be main-tained the same as that in the entrance region, such as the firstone or two straight tubes of route A or B (the marked entranceregion in Figure 4), the heat transfer will be enhanced enor-mously One possible technology is to separate liquid phasefrom liquid–vapor two-phase flow in time to reduce liquid
Trang 14(a) Liquid film thickness
(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
hor-izontal 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 = h1A1(Tf1− Tω)= h2A2η(Tω− Tf2) (3)
where, h1and h2are the heat transfer coefficients for inside and
outside tube, respectively; A1and A2denote the areas of inside
tube and outside fins;η is fin efficiency; and Tf1 and Tf2 are
inside vapor and outside air temperature, respectively
Figure 5 Test module.
For a condenser, Q, η, Tf1, and Tf2 all are predetermined
If the condensation heat transfer coefficient inside the tube, h1,
increases, (Tf1–T w ) and/or inside tube area, A1, will decrease
according to Eq (3), and accordingly T w and(T w —Tf2) increase
As a result, the outside fin area, A2, will decrease, which impliesthat heat transfer performance of the condenser is enhanced It isconcluded that if condensation heat transfer is greatly enhanced,wall temperature will increase and temperature difference at twosides will be redistributed, which enhances total performanceand reduce the heat transfer area
Experimental Evidence
To validate the condensation mode right in the entrance gion of a tube, a series of fundamental experiments was con-ducted The test module is schematically shown in Figure 5.Vapor from the evaporator entered into a test channel with rect-angle cross section, 10 mm in width and 12 mm in height.The bottom of the test channel was a copper plate, 12 mm inwidth, 72 mm in length, and 2 mm in thickness, and the otherthree sides were covered by quartz plates A cooling circle wasequipped under the copper plate to exhaust the heat transferredfrom upper vapor condensation In total, five T-type thermo-couples embedded in copper plate were equipped to portray thetemperature evolution during vapor condensation on the copperplate The uncertainty of the temperature measurement was lessthan 0.1 K Meanwhile, a high-speed CCD camera was utilized
re-to capture the whole dynamic phenomena
The inlet vapor Reynolds number is calculated as
where U and f are wetted perimeter and cross-section area of
the test channel, respectively The condensation heat transfercoefficient is
h c= q m r
Trang 15Figure 6 Condensation behavior: (a) Re v = 1670, (b) Re v = 2600, (c) Re v=
5600, (d) Re v= 8960.
where q mrepresents mass flow rate of the condensed liquid, and
r is latent heat Acdenotes the copper plate area, and Tsand Tw
are vapor saturated temperature and inner wall surface
tempera-ture (obtained using a one-dimensional [1-D] conduction model
from the averaged value of five thermocouples measuring
out-side 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
conden-sation, droplet condensation always emerged During
condensa-tion, droplets gradually grew and experienced coalescence, and
later large droplets formed and spread on the wall surface At
a low Re vin Figure 6a, since the shear force induced by vapor
had little effect on the liquid film, a steady thin film formed
disturbance and rupture, and the film became thicker along the
channel as shown in Figure 3 At a relatively higher Re v, 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
mo-ments, 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
turbu-lent 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
dif-ferent 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
temporar-ily experienced a high value and decreased to a low value inFigure 7a, as condensation varied from droplet to thin film
temperature experienced fluctuation at a relatively high valueand finally reached a steady value This evolution character-istic corresponds to condensation behavior in Figures 6b and6c As vapor velocity increased, steady thin film condensationgradually converted to periodical thin film turbulent and dropletcondensation
The heat transfer coefficient is plotted as a function of Re vin
Figure 8, significantly increasing with Re v, which has the samevariation trend as the work of Annaiev et al [17] and is ex-
pected to be induced by condensation-mode evolution As Re v
increased, the stable thin film became instable and the thicknessdecreased gradually due to the enhanced vapor shear force As
a result, periodically fluctuating thin film and droplet tion emerged and coexisted, which greatly enhanced heat trans-fer performance Due to the increase of the condensation heattransfer coefficient, the wall temperature accordingly increased
condensa-as shown in Figure 7, which is consistent with the analysis inthe previous section, and heat transfer performance is expected
to enlarge As the value of hcchanged from 10 kW/m2-K to 15.7kW/m2-K, corresponding to changes in Re vfrom 2600 to 3600,the increase gradient was relatively higher due to condensationtransition from steady film condensation to coexistence of un-steady thin turbulent film and droplet condensation, consistentwith that in Figures 6 and 7 Apparently, vapor velocity wouldplay an important role in enhancing the condensation heat trans-fer It is expected that the condensation mode and correspondingheat transfer characteristics observed in a rectangular channelwould be similar to that in a circular tube
Novel Idea
The idea for enhancing the performance of an air coolingcondenser is illustrated in Figure 9 A short tube only contain-ing the entrance effect of condensation is employed everywhere
in the whole condenser, and a liquid–vapor separator is duced to drain the condensed liquid between two adjacent shorttubes This idea can ensure that the condensation in the wholecondenser is in the expected modes, or droplet and unstable thinfilm condensation mode
How-by liquid–vapor separators in time and pure vapor enters the
Trang 16D WU ET AL 977
Figure 7 Wall surface temperature evolution: (a) Re v = 1670, (b) Re v= 2600,
(c) Re v= 5600.
succeeding tubes, or the condensation is kept as an unsteady
thin liquid film and/or droplet mode on the wall If the
reason-able structure is designed to remain with invariant high vapor
velocity everywhere in the condenser, the condensation heat
transfer would be further enhanced
Figure 8 Condensation heat transfer coefficient.
An innovative condenser based on the proposed novel ideawas re-manufactured from the original condenser of an airconditioner with refrigerating capacity of 2300 W, shown inFigure 10 The coiled tube of the original one was composed
of 16 straight tubes and 15 return (or U) bends One passagetube length of the original one is about 690 mm For the in-novative 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 notchanged The flow passed was divided into two zones, con-densation and condensate subcooled zone In the condensationzone the straight tubes were connected using two manifolds
at two ends In these two manifolds several liquid–vapor arators were included to separate vapor from the two-phasemixture there and form several flow passages in the conden-sation zone This ensures that pure vapor enters the next flowpassage, so the high-performance condensation modes can bereached in the tubes Keeping an almost invariant vapor velocity
sep-at all inlets of straight tubes is another important technology
in order to disturb the thin liquid film or even achieve dropletcondensation mode This requires each flow passage to have dif-ferent straight tube numbers, referring to Figure 10b The down-stream serpentine subcooled zone was almost similar to originalcondenser
During operation, refrigerant vapor would first enter the densation zone Different from the original one, the condensa-tion zone has left and right manifolds rather than return bends atthe two sides of straight tubes After condensation in one tubepass, condensed liquid separates from the liquid–vapor mixtureautomatically and flows into the downstream serpentine tube.Consequently, a single vapor phase enters without condensedliquid, and droplet condensation or thin liquid condensation isexpected to occur in the succeeding flow passage, similar to
con-Figure 9 Novel idea for condenser design.
Trang 17Figure 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
Figure 11 Experimental system.
in Figure 11 Utilizing electrical heaters and humidifiers, twoseparated rooms provided the standard indoor (dry-bulb temper-ature 27◦C, wet-bulb temperature 19◦C) and outdoor conditions(dry-bulb temperature 35◦C, wet-bulb temperature 24◦C) forindoor and outdoor unit, respectively Measuring dry-bulb andwet bulb-temperature of the air, before and after flowing throughthe evaporator, the enthalpy difference was calculated Togetherwith measurement of air flow rate, refrigerating capacity was de-termined 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 arevery important parameters that influence the performance of anair conditioner In each experiment, both capillary length andrefrigerant charge were optimized to obtain the highest EER andrefrigerating capacity Table 1 lists the experimental results of anair conditioning system with original and innovative condenser,respectively
Utilizing liquid–vapor separators to separate condensed uid without any leak of refrigerant vapor and appropriate tubedistribution to keep inlet velocity high and nearly invariant foreach vapor flow passage, the vapor phase is expected to con-dense in the mode of droplet or unsteady thin liquid condensa-tion occurring in short straight tubes Apparently, heat transfercapacity of the condenser was greatly enhanced, about 165 W,
liq-or from 3127 W to 3292 W in Table 1, though the heat transferarea had about 37% decrease Meanwhile, refrigerating capac-ity 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
Innovative condenser
Original condenser
Trang 18D WU ET AL 979
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
de-crease 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
sig-nificantly improved by means of liquid–vapor separation
tech-nology, together with reasonable design of tube route
distribu-tion in condensadistribu-tion zone In the present case, heat transfer area
can be significantly reduced, almost 37% or even more,
com-pared to original one This reduction is expected for any other
air cooling condensers
CONCLUSIONS
In the present investigation, an innovative idea and
technol-ogy were introduced to design a new kind of high-performance
air cooling condensers This kind of condenser would
automati-cally separate liquid from the vapor–water mixture or two-phase
flow in time by the innovative structure design The key
technol-ogy of structure design includes liquid–vapor separator
struc-ture and the tube number distribution of each flow passage in
condensation zone Consequently, the single vapor phase
en-ters without any condensed liquid, and droplet condensation or
thin liquid condensation is expected to occur The heat
trans-fer 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 Re v,
which is expected to be induced by condensation mode
transi-tion from steady film condensatransi-tion 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
sys-tem were significantly improved by means of liquid–vapor
sep-aration 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
innova-tive design idea in present work have great potential for
prac-tical applications, and particularly can be extended from conditioning systems to other kinds of air cooling condensers,such as that utilized in power plants, chemical engineering, and
h c heat transfer coefficient of condensation, W m−2K−1
h x average heat transfer coefficient, W m−2K−1
q m mass flow rate of condensed liquid, kg s−1
δx local liquid film thickness, m
[2] Lienhard, J H IV, A Heat Transfer Textbook, 3rd ed.,
Phlogiston Press, Lexington, MA, 2005
[3] Cengel, Y A., Heat Transfer: A Practical Approach, 2nd
ed., McGraw-Hill, New York, 2002
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for Evaporation, International Journal of Heat and Mass
Transfer, vol 39, no 9, pp 1827–1838, 1996.
[5] Kuo, C S., and Wang, C C., Evaporation of R22 in a 7mm
Microfin Tube, ASHRAE Transactions, vol 101, no 1, pp.
1055–1061, 1995
[6] Kuo, C S., and Wang, C C., In-Tube Evaporation of
HCFC22 in a 9.52 mm Microfin/Smooth Tube,
Interna-tional Journal of Heat and Mass Transfer, vol 39, no 1,
pp 2559–2569, 1996
[7] Ellison, P R., Crewick, F A., Fischer, F K., and Jackson,
W L., A Computer Model for Air-Cooled Refrigerant
Con-denser With Specified Refrigerant Circuiting, ASHRAE
Transaction, vol 87, no 1, pp 1106–1124, 1981.
[8] Liang, S Y., Wong, T N., and Nathan, G K., Study on
Refrigerant Circuitry of Condenser Coils With an Energy
Destruction Analysis, Applied Thermal Engineering, vol.
20, pp 559–577, 2000
[9] Wang, C C., Jiang, J Y., Lai, C C., and Chang, Y J., Effect
of Circuit Arrangement on the Performance of Air Cooled
Condensers, International Journal of Refrigeration, vol.
22, pp 272–282, 1999
[10] Peng, X F., Wu, D., and Zhang, Y., Applications and
Prin-ciple of High Performance Condensers, Chemical Industry
and Engineering Progress, vol 26, no 1, pp 97–104, 2007.
[11] Peng, X F., and Jia, L., Equal Velocity Steam–Liquid Heat
Exchange, China, 02130914.0[P], 2004
[12] Peng, X F., Jia, L., and Cheng, Z S., Equal
Ve-locity Steam–Liquid Heat Exchanger, Taiwan, China,
188060[Patent], 2003 http://webpat.twipr.com
[13] Peng, X F., Wu, D., Lu, G., Wang, Z., and Huang,
M., Liquid–Vapor Separation Air Condenser, China,
200610113304.4[Patent], 2006 http://www.sipo.gov.cn
[14] Peng, X F., Wu, D., Wang, Z., and Lu, G.,
Multi-stage Condensation and Liquid–Vapor Separation Air
Condenser, China, 200710064952.X[Patent], 2007 http://
www.sipo.gov.cn
[15] Liu, H B., Lin, Z Y., and Zhang, Y., Design and Test of
a Novel Condenser, 7th National Conference of Industrial
Furnaces, China, 13–16 August, pp 146–152, 2006.
[16] Yang, Y., Peng, X F., Wang, X D., and Wang, B X.,
Con-densation in Entrance Region of Rectangular Horizontal
Channel, Journal of Thermal Science and Technology (in
Con-Di Wu received a B.E degree in mechanical
engi-neering from Tsinghua University, Beijing, China in
2006 He is a Ph.D student in the Department of mal Engineering, Tsinghua University His current interests include transport phenomena in porous me- dia and microscale heat transfer (with/without phase change).
Ther-Zhen Wang received a B.E degree in mechanical
en-gineering 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 be- havior in the nucleate boiling of binary mixtures and evaporation of binary mixtures.
Gui Lu received a B.E degree in mechanical
engi-neering 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 versity, Beijing, China He is a professor in the De- partment of Thermal Engineering, Tsinghua Univer- sity His research interests include micro heat trans- fer, heat transfer in porous media, and phase-change transport phenomena.
Trang 20CopyrightC 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,4and JUNG-IN YOON5
1Department of Refrigeration and Air-Conditioning Engineering, Pukyong National University, Pusan, South Korea
2Department of Mathematics, Bangladesh University of Engineering and Technology, Dhaka, Bangladesh
3Daeil Co., Ltd., Pusan, South Korea
4Pukyong National University, Pusan, South Korea
5School of Mechanical Engineering, Pukyong National University, Pusan, South Korea
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
closed-circuit 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
fac-tors responsible for power consumption A cooling water pump
Address correspondence to Professor Jung-In Yoon, School of
Mechani-cal Engineering, Pukyong National University, Pusan, South Korea E-mail:
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 showsthe concept of multiple paths
In the relevant literature, no results have been reported sofar involving the CWCT with multiple paths In this scenario,the objective of this article is to obtain basic data from thisexperimental study on a small-size CWCT with multiple pathsand analyze the performance characteristics
EXPERIMENTAL APPARATUS AND METHOD
A schematic diagram of the experimental apparatus used inthis study is shown in Figure 2 In the experiment, the prototype
992
Trang 21Figure 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
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
cool-ing 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
cool-ing tubes The coolcool-ing 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
Figure 2 Schematic diagram of the experimental apparatus.
Table 1 Specifications of heat exchanger and experimental conditions
Inlet wet-bulb temperature 23–29 [ ◦C]
heat and gets stabilized to a certain temperature while passingthough the constant-temperature tank Then it recirculates to theCWCT The spray water is uniformly distributed at the upperpart of the coils by pump and circulates in the tower The lowerwater tank section consists of spray-water collecting tank andambient-air forcing fans Ambient air constraints were main-tained to the required state with the help of cooler, air heater,and humidifier Table 1 gives the experimental conditions andthe tower geometry Under the experimental conditions given
in Table 1, the experiment was conducted with changing theflow rate and inlet temperature of cooling water, flow rate ofspray water, and wet-bulb temperature and velocity of inlet air.After running the experiment for a while, the temperature ofspray water and cooling water get stabilized and experimentaldata are read after all conditions are normalized The data wererecorded with the help of an automatic data logger (MX100),and all the readings at all inlets and outlets were collected afterthe experiment stabilized in a steady state The duration of thetest run can be no less than 1 h The uncertainties of the mea-sured and calculated parameters are estimated by following theprocedures described in ASME PTC-23 [10] (with a level ofconfidence of 95%) The experimental uncertainties are associ-ated 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 surements The uncertainty values are 1.14% and 2.45% for thewater flow rate and total systematic and random uncertainty,respectively
mea-THEORETICAL BACKGROUND
Heat transfers from a hot process fluid inside tubes to spraywater and to air through a water film Heat transfers from spray
Table 2 Measuring devices specifications
Water temperature T type –200 to 400 [ ◦C] ±0.1[ ◦C]Water flow rate Dwyer, series VFB 0–70 [L/m] ±1[%]
Air temperature T type –200 to 400 [ ◦C] ±0.1[ ◦C]
Trang 22994 G.-J SHIM ET AL.
water to air in latent and sensible forms The rate of heat lost by
cooling water is given by:
dq c = m c c pc dt c = U o (t c − t s )d A (1)
where U o is the overall heat transfer coefficient based on the
outer area of the tube To calculate U o, the following equation
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:
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
coeffi-cients 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
re-liability 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 capacity of the cooling water [kW]
+15%
-15%
Figure 3 Heat balance of the experimental apparatus.
43% and 17% higher than those having one path when air locities were 1 m/s and 3.5 m/s, respectively In Figure 5, in thecase of the CWCT using two paths, mass transfer coefficientsare also higher than the other with regard to the flow rate ofspray water per unit breadth
ve-Figures 6 and 7 show the heat transfer coefficient, h o, versusflow rate of spray water per unit breadth and air velocity In bothfigures, heat transfer coefficients found for the CWCT using onepath can be claimed to be highly similar to the correlation given
by Nitsu et al [5] Furthermore, heat transfer coefficients for twopaths are similar to the correlation by Parker and Treybal [4].This indicates that heat transfer coefficients increase with theincrease of flow rate of spray water per unit breadth but showhardly any increase with respect to air velocity Heat transfer
1.0
19.05mm, One path 19.05mm, Two paths Parker and Treybal [4]
Figure 4 Mass transfer coefficient k as a function of air velocity.
Trang 23Flow rate of spray water per unit breadth [kg/ms]
Figure 5 Mass transfer coefficient k as a function of flow rate of spray water
per unit breadth.
Flow rate of spray water per unit breadth [kg/ms]
Figure 6 Heat transfer coefficient hoas a function of flow rate of spray water
per unit breadth.
Figure 7 Heat transfer coefficient hoas a function of air velocity.
Cooling water temperature [o
19.05mm, One path 19.05mm, Two paths
mc: 1,560 kg/h
ms : 1,080 kg/h
v : 3 m/s
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
Temperature range (drop) with respect to a variable water inlet temperature (CWIT) and wet-bulb temperature(WBT) are shown in Figures 8 and 9 It is evident that rangeincreases almost linearly with the increasing temperature ofcooling water At the standard design condition, ranges thatwere measured in the CWCT using one and two paths are 4.2◦Cand 5.1◦C, respectively The range in the CWIT using two paths
cooling-is approximately 20% higher than that with one path From thetemperature range against variable inlet wet-bulb temperature, it
is clear that the range decreases with the increase of the wet-bulbtemperature This is because when the wet-bulb temperature atthe inlet increases, the temperature difference between the inletcooling water and air decreases Thus, the vaporization of thespray water outside the pipes decreases so that the falling ofthe temperature of the cooling water flowing inside the tubes
7
19.05mm, One path 19.05mm, Two paths
mc : 1,560 kg/h
ms: 1,080 kg/h
v : 3 m/s
Figure 9 Temperature range as a function of inlet wet-bulb temperature.
Trang 24Figure 10 0 Cooling capacity as a function of air velocity.
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
Cooling water flow rate [l/min]
The fundamental study of the performance characteristics
of the closed-circuit wet cooling tower with multiple paths hasbeen done experimentally with a rated capacity of 9 kW Theresults can be summarized as follows:
Heat and mass transfer coefficient of the CWCT using onepath was found to conform well to the already reported resultsfor almost all cases considered
In the optimum level, mass transfer coefficients for variableair velocity and spray water flow rate of the CWCT having twopaths are respectively about 43% and 28% higher than thosehaving one path
The temperature drop of the cooling water for the CWIThaving two paths is nearly 20% higher than that with one path
h convective heat transfer coefficient, W m−2K−1
h i heat transfer coefficient for water inside the tubes, W m−2
K−1
h o heat transfer coefficient between tube external surface andspray water film, W m−2K−1
k mass transfer coefficient, kg m−2s−1
Trang 25[1] Bedekar, S V., Nithiarasu, P., and Seetharamuz, K N.,
Ex-perimental Investigation of the Performance of a
Counter-Flow, 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
Indirected-Contact Evaporative Cooling-Tower Behaviour, Applied
Energy, vol 78, pp 433–451, 2004.
[4] Parker, R O., and Treybal, R E., The Heat, Mass
Trans-fer Characteristics of Evaporative Coolers, Chemical
En-gineering Progress Symposium Series, Buffalo, NY, pp.
138–149, 1961
[5] Nitsu, Y., Naito, K., and Anzai, T., Studies of the
Char-acteristics and Design Procedure of Evaporative Coolers,
Journal of the Society of Heating, Air-Conditioning,
Sani-tary 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
Cor-relations 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
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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.
M M A Sarker is an associate professor of
mathe-matics at Bangladesh University of Engineering and Technology, Dhaka, Bangladesh He received his M.Sc (applied mathematics) degree from the Uni- versity of Dhaka, and an advanced studies in mas- ter of statistics degree from Katholieke Uviversiteit Leuven, Belgium He completed a Ph.D from Puky- ong National University, Pusan, South Korea He has been teaching at BUET since 1994 His research con- tributions were in the field of refrigeration and air- conditioning 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 previ- ously in charge of a laboratory involved in character- istics of heat and mass transfer on absorption refriger- ator 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
Na-tional 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 per- formance characteristics of refrigeration systems us- ing hydrocarbon refrigerants He was a researcher on characteristics of oil properties from discharge refrig- erants 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
Me-chanical Engineering at Pukyong National sity, Pusan, South Korea He received his M.Sc de- gree from Pukyong National University and his Ph.D degree in 1995 from Tokyo University of Agricul- ture & 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 re- search contributions are in the field of refrigeration and air-conditioning engineering He is currently working on the development
Univer-of the LNG refrigeration cycle, performance Univer-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.