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: xzhao@dmu.ac.uk (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:xzhao@dmu.ac.ukwww.sciencedirect.com/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

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

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3000

3500

4000

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0.1

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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

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27

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200

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Dimensionless dry channel length (L/H)

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TemperatureCooling capacity

0 100 200 300 400 500 60050

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0.3

0.4

0.5

0.6

0.7

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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

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25

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Dimensionless wet channel length (L/H)

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]

TemperatureCooling capacity

0 100 200 300 400 500 60050

100

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0.3

0.4

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1.1

CO

P

Dimensionless wet channel length (L/H)

We

t-b

ulb

e

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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).

References

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http://www.idalex.com/technology/how_it_works_engineering_perspective.htmhttp://www.idalex.com/technology/how_it_works_engineering_perspective.htmhttp://www.fchart.com

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