Experimental and numerical investigation of a high-efficiency dew-point evaporative cooler

Abstract This paper investigated the cooling performance of a high-efficiency dew-point evaporative cooler with optimised air and water flow arrangement using the combined experimental and numerical simulation method. The experimental results showed that the wet-bulb efficiency of the dew-point evaporative cooler was increased by 29.3% and COP was increased by 34.6%, compared to the existing commercial dew point air cooler of the same capacity. An improved two-dimensional, multi-factor engaged numerical model which can scale up and optimize the size and capacity of the cooler was developed. The numerical predictions agreed well with the experimental results, indicating that the cooling rate of the dew-point evaporative cooler is influenced by the dew-point evaporative cooler structure. The cooling efficiency of the dew-point evaporative cooler with corrugated plates is more than 10% higher than with flat plates and the cooling efficiency of the dew-point evaporative cooler with the actual flow arrangement is only 62%–67% that of a dew-point evaporative cooler with an ideal counter-flow arrangement. The cooling efficiency can be improved by increasing the channel length and the air entrance length, and decreasing the channel width and channel gap within a reasonable range.


Introduction
Evaporative cooling is an energy-saving and environmentallyfriendly cooling technology where the heat is absorbed and the air is cooled through water evaporation. Evaporative cooling methods can be divided into direct evaporative cooling (DEC) and indirect evaporative cooling (IEC) according to whether or not the product air is in contact with water. The IEC systems have an additional sensible heat exchange channel which is called the dry channel compared to DEC systems which have only a wet channel where the air contacts the water directly. Therefore, IEC systems have more applications than DEC systems because the product air (also called the supply air) is cooled without being added with moistures [1] . The air in the dry channel is called the primary air while the air in the wet channel is called the secondary air or working air. The product air can be cooled to the wet-bulb temperature of the incoming air for an ideal system, but the actual wet-bulb efficiencies of DEC systems are 40-80% and the cooling capacities are largely dependent on the incoming air properties [2] . The cooling efficiency can be improved by using dew-point evaporative cooling (also called M-cycle IEC) that improves the IEC design [3][4][5] . In this * Corresponding author.
E-mail addresses: lijm@tsinghua.edu.cn , lijm@mail.tsinghua.edu.cn (J. Li). design, the secondary air is first cooled in the dry channel before entering the wet channel so that the product air temperature can be lowered to the dew point of the incoming air for an ideal design. The wet-bulb efficiencies of M-cycle IEC systems are 10-30% higher than those of IEC systems [4] .
The cooling rates and energy efficiencies of M-cycle IEC systems are influenced by the M-cycle IEC system design, so many researchers have proposed various structures for M-cycle IEC systems. Hsu et al. [6] compared the efficiencies of three different heat exchanger structures and found that the product air can easily be cooled to below the wet-bulb temperature of the incoming air and can be cooled close to the dew-point temperature of the incoming air by optimizing the system design. Lee et al. [7] optimized the dew-point evaporative cooler design and compared the efficiencies of three types of dew-point evaporative coolers with cross-flow flat plates, cross-flow corrugated plates and countercurrent flow finned channels. Zhan et al. [8] studied the efficiencies of cross-flow and counter-flow dew-point evaporative coolers numerically and experimentally. Their results showed that the cooling capacity of a counter-flow dew-point evaporative cooler is about 20% higher than of a cross-flow type for the same operating conditions and same cooler volume and that the dew-point and wet-bulb efficiencies of counter-flow coolers are 15% −23% higher than those of cross-flow coolers. However, the cross-flow https s  Saturation near the water surface  t  Total  w  Wet channel air  wb Wet-bulb dew-point evaporative cooler has a higher energy efficiency COP, which is 10% higher than that of a counter-flow type. Riangvilaikul et al. [9,10] measured the performance of a counter-flow flat dew-point evaporative cooler for various incoming air temperatures and humidities. Their results showed that the wet-bulb efficiency was between 92% and 114% and the dew-point efficiency was between 58% and 84%. Kabeel and Abdelgaied [11] experimentally studied a counter-flow dew-point evaporative cooler with a baffle in the dry channel. Lee and Lee [12] experimentally studied the cooling in a dew-point evaporative cooler with ribs in the wet and dry channels. Xu et al. [13] experimentally studied the cooling efficiency and energy efficiency of a high-performance counterflow dew-point evaporative cooler for typical meteorological conditions of various regions. Their results showed that the wet-bulb efficiency range was 100-109.8% while the dew-point efficiencies were 67-76.3%. Anisimov and his team have optimized the designs of various dew-point evaporative coolers, including cross-flow and counter-flow patterns, designs with various channel structures, and various secondary air distribution methods [14][15][16][17][18][19][20][21] . As for the existing commercial dew-point evaporative cooler, the Coolerado Cooler TM with a cross-flow configuration, manufactured by Idalex, has been marketed to the residential and small commercial sector [22,23] . The dew-point evaporative cooler products of the Coolerado have developed to M50, C60 from the first product of M30. According to the lab evaluation of the Coolerado, the wet-bulb efficiency of the cooler ranged from 81% to 91% (averaging 86%) with the COP ranging from 9.3 to 40.5. [22,23] .Though many researchers have proposed various dew-point evaporative cooler systems, the efficiencies of these dew-point evaporative coolers are still not very high which leads to large, costly designs. Thus, the dew-point evaporative cooler designs need to be further optimized. The fabrication of a dew-point evaporative cooler is a complicated, time-consuming task, so the influences of various structures should be studied numerically rather than experimentally after the numerical models have been bell verified against experimental data. Zhao et al. [24] numerically studied the influences of channel length, width and gap on the efficiency of a counterflow dew-point evaporative cooler using an algebraic model. Cui et al. [25,26] improved that model by considering more factors and numerically studied the cooling rates of dew-point evaporative coolers with various structures using the Ansys Fluent solver. Lin et al. [27,28] proposed an unsteady one-dimensional numerical model to study a dew-point evaporative cooler that neglected the influence of the water. Many models have been proposed by various researchers [29][30][31][32][33][34][35] , but those models neglected many factors such as the influence of the water, the uneven flow of the air and conduction along the air flow direction. So, the structure optimization could not be studied using these models. More accurate models are needed to effectively optimize dew-point evaporative coolers.
Overall, the existing commercial dew-point evaporative coolers are relatively less energy efficient, leading to a large-sized and high cost operation because of unreasonable air and water flow arrangement. On the other side, the dew-point evaporative coolers previously studied in lab have smaller cooling capacities and thus are inappropriate to site application. In this paper, a highefficiency dew-point evaporative cooler was designed and established with improved air and water flow arrangement where the primary air and water and secondary air are closer to counter flow compared to existing commercial dew-point evaporative coolers and aluminum corrugated heat exchange plates and high wettability porous fiber placed on wet channel wall are used to increase the heat and mass transfer area and the channel supports in channels which can lead to air flow resistance are removed for the proposed dew-point evaporative cooer. To further enable the large scale and wide range applications of the dew-point evaporative coolers in real projects, this research also developed an improved two-dimensional, multi-factor engaged numerical model which can scale up and optimize the size and capacity of the unit to meet various requirements of real projects. The model itself has a few distinct advantages compared to existing models, namely, the air momentum equation and water flow equation are included and solved simultaneously with the air energy and mass transfer equations, and specific heat and mass transfer correlations for wet air is adopted in the energy and mass transfer equations. Further, the model was experimentally validated and refined by using the data obtained from coolers' testing by the authors, and thus, can guarantee its reliability and accuracy in the prediction and optimization.

Dew-point evaporative cooler experimental set-up
The cooling performance of a dew-point evaporative cooler design can be improved by (1) enlarging the heat exchanger heat and mass transfer area, (2) reducing the air flow resistance in the channel, and (3) making the air flow as close to ideal counter-flow as possible. These guidelines were used to design a counter-flow dewpoint evaporative cooler with corrugated plates.

Dew-point evaporative cooler
A high wettability porous fiber was placed in the wet channel wall to enlarge the wetted area with the corrugated plates used to increase the volumetric heat and mass transfer area. The heat exchanger was made of aluminum to reduce the weight and to eliminate the channel supports which increase the flow resistance. The corrugated plate geometry is shown in Fig. 1 and the entire dewpoint evaporative cooler design with the corrugated plates shown in Fig. 2 . The bottom part of the plate was flat for the air inlet. The primary air entered the dry channel from the bottom and flowed along the dry channels while being cooled by the water film flowing down along the plate in the adjacent wet channels. Then, part of primary air was delivered to the conditioned space from the upper end of the dry channel as the product air while the other part flowed into the adjacent wet channel as the working air through the interconnecting holes in the upper part of the plate. The water flowed downwards from the water distributor at the top of the wet channel along the wet surface covered with the porous fiber where the water contacts and transfers heat and mass with working air and absorbs the sensible heat from the primary in dry channel through the wall. The enthalpy and humidity of the working air both increase until the air is discharged to the outdoors at the bottom of the wet channel as exhaust air. The detailed parameters of the dew-point evaporative cooler are listed in Table 1 .

Experimental set-up
A schematic of the testing system is shown in Fig. 3 and the fabricated dew-point evaporative cooler is shown in Fig. 4 . The system was tested in a constant temperature and humidity lab where the temperature and humidity could be adjusted as required. The product air and the working air flow rates were controlled by adjusting the supply and exhaust air fan speeds which were both backward-curved centrifugal fans located in the product and working air outlet ducts. The product and exhaust air flow volumes were calculated by measuring the static pressure drop across converging-diverging nozzles. The air flow volumes were then calculated as: where Q is the volumetric air flow rate in m 3 /h, k is the nozzle calibration factor which is relevant to the inlet nozzles, P is the static pressure drop across the nozzle in Pa, ρ s is the standard air density of 1.2 kg/m 3 , and ρ is the actual air density in kg/m 3 .
The pressure drops were measured using micro manometers (measurement range: 0-20 0 0 Pa, accuracy: 10 Pa). The exhaust air was discharged to the outdoor through a duct, which gave an additional flow resistance. Thus, a variable-speed supplement fan was installed at the duct outlet to maintain the static pressure at room pressure at the dew-point evaporative cooler outlet. The temperature and humidity of the incoming air, product air and exhaust air were all measured using temperature and humidity sensors (HC2A-S, accuracy: ±0.1 • C, ±0.8% RH). The water temperature was measured using an RTD (accuracy: ±0.2 • C). The volumetric water flow was measured by a water flow meter (measurement range: 12.5-600 L/h, accuracy ±5%). The power consumption of the fans and water pump were measured by power meters (accuracy: ±0.5%).

Performance parameters
The cooling of a dew-point evaporative cooler is usually evaluated by the cooling efficiency and the coefficient of performance. The cooling efficiency can be denoted by the wet-bulb efficiency, ε wb , and dew-point efficiency, ε dp , with the energy efficiency usually given by the coefficient of performance (COP).

Uncertainty of the experimental results
The relative uncertainties of the volumetric air flow rate, cooling capacity, cooling efficiency and COP were calculated from the measured temperatures and humidities as: where y is one of the calculated parameters and x represents the directly measured values.
The calculation results showed that the highest relative uncertainty of the product air temperature was ±0.6%, highest relative uncertainty of the air flow volume was ±6.3%, the highest relative uncertainty of the cooling capacity was ±6.4%, the relative uncertainty of the COP was ±6.42%, the relative uncertainty of the cooling efficiency was ±1.7%.

Mathematical modeling
The physical model was based on half of the dry channel and half of the wet channel in the x -y plane as shown in Fig. 5 . The heat transfer in the dry channel and the heat and mass transfer in the wet channel are related through the energy, momentum and diffusion equations for the primary air, water and working air. The model was based on the following assumptions: For primary air in the dry channel: Mass equation: Momentum equation: Energy equation: where, h t is calculated as: For water, the momentum and heat transfer equations lead to [37] : Where, The heat transfer coefficient for the falling water film is given by [37] : The energy equation for the water film is: where, where d s (kg/kg) is the humidity of the wet air at the water film surface where the partial pressure of the water vapour is equal to the saturation pressure at the water film temperature and can be calculated as: where P t is the total pressure of air and P s is the saturation pressure of the water vapour which can be calculated using the correlation proposed by Goff-Gratch [38] , where, The moisture transport in the air equation is: where, The heat transfer coefficient of the air in the wet channel is calculated using the equation proposed by Dowdy et al. [39] which was obtained from the experiment results of evaporative cooling with the surface covered by a rigid cellulose media saturated with water: The mass transfer coefficient is then calculated using the Lewis correlation as The convective heat transfer coefficient of the primary air in the dry channel is calculated using the correlation proposed by Awad [40] which applies to high width-to-height square channels which is widely used in heat and mass exchangers for dew-point evaporative coolers.   where y * is the dimensionless length.

Boundary conditions
Boundary conditions for the primary air in the dry channel: Inlet: Outlet: The wall is assumed to be adiabatic: Boundary conditions for the working air in the wet channel: Inlet: Outlet: The working air flow in the wet channel is assumed to be uniform.
Boundary conditions for the water flow in the wet channel: Inlet: Outlet:

Simulation method
The governing equations for the primary air, working air and water film are coupled, so all the equations were solved simultaneously. The primary air flow in the dry channel was assumed to be laminar flow because of the low Reynolds number. The governing equations were discretized using the finite element method and solved using COMSOL. The mesh was generated in COMSOL and the grid independent was checked for each case. One of the grid independent checks is shown in Fig. 6 .

Model validation
The experiments were run for the various operating conditions listed in Table 2 , where the highest relative uncertainty of the product air temperature was ±0.6%. All the experimental results in Table 2 are compared with the numerical results in Fig. 7 which shows that the numerical results agreed well with the experimental results for the dew-point evaporative cooler. The numerical model was further validated for various dew-point evaporative cooler designs against the experimental results of Riangvilaikul and Kumar [10] and Duan et al. [41] shown in Figs. 8 and 9 . The relative error between the numerical and experimental results were all within 3%.

Experimental results discussion
Comparisons of the experimental results of tests 1 and 2, 3 and 4, and 12 and 13 show that the product air temperature decreases with increasing working-to-primary air ratio. The reason is that the heat and mass transfer coefficients of the working air increase with increasing working air flow rate as which can deduced from Eq. (24) . However, the cooling capacity did not always increase because reducing the product air flow rate and reducing the product air temperature cause the cooling capacity to change in opposite directions.
Comparisons of the experimental results of tests 2 and 3 and 5 and 6 show that the product air temperature decreases with decreasing water volumetric flow rate for a given inlet water temperature. Comparison of the experimental results of tests 6, 7 and 8 shows that the product air temperature decreases with decreasing water temperature for a given volumetric water flow rate. These influences of the water on the cooling of dew-point evaporative coolers agree with the effects seen in the numerical study by Liu et al. [36] .
The highest wet-bulb efficiency among the listed tests was 107.5% while the highest dew-point efficiency was 77.3%. The experimental results show that the cooling efficiency can be improved by decreasing the volumetric water flow rate and the inlet water temperature and increasing the working-to-primary air ratio.
During the tests, an interesting phenomenon was found after the water pump was turned off that the product air temperature decreased to a minimum and then increased again. Two sets of temperature data for the product air, incoming air and water after the water pump was turned off are shown in Figs. 10 and 11 In test 4 with a product air temperature of 22.1 • C and a volumetric water flow rate of 384 L/h, after the water pump was turned off at time of 09:31, the product air temperature first decreased to 20.9 • C and then increased again as the wet surface slowly dried. After the water pump was turned on again, the product air temperature again remained constant at 22.3 • C (Test 5). The product air temperatures again decreased after the water pump was turned off after test 5 and test 8. The lowest product air temperature is 17.8 • C as shown in Fig. 11 , where the cooling capacity is 10,829 W, the wet-bulb efficiency is 114%, the dew-point efficiency is 68% and the COP is 42.8 for the conditions of test 8 which is the design operating conditions. Also, the wet-bulb efficiency can be higher than 138% when the working-to-primary air ratio is 0.7, but the cooling capacity and COP are both relatively low.
The product temperature initially decreases after the water pump is turned off because there is still the water in the porous layer of the wet channel surface, and the water film thickness is decreasing with time because of the flow and evaporation, which will reduce the thermal resistance of the water film. When there is not enough water to wet the surface, the heat and mass transfer will decrease and the product air temperature will rise. So, there must be a best water flow volume for a given water temperature as discussed in [36] .

Comparison between the existed commercial dew-point evaporative cooler and the present one
The cooling performance of the Coolerado M30 which is arranged as cross flow for air and water was studied in the lab. The comparison between the M30 and the present prototype is shown in Table 3 . It is seen that under the similar conditions, the wet-bulb efficiency and COP are both improved obviously with   to 42.8) with smaller air flow rate of the present prototype for the similar inlet air parameters.

Numerical results and discussion
The numerical model was verified against the experimental results from this and other studies. The influence of the structure on the dew-point evaporative cooler is difficult to study experimentally, but can be easily studied numerically. Therefore, the dew-point evaporative cooler design was improved to increase the cooling by using the numerical model to study the influences of various design parameters on the dew-point evaporative cooler cooling capability.   Fig. 12 compares the product air temperatures and wet-bulb efficiencies of dew-point evaporative coolers with corrugated plates and with flat plates for various inlet air conditions for an average primary air velocity of 2 m/s, working-to-primary air ratio of 0.44, water flow rate of 0.48 L/h per cooling unit and water inlet temperature of 20 • C. The wet-bulb efficiency of the cooler with the corrugated plates is over 10% higher than that with the flat plates.

Comparison between dew-point evaporative coolers with flat plates and corrugated plates
The corrugated plates improve the cooling efficiency by increasing the cooler surface area and by increasing the heat and mass transfer coefficients. Thus, the corrugated plates are a simple and efficient way to improve the cooling capability of the dew-point evaporative cooler instead of flat plates.

Influence of channel length
The influence of the channel length on the cooling efficiency of the dew-point evaporative cooler is shown in Fig. 13 for the same operating conditions as in Section 5.1 . The phenomenon that the cooling efficiency is improved by increasing the channel length is easy to understand. A longer channel give more time for heat and mass transfer in the dew-point evaporative cooler. However, a longer channel also gives a larger air flow resistance (which leads to a higher fan power), higher material costs and a larger volume.  In addition, the cooling efficiency improvements brought by the increasing channel length become smaller with increasing channel length as shown in Fig. 13 . Thus, the optimal channel length should be chosen based on the specific requirements of the conditioned space.

Influence of channel width
The influence of channel width on the cooling efficiency of the dew-point evaporative cooler is shown in Fig. 14 for the same operating conditions as in Section 5.1 . The channel width influences the air flow in the dry channel, because the incoming air enters into the dew-point evaporative cooler from below the plate and flows out of the cooler from the top as shown in Fig. 2 . The cooling efficiency increases with decreasing channel width because the air flow is more uniform in a dry channel with a smaller channel width. However, a smaller channel width means a smaller volumetric air flow rate which will leads to a smaller cooling capacity. Also, the results in Fig. 15 then show that the best working-toprimary air ratio that gives the highest cooling capacity varies with the channel width.

Influence of channel gap width
The influence of the channel gap on the cooling efficiency of the dew-point evaporative cooler shown in Fig. 16 shows that the cooling efficiency increases with decreasing channel gap. As the channel gap width decreases, the volumetric air flow in each cooling unit decreases, so the air temperature decreases when the  total heat transfer for each cooling unit does not change a lot. But when the channel gap width decreases, the same total volumetric air flow of the cooler would then require more plates and the resulting dew-point evaporative cooler would be heavier. Also, for channel gaps smaller than 4 mm, the cooling efficiency increases become quite small as the channel gap further decreases.

Influence of air inlet arrangement
For the dew-point evaporative cooler analyzed in this study, the incoming air enters the dry channel from the bottom along a flat plate and flows along the dry channels in the corrugated plate, so the air inlet design strongly influences the air distribution. When the inlet length is less than 0.6 m, the cooling efficiency increases as the air inlet length increases as shown in Fig. 17 . Since the air inlet cannot be arranged exactly at the bottom of the cooler due to the air and water separation method as shown in Fig. 2 , the air flow in the dew-point evaporative cooler is not exactly counter flow. The cooling efficiency difference between the actual flow and an ideal counter flow pattern is shown in Fig. 18 which shows that the cooling efficiency of the dew-point evaporative cooler with the actual flow arrangement is only 62% −67% of that with the ideal counter-flow arrangement. Thus, the cooling efficiency can be significantly improved by optimizing the dew-point evaporative cooler design.

Conclusions
This paper investigated the cooling performance of a highefficiency dew-point evaporative cooler using the combined experimental and numerical simulation method. This cooler, characterized by the enhanced cooling effectiveness and significantly improved energy efficiency, can achieve 34.6% higher COP and 29.3% higher wet-bulb efficiency compared to the existing commercial dew-point evaporative cooler of the same capacity. An improved two-dimensional, multi-factor engaged numerical model where the air momentum equation and water flow equation are included and solved simultaneously with the air energy and mass transfer equations, and specific heat and mass transfer correlations for wet air is adopted in the energy and mass transfer equations, featured with advantages of improved accuracy and extended applicability, was developed and verified with experimental results and was used to optimize the dew-point evaporative cooler structure.
The testing results showed that the selected experimental dew point air cooler can achieve the wet-bulb efficiency of 114% and dew-point efficiency of 68.4%, cooling capacity of 10,829 W and maximum COP of 42.8 under the standard operational condition. The volumetric water flow rate and the inlet water temperature can be appropriately adjusted to give the best cooling output.
The numerical results showed that the cooling efficiency of the dew-point evaporative cooler with corrugated plates is more than 10% higher than with flat plates. The cooling efficiency of the dew-point evaporative cooler with the actual flow arrangement is only 62% −67% of the dew-point evaporative cooler with an ideal counter-flow arrangement. The cooling efficiency can be improved by increasing the channel length and the air inlet length and reducing the channel width and the channel gap within a reasonable range. The optimal parameters should be chosen based on the specific requirements of the conditioned space.

Conflict of interest statement
We declare that we have no financial and personal relationships with other people or organizations that can inappropriately influence our work, there is no professional or other personal interest of any nature or kind in any product, service and/or company that could be construed as influencing the position presented in, or the review of, the manuscript entitled, "Experimental and Numerical Investigation of a High-efficiency Dew-point Evaporative Cooler".