-
lable at ScienceDirect
Building and Environment 46 (2011) 657e668Contents lists
avaiBuilding and Environment
journal homepage: www.elsevier .com/locate/bui ldenvNumerical
study of a M-cycle cross-flow heat exchanger for indirect
evaporativecooling
Changhong Zhan a,b, Xudong Zhao c,*, Stefan Smith c, S.B. Riffat
a
aDepartment of the Built Environment, University of Nottingham,
University Park, Nottingham NG7 2RD, UKb School of Civil
Engineering, Northeast Forestry University, Harbin 150040, Chinac
Institute of Energy and Sustainable Development, De Montfort
University, The Gateway, Leicester LE1 9BH, UKa r t i c l e i n f
o
Article history:Received 16 June 2010Received in revised form15
September 2010Accepted 21 September 2010
Keywords:Evaporative coolingCross-flowHeat and mass
transferNumerical simulation* Corresponding author. Tel.: 44 116
257 7971; faE-mail address: [email protected] (X. Zhao).
0360-1323/$ e see front matter 2010 Elsevier
Ltd.doi:10.1016/j.buildenv.2010.09.011a b s t r a c t
In this paper, numerical analyses of the thermal performance of
an indirect evaporative air coolerincorporating a M-cycle
cross-flow heat exchanger has been carried out. The numerical model
wasestablished from solving the coupled governing equations for
heat and mass transfer between theproduct and working air, using
the finite-element method. The model was developed using the
EES(Engineering Equation Solver) environment and validated by
published experimental data. Correlationbetween the cooling
(wet-bulb) effectiveness, system COP and a number of air
flow/exchanger param-eters was developed. It is found that lower
channel air velocity, lower inlet air relative humidity, andhigher
working-to-product air ratio yielded higher cooling effectiveness.
The recommended average airvelocities in dry and wet channels
should not be greater than 1.77 m/s and 0.7 m/s, respectively.
Theoptimum flow ratio of working-to-product air for this cooler is
50%. The channel geometric sizes, i.e.channel length and height,
also impose significant impact to system performance. Longer
channel lengthand smaller channel height contribute to increase of
the system cooling effectiveness but lead to reducedsystem COP. The
recommend channel height is 4 mm and the dimensionless channel
length, i.e., ratio ofthe channel length to height, should be in
the range 100 to 300. Numerical study results indicated thatthis
new type of M-cycle heat and mass exchanger can achieve 16.7%
higher cooling effectivenesscompared with the conventional
cross-flow heat and mass exchanger for the indirect evaporative
cooler.The model of this kind is new and not yet reported in
literatures. The results of the study help withdesign and
performance analyses of such a new type of indirect evaporative air
cooler, and in further,help increasing market rating of the
technology within building air conditioning sector, which
iscurrently dominated by the conventional compression refrigeration
technology.
2010 Elsevier Ltd. All rights reserved.1. Introduction
Air conditioning of buildings is currently dominated
byconventional compression refrigeration system, which takes
over95% of the market share in this sector. This kind of system is
highlyenergy intensive due to extensive use of electricity for
operation ofthe compressor, and therefore, is neither sustainable
nor environ-mentally friendly. The use of indirect evaporative
cooling has a highpotential for meeting air conditioning needs at
low energy costs.This, however, is dependent on the capacity of
additional watervapour that can be held by the cooling air stream.
Whilst morecommonly applied in hot, arid climatic regions such as
the MiddleEast, part of the Far East, North/South America and
Europe, there isx: 44 116 257 7981.
All rights reserved.an increasing trend for such systems to be
applied in low energybuilding designs in less suited climatic
regions such as in the UK.Recent research associated with projected
future climate in the UKshows at least a probable increased
potential for evaporativecooling in this region, particularly when
being jointly operatedwith desiccant dehumidification [1,2].
Indirect evaporative cooling systems have the advantage ofbeing
able to lower the air temperature without increasinghumidity of the
conditioned space and avoid potential health issuesfrom
contaminated water droplets entering occupied spaces (asassociated
with direct evaporative cooling systems). These systemsusually
require much less electric power that mechanical vapourcompression
uses for air conditioning [3]. Therefore, such systemswill help
reduce electricity consumption, and thus contribute toreducing
greenhouse gas emissions. It has widely been used asa low energy
consuming device for various cooling and air
mailto:[email protected]/science/journal/03601323http://www.elsevier.com/locate/buildenvhttp://dx.doi.org/10.1016/j.buildenv.2010.09.011http://dx.doi.org/10.1016/j.buildenv.2010.09.011http://dx.doi.org/10.1016/j.buildenv.2010.09.011
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Nomenclature
A heat transfer area, m2
cp specific heat of air, J/kg CCOP energy efficiency of the IEC
(Indirect Evaporative
Cooler)d equivalent diameter of the air passage, mh convective
heat transfer coefficient, W/m2 Chm mass transfer coefficient, m/si
specific enthalpy of air, J/kgL length, mLe Lewis numberm air mass
flow rate, kg/sNu Nusselt numberP theoretical fan power, WPr
Prandtl numberQ heat flux, W/m2
Re Reynolds numbert temperature, Cu velocity, m/s
V air volume flow rate, m3/sw humidity ratio of moist air, kg/kg
dry airDp pressure loss, Pag latent heat of water evaporation,
J/kg3 effectiveness, %h dynamic viscosity , Pa sr density,
kg/m3
F0 cooling capacity, W
Subscripts1 dry side2 wet sidea,f air flowdb dry bulbin inletl
latentsu supply airw wallwb wet-bulbwk working air
Fig. 1. Schematic of the traditional cross-flow heat and mass
exchanger for indirectevaporative cooling.
C. Zhan et al. / Building and Environment 46 (2011)
657e668658conditioning applications in industrial, agricultural and
residentialsectors [4e7] for providing low temperature fluids (e.g.
air, water).Indirect components can also be combinedwithmechanical
vapourcompression air conditioning systems to achieve very high
effi-ciencies while delivering comfort cooling that is equal to
conven-tional air conditioning.
In an indirect evaporative cooler (IEC), a primary (also
calledproduct) air stream is cooled by simultaneous heat and
masstransfer between a secondary (also called working) air stream
andawet wall surface. The latent heat transport, in connectionwith
thevaporization of the liquid film, plays an important role in the
heattransfer process [8,9]. Most commercially available IECs
areequipped with standard cross-flow heat exchangers that havea
stacked structure of heat and mass transfer plates as shown inFig.
1. In principle, the structure allows the product air to flow
overthe dry side of a plate and the working air to flow
perpendicular tothe product flow direction over the opposite wet
side of the plate.The wet side absorbs heat from the dry side by
evaporating waterand therefore cooling the dry side, while the
latent heat of vapor-izing water is given to the wet side air. In
an ideal operation, theproduct air temperature on the dry side of
the plate will reach thewet-bulb temperature of the incoming
working air, and tempera-ture of the working air on the wet side of
the plate will increase tothe incoming product air dry-bulb
temperature andwill reach 100%saturation. However, practical
systems are far from ideal. It hasbeen suggested that only 50e60%
of the incoming working air wet-bulb temperature can be achieved
for a typical indirect evaporativecooling device [10], while in
most systems theworking and productair come from the same source
(i.e. ambient air) and therefore havethe same temperature level.
This type of exchanger has beencomprehensively studied and
developed, as suggested in the liter-ature [8e13], with no great
potential to further improve the coolingeffectiveness (efficiency)
of the exchanger.
In recent years, a new type of heat and mass exchanger (Fig.
2a)utilizing the benefits of the Maisotsenko cycle [14] has
beendeveloped commercially [15e17]. In this type of exchanger, part
ofthe surface on the dry side is designed for the working air to
passthrough and the rest is allocated to the product air. Both
theproduct and working air are guided to flow over the dry side
alongparallel flow channels. There are numerous holes
distributedregularly on the area where the working air is retained
and each ofthese allows a certain percentage of air to pass through
and enterthe wet side of the sheet. The air is gradually delivered
to the wetside as it flows along the dry side, thus forming an even
distributionof airstreams over the wet surface. This arrangement
allows theworking air to be pre-cooled before entering the wet side
of thesheet by losing heat to the opposite wet surface. The
pre-cooled airdelivered to the wet side flows over the wet surface
along channelsarranged at right angles to the dry side channels,
absorbing heatfrom the working and product air. As a result, the
product air iscooled before being delivered to spaces where cooling
is required,and the working air is humidified, heated and
discharged to theatmosphere. Owing to effect of pre-cooling, the
working air in thewet side (working air wet channel) has a much
lower temperatureand therefore, is able to absorb more heat from
its two adjacentsides, i.e. the dry working air flow side and the
dry product air flowside. As a result, the cooling (wet-bulb)
effectiveness of the newstructure would be higher than that in the
traditional cross-flowexchanger (Fig.1). The cooling process is
shown on a psychometricchart in Fig. 2b. The manufacturers data has
indicated that theexchanger, namely M-cycle heat exchanger, could
obtain a wet-bulb effectiveness of 110% to 122%. [16,17]
Although significant progression has been achieved in
industrialand manufacturing exercise of such a new type of
M-cycleexchanger, to the authors knowledge there is no numerical
study ofthe new design being so far reported. To overcome the
shortfall inthe theoretical study of the exchanger and to further
enable
-
Fig. 2. Air flow and heat/mass transfer associated with the new
heat and mass exchanger. (a) Air flow profile. (b) air treatment
process (psychrometric indication).
C. Zhan et al. / Building and Environment 46 (2011) 657e668
659optimization of the exchanger performance, a numerical model
hasbeen developed to enable solving the coupled governing
equationsof the heat and mass transfer between two adjacent
airstreamsusing EES software [18]. Based on this development, the
effect ofvarious exchanger operating parameters to the system
perfor-mance have been investigated. This work is of significant
impor-tance to optimization of system configuration and development
ofthe solutions towards the better performance of the system
oper-ation. The work is expected to achieve high level of impact in
termsof increasing energy efficiency of the indirect evaporative
coolingsystems, extending its market share in building air
conditioningsector, and thus contributing to achieve the global
targets in energysaving and carbon reduction measures.
2. Description of the cooler with new type of heat and
massexchanger
Fig. 3a presents the structure of the M-cycle exchanger in
anISAW [17] indirect evaporative cooler e tac-150. This type
ofexchanger consists of numerous sheets of a fibre designed to
wickfluids evenly. The sheets are stacked together, separated by
channelguides located on one side of the sheet. One side of each
sheet is alsocoated with polyethylene to avoid penetration of
water. The guidesare fabricatedwith aplasticmaterial, and run along
the lengthof oneFig. 3. Schematic of the heat and mass exchanger in
ISAW TAC-150. (asheet, and thewidthof thenext sheet to forma
cross-flowwithin theexchanger. There are numerous regularly
distributed holes madealong the dry air flow paths, which are
located at the working airflow area. This configuration gradually
diverts air from the drychannel to the wet channels e the air flow
is perpendicular to thewet channels and has an even velocity
distribution. With heat andmoisture exchange this warmer and more
highly saturated air isdischarged to the atmosphere. In the
meantime, the product air isbeing cooled along its flow path. The
pre-cooling of the working airprovides a greater temperature
difference between the dry and wetchannel air, so improving the
cooling effectiveness of the system. Inthis studied case, the fibre
of the exchanger is 0.24 mm in thickness;and thewhole package
incorporates a total of 35drypassages and34wet passages, each of 4
mm in height.
Air flow distribution across the channels is shown
schematicallyin Fig. 3b, with heat andmass transfer taking place
between the dryandwet air channels. All the incoming air is
initially led into the drypassages (from Nos. 3 to 6 for product
and Nos. 1 to 2 for workingair), with the working air being
gradually diverted into the wetchannels, via the dedicate-designed
holes. This channel layoutcontributes to even distribution of the
air flow across the wetchannels without imposed flow adjustment at
the outlets of thesupply and exhaust air. Heat and mass transfer
will take placebetween the dry and wet channel air.) Structure
view. (b) Air flow distribution in the dry/wet channels.
-
a b
Fig. 4. Cell element applied for numerical simulation. (a) Cell
element for simulation. (b) Differential illustration.
Fig. 5. The calculating grids/meshes.
C. Zhan et al. / Building and Environment 46 (2011) 657e6686603.
Simulation approach
3.1. Heat and mass transfer mechanisms emathematical
indication
The cell element selected for numerical analyses is shown inFig.
4. The element consists of half the height of the dry channel,
theplate wall and half the height of the wet channel. Energy
balanceequations were applied to each single element, with
considerationof a pre-set boundary condition. This allowed the
temperature andhumidity distribution across the dry and wet channel
sections to beestablished.
To simplify the modelling process and mathematical analysis,the
following assumptions were made:
1. The heat and mass transfer is in steady state. The IEC
enclosureis considered as the system boundary.
2. The wet surface of the fibre sheet is completely saturated.
Thewater vapour is distributed uniformly within the wet
channel.
3. A temperature gradient for the channel cross-sectionwas set
tozero. Heat transfer in the separating plate is considered in
thevertical direction only. Within the working fluid, the
cross-stream convective heat transfer is considered as the
dominantmechanism of heat transfer.
4. Each element has a uniform wall surface temperature.
Ananalysis carried out by Zhao et al. [9] showed that the
thermalconductivity of the plate wall has little impact on the
magni-tude of the heat and mass transfer rates, owing to its
smallthickness (0.24 mm). The temperature difference between dryand
wet sides of the wall can be ignored.
5. Air is treated as an incompressible gas.
By applying principles of mass and energy conservation [19]
intothe differential element shown in Fig. 4, the heat and mass
transferprocesses in an IEC can be described with the following set
ofdifferential equations.
(1) The mass balance in the wet channel
The level of moisture in the working air could be calculated
asfollows:ma;f22
dwa;f2 hm
rw;a2 ra;f2
dA (1)
(2) The general energy balance within the element in Fig. 4 can
beexpressed as:dQl dQ1 dQ2 (2)(3) The energy balance in dry
passages
Dry passage air involves the forced convective heat
transfer,leading to change of the enthalpy of the air. Energy
balance in a drypassage could be written as,
dQ1 h1ta;f1 tw
dA
ma;f12
dia;f1 (3)
(4) The energy balance in wet passages
Wet passage air involves the forced heat and mass exchange,which
leads to a change of enthalpy of the air within the passages.The
energy balance within the passages can be written as,
dQl dQ2 ma;f22
dia;f2 (4)
where, for the forced convective heat and mass transfer
occurringin the wet passages,
dQ2 h2ta;f2 tw
dA (5)
dQl hmrw;a2 ra;f2
g dA (6)
The air flow within the pipes remains in a laminar flow
statewhen ReD< 2300 and becomes turbulent flow when ReD>
4000.Due to the passage size and air velocity, the air flow within
the
-
Fig. 6. Experimental validation e supply air temperature. (a)
Case 1. (b) Case 2.
C. Zhan et al. / Building and Environment 46 (2011) 657e668
661passage is considered to be laminar. In this case, the thermal
entrylength for laminar flow can be calculated as follows [20]:
Ld 0:05Re Pr (7)
For both entry region and fully developed flow conditions,
theNusselt number can be calculated using the following
equation:
Nu 1:86Re PrL=d
1=3 ha;fhw;a
!0:14(8)
The thermal entrance Nusselt numbers are higher than those
forthe fully developed case. For the developing flow conditions in
theentry region, the Nusselt number can be calculated as below.
Nu 3:660:0688 Re Pr
dL
1 0:04
hRe Pr
dL
i2=3 (9)The mass transfer coefficient between wet passage air
flow and
the wet surface of the wall may be calculated using the
followingequation:
hhm
rcpLe2=3 (10)
The mathematical expression for wet-bulb effectiveness can
bewritten as follows
3wb tdb;wk;in tdb;su
tdb;wk;in twb;wk;in(11)
It should be stressed that thewet-bulb effectiveness is
themajorparameter for evaluating the performance of the exchanger
andcooler, which represents the extent of the outlet air
temperature toapproach its relative wet-bulb of the inlet air.Fig.
7. Experimental validation e wet-bulThe theoretical energy
efficiency of the system can be defined asthe ratio of cooling
capacity to fan power consumption:
COP f0P
(12)
It should be stressed that cost of the water consumed in
thesystem was neglected owing to its minor value compared to
thecost of the electricity.
Cooling capacity, 40, can be expressed as:
f0 mptiwk;in ipt
(13)
The theoretical fan power, P, can be written as:
P DpwkVwk DpptVpt (14)It should be emphasized that the energy
efficiency obtained
from the simulation is an ideal value, which involves use of
thetheoretical fan power. Actual fan power will be 120e170% of
theideal value, leading to a drop in the calculated efficiency by
60e80%[21]. It should be noted that in this paper all the
subsequent figuresrelated to COP are ideal rather than practical
values.
By solving the above coupled differential equations, values
oftemperature and moisture content of the air at each single
elementcan be obtained, which results in a solution of the wet-bulb
effec-tiveness. A computer model incorporating the above equations
wasdeveloped in EES, by employing the finite-element approach. Fig.
5presents the air flow profile across a plate wall; the upper side
ofthe plate is arranged with wet passages, and the underside the
drypassages. In terms of a single element, the following
assumptionswere made: (1) each element has a uniform wall surface
temper-ature; (2) at the inlet and outlet of the dry or wet
channel, the airhas a uniform temperature and moisture content.
The trial computation results showed that under the
specifiedconditions a change of 0.02 C (0.2%) in the supply air
temperatureresulted from increasing the mesh grid from 1212 to 24
24. Theb effectiveness. (a) Case 1. (b) Case 2.
-
Fig. 8. Experimental validation e supply air moisture content.
(a) Case 1. (b) Case 2.
Table 1Operational conditions of the TAC-150 air cooler.
Inlet air dry-bulbtemperature (C)
Inlet air relativehumidity (%)
Inlet air wet-bulbtemperature (C)
Supply airflow rate
(m3/h) (kg/s)
30 50 22.0 150 0.0475
C. Zhan et al. / Building and Environment 46 (2011)
657e668662significant increase in computing time for the 24 24 grid
was notconsidered a rational modelling burden for the relatively
smalltemperature change. The mesh grid of 1212 was considered
toprovide sufficient accuracy for engineering applications and
was,therefore, adopted in the model set up. The Newton
Iterativemethod was used to solve a set of 6,357 equations in
relation tofluid flow and heat and mass transfer within the
passages of theheat exchanger.
4. Validation of the model accuracy using the
existingexperimental data
The model was set to the same operating conditions as
forexperimental cases 1 & 2 (i.e. the same inlet air parameters
andflow rates). Comparison between EES modelling results and
testingdata obtained from [22] was carried out, and differences
betweenthe results were analysed. This analysis established the
accuracy ofthe model in predicting the performance of the real
system.
The experimental cases (1 and 2) selected for validation refer
tothe testing to a ISAW TAC-150 cooler, as shown in Fig. 3, which
wascarried out by Qiu [22] at the standard environmental
chamberconditions. Case 1 was carried out at the controlled inlet
airconditions of 35% RH, 25 to 40 C dry bulb and 130 m3/h air
flowrate; whereas case 2 was at the condition of 50% RH, 25 to 40 C
drybulb and 130 m3/h air flow rate. During the testing, T-type
ther-mocouple probes and PT100 humidity sensors were installed
tomeasure the temperatures and relative humidity of air
flow,whereas the Testo 425 type handheld hotwire anemometer used
tomeasure the air velocity which resulted in calculation of the
airflow rate. All these measurement sensors were linked to a
DT500data logger and a computer for data recording and analyses.
Aprogramme was established in the Datatakers software to controlthe
Datataker to scan signals and report them at 5 second intervals.The
data was also saved at the same intervals as well.Table 2Results of
simulation.
Supply air flowrate (m3/h)
Average air speed indry channel (m/s)
Wet-bulbeffectiveness (%)
Supply atempera
150 1.77 51.1 25.8Modelling and experimental data regarding the
supply airtemperature, wet-bulb effectiveness and air moisture are
shownrespectively in Figs. 6e8. For case 1, the difference
betweenexperimental and simulated supply air temperature
is0.69e1.28 C, but for case 2 much closer agreement is shown witha
difference of as small as 0.02e0.05 C. For the two cases,
thehighest deviation in simulated to experimental supply
airtemperature is 3.4%. Case 2 shows greater agreement
betweenexperimental and simulated wet-bulb effectiveness, with case
1showing a difference in a range of 7.3%e9.4% and case 2 a
differenceof 0.2%e0.4%.
In theory, the moisture content of supply air should be equal
tothat of the inlet air. Fig. 7 shows good agreement to this in
theexperimental data for case 2, and shows supply air moisture to
beclearly higher than inlet air moisture for case 1. With the
reportedunsteady behaviour in the experiment of [22] and the
indicationthat the experimental data of case 2 is more accurate,
the level ofagreement shown between the experimental results and
simula-tion is considered to offer sufficient confidence in the
modellingprocess for the IECs air flow, heat and mass transfer.
5. Simulation results and analyses
5.1. Start-up operation and system performance
After being validatedwith the experimental data, themodel
wasutilized to investigate the effect of various operational
factors tosystem performance. The recommended operating conditions
ofthe TAC-150 unit (see Table 1) were used to set the initial
conditionsof the model [17]. The simulation indicated that the
system canachieve a cooling capacity of 200 W. Further information
resultingfrom the simulation is presented in Table 2.
For this operating condition, the temperature profiles of dry
air,wet air and the exchanging wall are presented in Fig. 9aec, the
heatflux in Fig. 10aec, and the moisture content profile of the dry
andwet air in Fig. 11a, b.
In Fig. 9a, it can be seen that the temperature of supply air in
drychannels (Nos. 3e6, referring to Fig. 3b) decreases along its
direc-tion of flow, and the temperature of working air in the dry
channels(Nos. 1 and 2, referring to Fig. 3b) has a bigger drop
because of thereduction of air mass flow along the way. Wet air
temperatureshows a different trend. As shown in Fig. 9b, the
temperature ofworking air in the wet channels of Nos. 1e3 are lower
than those ofNos. 4e6. Moreover, both of them initially fall before
rising againir averageture (C)
Coolingcapacity (W)
Total inlet airflow rate (m3/h)
Exhaust airflow rate (m3/h)
200.4 228 78
-
Fig. 9. Temperature distribution across the exchanger plate. (a)
Dry side (dry passages). (b) Wet side (wet passages). (c) Wall.
C. Zhan et al. / Building and Environment 46 (2011) 657e668
663along the wet air flow direction. The temperature of working air
inwet channels Nos. 4e6 has a smaller increase to the end of
thechannels - relative to the other three wet channels. Shown
inFig. 9c, the general trend of wall temperature is that it
decreasesalong the dry-airs direction of flow and increases along
the wet-airs flow direction.
As shown in Fig. 10a, the convective heat transfer
decreasesalong the flow path of dry air as a result of the observed
(see Fig. 9aand c) decrease in the temperature difference between
the drychannel air and the wall. The heat transfer rate in dry
channels Nos.3e6 is higher than that of Nos. 1 and 2 as they have
different airmass flow rates (see Table 2).
Referring to Fig. 10b, the wet air is not initially saturated
and hasa higher temperature than the wet wall close to the entrance
of thewet channels (comparison of Fig. 9b and c). This results in
heat beingtransferred to thewater reservedon thewet sideof thewalle
leadingto the evaporation of thewater. After travelling to a
critical point, thetemperature of wet air is lower than that of the
wet wall, so theconvective heat flux has become negative (as shown
in Fig. 10b),which means that the wet air picks up both sensible
and latent heatfrom the wall.
From Fig. 10c, the working air in wet channels 1e3 is
pre-cooledover a very short distance in dry channels 1e2 to a
temperaturelower than that of the wall surface. Below this
temperature theworking air absorbs heat from the wet wall.
The moisture content of the air in dry channels keeps
constant,shown in Fig. 11a. As shown in Fig. 11b, the moisture
content of theair in eachwet channel increases along themain
direction of flowofthewet air. This is due to the continuous
addition of moisture to theair along the direction of flow. This
results in a reduction in thedifference of moisture concentration
between the wall surface andthe air, ultimately leading to a
smaller driving force for evaporationand a smaller associated heat
flux (Fig. 10c). This behaviour isFig. 10. Heat transfer rate
across the exchanging plconsistent in each wet channel; the heat
flux in all wet channelsdecreases along the main direction of air
flow. The air in each wetchannel does not reach saturation; instead
the maximum relativehumidity is about 90%.
The results from the simulation allow determination of wet-bulb
effectiveness. Changing the values of air flow rate, ratio
ofworking-to-product air flow rates, temperature, and
moisturecontent allows different sets of simulation results to be
obtained.Further analyses of the results will allow the impact of
these vari-ables on cooling effectiveness to be determined.
5.2. Inlet air temperature impact
Varying the inlet air temperature between 20 C and 40 Cwhileall
other parameters remain unchanged, the simulationwas carriedout
using the above established computer model and the results
arepresented in Fig. 12a and b. A trend in increasing supply
airtemperature, cooling capacity, wet-bulb effectiveness and COP
ofthe evaporative cooler coincides with increasing inlet air
temper-ature. This is due to a higher inlet air temperature
resulting ina larger temperature difference between the inlet air
and the water.Although the temperature of inlet air has doubled
(from 20 C to40 C), the wet-bulb effectiveness increased from 46.5%
to 56.3%,and COP increased from 230 to 440. This shows the IEC is
moreefficient at higher temperatures, suggesting it is more suited
toa high-temperature environment.
5.3. Air relative humidity impact
Keeping other parameters unchanged, the impact of
relativehumidity of inlet air can be seen in Fig. 13a and b. When
the relativehumidity of inlet air increased from 0.1 to 0.9,
accordingly the wet-bulb effectiveness increased from 48.3% to
54.1%, but COP and theate. (a) Dry side. (b) Wet side. (c) On the
wall.
-
Fig. 11. Moisture content distribution across the exchanging
plate. (a) Dry side. (b) Wet side.
C. Zhan et al. / Building and Environment 46 (2011)
657e668664cooling capacity dropped from 640 to 60 and from 388.2W
to38.17W, respectively. At RH 0.9 the temperature drop across the
IECcooler was only 0.76 C, so showing the evaporative cooler to
beredundant inahumidclimate. Therefore, the inlet airhumidity
shouldnot be higher than 65%, corresponding to a supply air
temperature of26 C, which is advised in the European standard of
[23]. As the wet-bulb effectiveness shows similar trends in both
increasing airtemperature and increasing RH, butwith different
outcome to the IECperformance, the wet-bulb effectiveness cannot be
considered toindependently characterize the performance of the
IEC.
5.4. Impact of air speed
Simulations were carried out to investigate the effect of
airspeed on the performance of the cooler. When the total inlet
airflow rate increases, the working air flow rate and product air
flowrate will increase in proportion, so does the air speed in
drychannels or wet channels. Varying the total inlet air flow rate
from50 to 500 m3/h while keeping other parameters unchanged,
thesimulation results are shown in Fig. 14a and b. Fig. 14a shows
thecooling capacity (W) to increase with increased flow rate
(i.e.increase in the air speed). The sensitivity of cooling
capacity,however, reduces at higher inlet air flow rates. Whilst
the coolingcapacity increases (from 83W to 270W) the supply air
tempera-ture also increases, which may be a limiting factor to
achievingdesired internal environmental conditions.
Fig. 14b shows a steep decline in both thewet-bulb
effectivenessand the COP (more so for COP) as inlet air flow rate
increases. Thepressure drop across the cooler (due to increased
flow rate) nega-tively impacts the COP e dropping from 3700 to 114
over theconsidered range of inlet flow rate. If the average supply
air20 24 28 32 36 400
5
10
15
20
25
30
35
40
0
50
100
150
200
250
300
350
400
TemperatureCooling capacity
Co
olin
g c
ap
ac
ity
[W
]
Su
pp
ly
a
ir te
mp
era
tu
re
[C
]
Inlet air temperature [C]
a b
Fig. 12. Impact of inletemperature is limited to a value of 26
C, the recommendedaverage air velocities in dry and wet channels
should be less than1.77 m/s and 0.7 m/s, respectively.
5.5. Impact of ratio of supply-to-total inlet air mass flow
rate
When the resistance of flow varies in either the dry or
wetchannels, the resulting variation in air velocity between the
dry andwet channels will influence the performance of the cooler.
Underthis situation, the resistance variation is hard to determine,
so theinfluence of the ratio of supply-to-total inlet air flow rate
on COPcant be given. For discrete total inlet air flow rates, the
influence ofdifferent average velocity inwet channels and dry
channels on bothwet-bulb effectiveness and cooling capacity were
investigated,through changing the ratio of supply air flow rate to
total inlet airflow rate from 0.1 to 0.9 by interval of 0.1, shown
in Fig. 15a and b.
Fig. 15a shows wet-bulb effectiveness decreases as the
consid-ered flow rate ratio increases, but with different paths.
The nature ofthese paths are such that above 228 m3/h (atwhich
point the supplyair flow rate reaches the specified maximum of 150
m3/h) the curvebecomes increasinglymore convex, and below200 m3/h
it becomesincreasingly more concave. The wet-bulb effectiveness can
evenachieve a value of 144% when the total inlet air flow is 50
m3/h witha ratio of 0.1, but the corresponding cooling capacity is
38.8 W,whichapparently is not applicable andeconomic for apractical
application.
At or belowan inlet flow rate of 228 m3/h, Fig.15b shows
coolingcapacity reaches a maximum in the range of 0.5 to 0.6 for
the ratioof supply air to inlet air (decreasing towards the limits
of theconsidered ratio). Above 228 m3/h the curves display two
maxima,but with differing measures of cooling capacity. The
maximumcooling capacities for flow rates of 250 m3/h and 300 m3/h
areInlet air temperature [C]
20 24 28 32 36 400
50
100
150
200
250
300
350
400
450
500
0.3
0.35
0.4
0.45
0.5
0.55
0.6
0.65
0.7
CO
P
COP
We
t-b
ulb
e
ffe
ctiv
en
es
s
Wet-bulb effectiveness
t air temperature.
-
0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 10
5
10
15
20
25
30
Co
olin
g c
ap
ac
ity
[W
]
Inlet air Relative Humidity
Su
pp
ly
a
ir te
mp
era
tu
re
[C
]
Cooling capacity
Temperature
0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 10
100
200
300
400
500
600
700
800
900
CO
P
We
t-b
ulb
e
ffe
ctiv
en
es
Inlet air Relative Humidity
COPWet-bulb effectiveness
a b
Fig. 13. Impact of inlet air relative humidity.
C. Zhan et al. / Building and Environment 46 (2011) 657e668
665within the lower half of the considered ratio range, for 400
m3/hand 500 m3/h the maximum cooling capacities move towards
theupper limit of the considered ratio range.
Based on Fig. 15b, the cooler has an optimal (theoretical)
flowrate ratio for different inlet flow rates to achieve maximum
coolingcapacity. As total inlet air flow increases (up to 250
m3/h), the ratioof supply-to-inlet air flow rate decreases to
maintain maximumcooling capacity.
At the specified conditions (total inlet air flow rate is 228
m3/h,shown in Table 2), the practical ratio of the supply-to-inlet
air flowrate is 0.657, which result in a deviation of 7% for
cooling capacityfrom the best ratio value of 0.5 (Fig. 14b).
Moreover, the wet-bulbeffectiveness is decreasing from 70.52% to
50.4% when theconsidered ratio rises from 0.5 to 0.657, which means
that thesupply air temperature is higher according to Eq. (11).
Therefore,within the limits of the considered test conditions, the
IEC unit canbe said to be operating within 93% of the maximum
coolingcapacity, and within 71% of the wet-bulb effectiveness
corre-sponding to the maximum cooling capacity.
5.6. Impact of exchanger geometry e channel height, and
length
Simulations were carried out to investigate effect of channel
size(height and length) on the COP, wet-bulb effectiveness, supply
airtemperature and cooling capacity.
Varying the height from 2 to 20 mm while keeping otherparameters
unchanged (as shown in Table 1), different sets ofsimulation
results were obtained. As shown in Fig. 16a and b, it canbe seen
that both the wet-bulb effectiveness and cooling capacity0 0.5 1
1.5 2 2.5 3 3.5 4
0
5
10
15
20
25
30
Temperature
Cooling capacity
Co
olin
g c
ap
ac
ity
[W
]
Su
pp
ly
a
ir te
mp
era
tu
re
[C
]
Air speed in Dry channels (m/s)
0.17 0.33 0.50 0.66 0.82 0.98 1.14 1.30 1.46 1.62
Air speed in Wet channels (m/s)
Total inlet air flow rate (m /h) 3
a b
Fig. 14. Impact ofdecrease with increasing channel height.
However, a small channelheight results in increased flow resistance
and decreased energyefficiency. The COP reaches its maximum of 578
when the channelheight is 12 mm, but the wet-bulb effectiveness and
coolingcapacity are respectively 20% and 78.4 W, which are quite
low. Atthe given channel lengths as shown in Table 1, if the
channel heightis greater than 4 mm, the supply air temperature will
exceed theindoor thermal comfort temperature of 26 C. A
compromiseamong the cooling effectiveness and COP suggests that the
channelheight should not be greater than 4 mm.
Varying the length of the dry channel from 0.2 m to 2.4
m,leading to change of the dimensionless length, i.e., ratio of
length toheight, from 50 to 600, while keeping all other
parametersconstant, simulation was carried out to investigate the
impact ofchannel length to cooling performance. As shown in Fig.
17a and b,it can be seen that wet-bulb effectiveness and cooling
capacityincrease with increasing dry channel length, whereas the
COP andsupply air temperature present the adverse trend under this
vari-ation. When the dimensionless length exceeds 300, the
variationrates of the above parameters tend to slow down.
Considering thefactors of material use and cooling performance, it
is suggested thatthe dimensionless length should be controlled to
between 100 and300. For the exchanger of 4 mm channel height, the
length of thedry channel should be in the range 0.4e1.2 m.
Similar simulation work to wet channel was carried out and
theresults present similar trend of variation, as shown in Fig. 18a
and b.The wet channel dimensionless length should be in the
range100e300. For the exchange of 4 mm channel height, the length
ofthe wet channel should be in the range 0.4e1.2 m.0 0.5 1 1.5 2
2.5 3 3.5 4
0
500
1000
1500
2000
2500
3000
3500
4000
0
0.1
0.2
0.3
0.4
0.5
0.6
0.7
0.8
0.9
1
Air speed in Dry channels (m/s)
CO
P
COP
Total inlet air flow rate (m /h)
effectivenessWetbulb
0.17 0.33 0.50 0.66 0.82 0.98 1.14 1.30 1.46 1.62
Air speed in Wet channels (m/s)
3
We
t-b
ulb
e
ffe
ctiv
en
es
s
air velocity.
-
Fig. 15. Impact of the supply-to-total air ratio.
Fig. 16. Impact of air passage height.
0 100 200 300 400 500 60021
22
23
24
25
26
27
150
200
250
300
350
400
450
Dimensionless dry channel length (L/H)
Su
pp
ly
a
ir te
mp
era
tu
re
[C
]
Co
olin
g c
ap
ac
ity
[W
]
TemperatureCooling capacity
0 100 200 300 400 500 60050
100
150
200
250
300
350
400
450
500
0.3
0.4
0.5
0.6
0.7
0.8
0.9
1
1.1
CO
P
Dimensionless dry channel length (L/H)
We
t-b
ulb
e
ffe
ctiv
en
es
s
COPWet-bulb effectiveness
a b
Fig. 17. Impact of dry channels length.
C. Zhan et al. / Building and Environment 46 (2011)
657e6686665.7. Comparison between the new (M-cycle) exchanger
andconventional (cross-flow) exchanger
Based on the optimised geometrical sizes of the exchanger,
i.e.,4 mm channel height and 1.2 m channel length, comparison
wascarried out to examine the difference in the performance of
thenew type of M-cycle exchanger and conventional
cross-flowexchanger. To enable the comparison, it is assumed:
(1) Both exchangers have the same effective heat/mass
transferarea. However, the new type of exchanger has extraworking
airpre-cooling space.(2) The same inlet air parameters (shown in
Table 1).(3) The same working/product air flow rates.
The simulation results are listed in Table 3. In general, the
newtype of exchanger achieved much higher cooling performancethan
the conventional cross-flow exchanger. Under the givenconditions,
the supply air temperature of the new exchanger is1.4 C lower than
that in the conventional exchanger; the wet-bulb effectiveness is
15.7% higher than that in the conventionalexchanger. As a result,
the system cooling capacity can beincreased by 62 W which is 16%
higher than the conventionalexchanger.
-
Table 3Comparison between improved tac-150 (M-Cycle) and
conventional (cross-flow)exchanger.
Supply airtemperature (C)
Wet-bulbeffectiveness
Coolingcapacity (W)
New exchanger 20.7 116.4% 456.2Conventional exchanger 22.1 99.7%
394.2
0 100 200 300 400 500 60021
22
23
24
25
26
27
150
200
250
300
350
400
450
Dimensionless wet channel length (L/H)
Su
pp
ly
a
ir te
mp
era
tu
re
[C
]
Co
olin
g c
ap
ac
ity
[W
]
TemperatureCooling capacity
0 100 200 300 400 500 60050
100
150
200
250
300
350
400
450
500
0.3
0.4
0.5
0.6
0.7
0.8
0.9
1
1.1
CO
P
Dimensionless wet channel length (L/H)
We
t-b
ulb
e
ffe
ctiv
en
es
s
COP
Wet-bulb effectiveness
a b
Fig. 18. Impact of wet channels length.
C. Zhan et al. / Building and Environment 46 (2011) 657e668
6676. Conclusions
The research has established a computer model able to
simulatethe thermal performance of a M-cycle cross-flow heat
exchanger.By using the model, detailed analyses into relation
between thecooling (wet-bulb) effectiveness, system COP and air
flow/exchanger operational parameters were undertaken. These led
tosuggestion to the most favourite operating conditions including
airvelocity, inlet air temperature and humidity and ratio of
working-to-product air, and optimised exchanger configuration e.g.
channellength and height etc. The model was also validated by the
pub-lished experimental data which indicated that the sufficient
accu-racy in simulation could be obtained. The model is
thereforesuitable for use in design of the indirect evaporative
cooling systemand prediction of the system operational performance.
This workwill help with enhancing the energy efficiency of this
kind ofsystem, exploring its market share in building air
conditioningsector, and thus contribute to achieve the global
targets in energysaving and carbon reduction measures. Furthermore,
extension ofthe model could be used in simulating the performance
of othertypes of exchanger for indirect evaporative cooling, e.g.,
counterflow exchanger with no divisional holes along the air flow
paths,which is part of the follow-up project to develop a
commercialexchanger and will be detailed in a separate paper.
It is indicated that (1) the new type of M-cycle heat and
massexchanger is able to achieve 16.7% higher cooling
effectivenesscompared with the conventional cross-flow heat and
massexchanger for the indirect evaporative cooler; (2) a higher
channelair velocity in the new exchanger results in a relatively
lower wet-bulb effectiveness and system COP, though the smaller
system sizeis considered to be spatially and economically
beneficial topotential users. The recommended average air
velocities in dryand wet channels should be less than 1.77 m/s and
0.7 m/s,respectively; (3) at the specified conditions, the optimum
ratio (interms of cooling capacity) between the exhaust and supply
airflow rate is 1:1; (4) reducing the channel height led to an
increasein cooling capacity or wet-bulb effectiveness and decrease
of thesystem COP. A compromise among these performance
indicessuggests that the channel height should be set to no more
than4 mm; (5) increasing the channel (both dry and wet) length led
toimproved cooling effectiveness but reduced the system COP. It
issuggested that the dimensionless channel length should
becontrolled to between 100 and 300. For the exchanger of 4
mmchannel height, both dry and wet channel lengths should be in
therange 0.4e1.2 m; and (6) the system performance is
highlydependent on the climatic conditions where it is applied. For
thegiven inlet air condition, the system can achieve 4.22 kW
ofcooling capacity as per kg/s of supply air, which is 50% of
themaximum capacity the system can achieve.
It should also be stressed that the above analysis is based ona
small size unit (TAC-150) for domestic use. This type of
exchangercan be extended to the large-scale central air handle unit
whichwillresult in significantly higher energy saving and carbon
reductionpotential.
Acknowledgement
The authors would like to acknowledge the financial
supportprovided for this research by the EU FP7 Marie Curie
InternationalIncoming Fellowship (PIIF-GA-2008-220079).
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Numerical study of a M-cycle cross-flow heat exchanger for
indirect evaporative coolingIntroductionDescription of the cooler
with new type of heat and mass exchangerSimulation approachHeat and
mass transfer mechanisms mathematical indication
Validation of the model accuracy using the existing experimental
dataSimulation results and analysesStart-up operation and system
performanceInlet air temperature impactAir relative humidity
impactImpact of air speedImpact of ratio of supply-to-total inlet
air mass flow rateImpact of exchanger geometry channel height, and
lengthComparison between the new (M-cycle) exchanger and
conventional (cross-flow) exchanger
ConclusionsAcknowledgementReferences