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Title	Numerical study of a M-cycle cross-flow heat exchanger for indirect evaporative cooling
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Building and Environment 46 (2011) 657e668

Contents lists available at ScienceDirect

Building and Environment journal homepage: www.elsevier.com/locate/buildenv

Numerical study of a M-cycle cross-ﬂow heat exchanger for indirect evaporative cooling Changhong Zhan a, b, Xudong Zhao c, *, Stefan Smith c, S.B. Riffat a a

Department of the Built Environment, University of Nottingham, University Park, Nottingham NG7 2RD, UK School of Civil Engineering, Northeast Forestry University, Harbin 150040, China c Institute of Energy and Sustainable Development, De Montfort University, The Gateway, Leicester LE1 9BH, UK b

a r t i c l e i n f o

a b s t r a c t

Article history: Received 16 June 2010 Received in revised form 15 September 2010 Accepted 21 September 2010

In this paper, numerical analyses of the thermal performance of an indirect evaporative air cooler incorporating a M-cycle cross-ﬂow heat exchanger has been carried out. The numerical model was established from solving the coupled governing equations for heat and mass transfer between the product 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. Correlation between the cooling (wet-bulb) effectiveness, system COP and a number of air ﬂow/exchanger parameters was developed. It is found that lower channel air velocity, lower inlet air relative humidity, and higher working-to-product air ratio yielded higher cooling effectiveness. The recommended average air velocities in dry and wet channels should not be greater than 1.77 m/s and 0.7 m/s, respectively. The optimum 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 length and smaller channel height contribute to increase of the system cooling effectiveness but lead to reduced system COP. The recommend channel height is 4 mm and the dimensionless channel length, i.e., ratio of the channel length to height, should be in the range 100 to 300. Numerical study results indicated that this new type of M-cycle heat and mass exchanger can achieve 16.7% higher cooling effectiveness compared 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 with design 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 is currently dominated by the conventional compression refrigeration technology. Ó 2010 Elsevier Ltd. All rights reserved.

Keywords: Evaporative cooling Cross-ﬂow Heat and mass transfer Numerical simulation

1. Introduction Air conditioning of buildings is currently dominated by conventional compression refrigeration system, which takes over 95% of the market share in this sector. This kind of system is highly energy intensive due to extensive use of electricity for operation of the compressor, and therefore, is neither sustainable nor environmentally friendly. The use of indirect evaporative cooling has a high potential for meeting air conditioning needs at low energy costs. This, however, is dependent on the capacity of additional water vapour that can be held by the cooling air stream. Whilst more commonly applied in hot, arid climatic regions such as the Middle East, part of the Far East, North/South America and Europe, there is * Corresponding author. Tel.: þ44 116 257 7971; fax: þ44 116 257 7981. E-mail address: [email protected] (X. Zhao). 0360-1323/$ e see front matter Ó 2010 Elsevier Ltd. All rights reserved. doi:10.1016/j.buildenv.2010.09.011

an increasing trend for such systems to be applied in ‘low energy’ building designs in less suited climatic regions such as in the UK. Recent research associated with projected future climate in the UK shows at least a probable increased potential for evaporative cooling in this region, particularly when being jointly operated with desiccant dehumidiﬁcation [1,2]. Indirect evaporative cooling systems have the advantage of being able to lower the air temperature without increasing humidity of the conditioned space and avoid potential health issues from contaminated water droplets entering occupied spaces (as associated with direct evaporative cooling systems). These systems usually require much less electric power that mechanical vapour compression uses for air conditioning [3]. Therefore, such systems will help reduce electricity consumption, and thus contribute to reducing greenhouse gas emissions. It has widely been used as a low energy consuming device for various cooling and air

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Nomenclature A cp COP d h hm i L Le m Nu P Pr Q Re t u

heat transfer area, m2 speciﬁc heat of air, J/kg C energy efﬁciency of the IEC (Indirect Evaporative Cooler) equivalent diameter of the air passage, m convective heat transfer coefﬁcient, W/m2 C mass transfer coefﬁcient, m/s speciﬁc enthalpy of air, J/kg length, m Lewis number air mass ﬂow rate, kg/s Nusselt number theoretical fan power, W Prandtl number heat ﬂux, W/m2 Reynolds number temperature, C velocity, m/s

conditioning applications in industrial, agricultural and residential sectors [4e7] for providing low temperature ﬂuids (e.g. air, water). Indirect components can also be combined with mechanical vapour compression air conditioning systems to achieve very high efﬁciencies while delivering comfort cooling that is equal to conventional air conditioning. In an indirect evaporative cooler (IEC), a primary (also called ‘product’) air stream is cooled by simultaneous heat and mass transfer between a secondary (also called ‘working’) air stream and a wet wall surface. The latent heat transport, in connection with the vaporization of the liquid ﬁlm, plays an important role in the heat transfer process [8,9]. Most commercially available IECs are equipped with standard cross-ﬂow heat exchangers that have a stacked structure of heat and mass transfer plates as shown in Fig. 1. In principle, the structure allows the product air to ﬂow over the dry side of a plate and the working air to ﬂow perpendicular to the product ﬂow direction over the opposite wet side of the plate. The wet side absorbs heat from the dry side by evaporating water and therefore cooling the dry side, while the latent heat of vaporizing water is given to the wet side air. In an ideal operation, the product air temperature on the dry side of the plate will reach the wet-bulb temperature of the incoming working air, and temperature of the working air on the wet side of the plate will increase to the incoming product air dry-bulb temperature and will reach 100% saturation. However, practical systems are far from ideal. It has been suggested that only 50e60% of the incoming working air wetbulb temperature can be achieved for a typical indirect evaporative cooling device [10], while in most systems the working and product air come from the same source (i.e. ambient air) and therefore have the same temperature level. This type of exchanger has been comprehensively studied and developed, as suggested in the literature [8e13], with no great potential to further improve the cooling effectiveness (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 been developed commercially [15e17]. In this type of exchanger, part of the surface on the dry side is designed for the working air to pass through and the rest is allocated to the product air. Both the product and working air are guided to ﬂow over the dry side along parallel ﬂow channels. There are numerous holes distributed regularly on the area where the working air is retained and each of these allows a certain percentage of air to pass through and enter

V w Dp

g 3 h r F0

air volume ﬂow rate, m3/s humidity ratio of moist air, kg/kg dry air pressure loss, Pa latent heat of water evaporation, J/kg effectiveness, % dynamic viscosity , Pa s density, kg/m3 cooling capacity, W

Subscripts 1 dry side 2 wet side a,f air ﬂow db dry bulb in inlet l latent su supply air w wall wb wet-bulb wk working air

the wet side of the sheet. The air is gradually delivered to the wet side as it ﬂows along the dry side, thus forming an even distribution of airstreams over the wet surface. This arrangement allows the working air to be pre-cooled before entering the wet side of the sheet by losing heat to the opposite wet surface. The pre-cooled air delivered to the wet side ﬂows over the wet surface along channels arranged at right angles to the dry side channels, absorbing heat from the working and product air. As a result, the product air is cooled before being delivered to spaces where cooling is required, and the working air is humidiﬁed, heated and discharged to the atmosphere. Owing to effect of pre-cooling, the working air in the wet side (working air wet channel) has a much lower temperature and therefore, is able to absorb more heat from its two adjacent sides, i.e. the dry working air ﬂow side and the dry product air ﬂow side. As a result, the cooling (wet-bulb) effectiveness of the new structure would be higher than that in the traditional cross-ﬂow exchanger (Fig.1). The cooling process is shown on a psychometric chart in Fig. 2b. The manufacturer’s data has indicated that the exchanger, namely M-cycle heat exchanger, could obtain a wetbulb effectiveness of 110% to 122%. [16,17] Although signiﬁcant progression has been achieved in industrial and manufacturing exercise of such a new type of M-cycle exchanger, to the authors’ knowledge there is no numerical study of the new design being so far reported. To overcome the shortfall in the theoretical study of the exchanger and to further enable

Fig. 1. Schematic of the traditional cross-ﬂow heat and mass exchanger for indirect evaporative cooling.

C. Zhan et al. / Building and Environment 46 (2011) 657e668

659

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).

optimization of the exchanger performance, a numerical model has been developed to enable solving the coupled governing equations of the heat and mass transfer between two adjacent airstreams using EES software [18]. Based on this development, the effect of various exchanger operating parameters to the system performance have been investigated. This work is of signiﬁcant importance to optimization of system conﬁguration and development of the solutions towards the better performance of the system operation. The work is expected to achieve high level of impact in terms of increasing energy efﬁciency of the indirect evaporative cooling systems, extending its market share in building air conditioning sector, and thus contributing to achieve the global targets in energy saving and carbon reduction measures. 2. Description of the cooler with new type of heat and mass exchanger Fig. 3a presents the structure of the M-cycle exchanger in an ISAW [17] indirect evaporative cooler e tac-150. This type of exchanger consists of numerous sheets of a ﬁbre designed to wick ﬂuids evenly. The sheets are stacked together, separated by channel guides located on one side of the sheet. One side of each sheet is also coated with polyethylene to avoid penetration of water. The guides are fabricated with a plastic material, and run along the length of one

sheet, and the width of the next sheet to form a cross-ﬂow within the exchanger. There are numerous regularly distributed holes made along the dry air ﬂow paths, which are located at the working air ﬂow area. This conﬁguration gradually diverts air from the dry channel to the wet channels e the air ﬂow is perpendicular to the wet channels and has an even velocity distribution. With heat and moisture exchange this warmer and more highly saturated air is discharged to the atmosphere. In the meantime, the product air is being cooled along its ﬂow path. The pre-cooling of the working air provides a greater temperature difference between the dry and wet channel air, so improving the cooling effectiveness of the system. In this studied case, the ﬁbre of the exchanger is 0.24 mm in thickness; and the whole package incorporates a total of 35 dry passages and 34 wet passages, each of 4 mm in height. Air ﬂow distribution across the channels is shown schematically in Fig. 3b, with heat and mass transfer taking place between the dry and wet air channels. All the incoming air is initially led into the dry passages (from Nos. 3 to 6 for product and Nos. 1 to 2 for working air), with the working air being gradually diverted into the wet channels, via the dedicate-designed holes. This channel layout contributes to even distribution of the air ﬂow across the wet channels without imposed ﬂow adjustment at the outlets of the supply and exhaust air. Heat and mass transfer will take place between the dry and wet channel air.

Fig. 3. Schematic of the heat and mass exchanger in ISAW TAC-150. (a) Structure view. (b) Air ﬂow distribution in the dry/wet channels.

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C. Zhan et al. / Building and Environment 46 (2011) 657e668

a

b

Fig. 4. Cell element applied for numerical simulation. (a) Cell element for simulation. (b) Differential illustration.

3. Simulation approach

dQl ¼ dQ1 þ dQ2

(2)

3.1. Heat and mass transfer mechanisms e mathematical indication (3) The energy balance in dry passages The cell element selected for numerical analyses is shown in Fig. 4. The element consists of half the height of the dry channel, the plate wall and half the height of the wet channel. Energy balance equations were applied to each single element, with consideration of a pre-set boundary condition. This allowed the temperature and humidity distribution across the dry and wet channel sections to be established. 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 enclosure is considered as the system boundary. 2. The wet surface of the ﬁbre sheet is completely saturated. The water vapour is distributed uniformly within the wet channel. 3. A temperature gradient for the channel cross-section was set to zero. Heat transfer in the separating plate is considered in the vertical direction only. Within the working ﬂuid, the crossstream convective heat transfer is considered as the dominant mechanism of heat transfer. 4. Each element has a uniform wall surface temperature. An analysis carried out by Zhao et al. [9] showed that the thermal conductivity of the plate wall has little impact on the magnitude of the heat and mass transfer rates, owing to its small thickness (0.24 mm). The temperature difference between dry and 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] into the differential element shown in Fig. 4, the heat and mass transfer processes in an IEC can be described with the following set of differential equations.

Dry passage air involves the forced convective heat transfer, leading to change of the enthalpy of the air. Energy balance in a dry passage could be written as,

ma;f 1 dia;f1 tw dA ¼ 2

dQ1 ¼ h1 ta;f1

(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 ¼

ma;f2 dia;f2 2

dQ2 ¼ h2 ta;f2 dQl ¼ hm rw;a2

tw dA

(5)

(6)

ra;f2 g dA

The air ﬂow within the pipes remains in a laminar ﬂow state when ReD < 2300 and becomes turbulent ﬂow when ReD > 4000. Due to the passage size and air velocity, the air ﬂow within the

The level of moisture in the working air could be calculated as follows:

ma;f2 dwa;f2 ¼ hm rw;a2 2

ra;f2 dA

(1)

(2) The general energy balance within the element in Fig. 4 can be expressed as:

(4)

where, for the forced convective heat and mass transfer occurring in the wet passages,

(1) The mass balance in the wet channel

(3)

Fig. 5. The calculating grids/meshes.

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C. Zhan et al. / Building and Environment 46 (2011) 657e668

Fig. 6. Experimental validation e supply air temperature. (a) Case 1. (b) Case 2.

passage is considered to be laminar. In this case, the thermal entry length for laminar ﬂow can be calculated as follows [20]:

The theoretical energy efﬁciency of the system can be deﬁned as the ratio of cooling capacity to fan power consumption:

L ¼ 0:05Re Pr d

COP ¼

(7)

For both entry region and fully developed ﬂow conditions, the Nusselt number can be calculated using the following equation:

!0:14 Re Pr 1=3 ha;f Nu ¼ 1:86 hw;a L=d

(9)

The mass transfer coefﬁcient between wet passage air ﬂow and the wet surface of the wall may be calculated using the following equation:

(10)

The mathematical expression for wet-bulb effectiveness can be written as follows

3wb ¼

tdb;wk;in tdb;su tdb;wk;in twb;wk;in

It should be stressed that cost of the water consumed in the system was neglected owing to its minor value compared to the cost of the electricity. Cooling capacity, 40, can be expressed as:

f0 ¼ mpt iwk;in

0:0688 Re Pr dL Nu ¼ 3:66 þ h i2=3 1 þ 0:04 Re Pr dL

h ¼ rcp Le2=3 hm

(12)

P

(8)

The thermal entrance Nusselt numbers are higher than those for the fully developed case. For the developing ﬂow conditions in the entry region, the Nusselt number can be calculated as below.

f0

(11)

It should be stressed that the wet-bulb effectiveness is the major parameter for evaluating the performance of the exchanger and cooler, which represents the extent of the outlet air temperature to approach its relative wet-bulb of the inlet air.

ipt

(13)

The theoretical fan power, P, can be written as:

P ¼ Dpwk Vwk þ Dppt Vpt

(14)

It should be emphasized that the energy efﬁciency obtained from the simulation is an ideal value, which involves use of the theoretical fan power. Actual fan power will be 120e170% of the ideal value, leading to a drop in the calculated efﬁciency by 60e80% [21]. It should be noted that in this paper all the subsequent ﬁgures related to COP are ‘ideal’ rather than ‘practical’ values. By solving the above coupled differential equations, values of temperature and moisture content of the air at each single element can be obtained, which results in a solution of the wet-bulb effectiveness. A computer model incorporating the above equations was developed in EES, by employing the ﬁnite-element approach. Fig. 5 presents the air ﬂow proﬁle across a plate wall; the upper side of the plate is arranged with wet passages, and the underside the dry passages. In terms of a single element, the following assumptions were made: (1) each element has a uniform wall surface temperature; (2) at the inlet and outlet of the dry or wet channel, the air has a uniform ...