Heat and Mass Transfer of a Decoupling Cooling System: A Desiccant-Coated Heat Exchanger and a Dew-Point Evaporative Cooler

Abstract

This study presents an innovative de-coupling cooling technology where latent and sensible cooling loads are handled separately by a desiccant-coated heat exchanger (DCHE)-based dehumidifier and a dew-point evaporative cooler (DEC). The DCHE first removes the moisture of humid outdoor air by adsorption process. Subsequently, the DEC sensibly cools down the dehumidified air, maintaining the humidity ratio. Their performances are investigated numerically by analyzing the heat and mass transfer. The cyclic average outlet values of DCHE are applied to the inlet conditions for DEC simulation. Comparison between the simulation results and the experimental data was carried out and showed good agreement and a similar trend with a maximum discrepancy of 8.6% for DCHE and 3.2% for DEC, respectively. Key results revealed that moisture removal capacity (MRC) and latent cooling capacity (QL) for DCHE are largely affected by varying air dry-bulb and air wet-bulb temperatures, while the almost constant COPth was observed regardless of the variation of temperatures. For the DEC, the higher dew-point effectiveness and wet-bulb effectiveness were observed at the higher dry-bulb temperature and higher humidity ratio, while the higher sensible cooling capacity was observed at the higher dry-bulb temperature and lower humidity ratio.

1. Introduction

The growing population and rising energy demand in many countries are imposing significant challenges to energy and environmental sustainability. The global energy and environmental scenarios are closely interlinked. That is, the problems of supply and use of energy are related to global warming and climate change [1]. The electrical energy consumption for air conditioning accounts for 20–40% of the total electricity used in buildings around the world today [2, 3]. Global energy demand for air conditioning is expected to triple by 2050, requiring new electricity capacity that is the equivalent to the combined electricity capacity of the United States, the EU, and Japan today. Accordingly, the global stock of air conditioners in buildings will grow to 5.6 billion by 2050, up from 1.6 billion today, which means new air conditioners are sold every second for the next 30 years [4].

In conventional air-conditioning systems, the compressor efficiency has shown a significant improvement chronologically from 1.2 kW/RT in the 1990s to 0.85 ± 0.05 kW/RT [5]. Since 2000, however, this improvement trend has leveled asymptotically, implying that efficiency improvement for compressor stages and heat exchangers has been saturated. Despite massive investment in the R&D of cooling technologies by chillers’ manufacturers, there are physical and material limits to which efficiency improvement of the major components in the cooling cycle can be attained. If one were to continue to improve the cooling efficiency of conventional chillers (with HFC refrigerants), the improvement in kW/RT may improve only marginally by less than 5% and no quantum efficiency improvement can be expected. Therefore, we believe that there is a need for an out-of-box solution for cooling where the consumption of energy and water can be reduced significantly, by as much as 50% so as to attain sustainable cooling for the future, addressing the space cooling needs of the present and future generations.

Recently, many studies have been focused on the development of a desiccant-coated heat exchanger (DCHE) to improve dehumidification performance by employing various desiccant materials and heat exchangers. Sunhor et al. [6] investigated experimentally the dehumidification behavior of AlPO-zeolite-coated heat exchanger by varying operating conditions. Chai et al. [7] presented a desiccant heat pump by combining a heat pump and a desiccant-coated heat exchanger. They employed a commercial silica gel to fabricate the desiccant-coated heat exchanger. Liang et al. [8] developed a desiccant-coated heat exchanger by using a microchannel heat exchanger and explored the performance under various operating conditions.

Many investigations also have been carried out mathematically and experimentally to study the effectiveness and COP of various dew-point evaporative coolers (DECs). Xu et al. [9] proposed a novel complex heat and mass exchanger (HMX) to improve the performance of a dew-point evaporative cooler. Duan et al. [10] explored the operational performance of a regenerative evaporative cooler under various conditions by analyzing the wet-bulb effectiveness and COP. Balyani et al. [11] conducted an economic, environmental, and thermodynamic analysis for the best cooling technology for various climates. Duan et al. [10] investigated experimentally the energy-saving potential of a prototype of a counter-flow regenerative cooler when it applied to China’s various regions. Jafarian et al. [12] developed GMDH-type neural networks for modeling and optimization of a flat plate dew-point counter flow IEC. Their results indicated that the COP was improved by 36.6% in hot and dry climate. Sohani et al. [13] conducted a multiobjective optimization to find the best design for two different dew-point evaporative coolers (i.e., counter-flow and cross-flow configurations).

In this study, firstly, we present an innovative decoupling cooling technology where latent cooling load and sensible cooling load are handled separately by a desiccant-coated heat exchanger (DCHE)-based dehumidifier and a dew-point evaporative cooler (DEC). Secondly, the performance is investigated numerically by analyzing the heat and mass transfer for both DCHE and the DEC.

2. De-coupling cooling system

In order to achieve a large reduction in the kW/RT of cooling, we propose a de-coupling cooling system that thermodynamically hybridizes a desiccant coated heat exchanger (DCHE)-based dehumidifier with a dew-point evaporative cooler (DEC). The decoupling cooling system handles a sensible cooling load and a latent cooling load separately. Figure 1 shows schematically the working principle of the decoupling system and its thermodynamic process on a psychrometric chart. The DCHE first removes the undesired moisture of humid outdoor air by the adsorption process. During this process, the desiccant adsorbs water vapor up to its equilibrium state, which is determined by the water vapor partial pressure and the temperature of the desiccant. At this point, the air temperature rises slightly due to adsorption heat generation. Cooling water (∼25°C) is supplied to remove the adsorption heat. Subsequently, the DEC cools down the dehumidified air up to the desired temperature, maintaining moisture level. The DEC consists of two different channels, namely a dry air channel and a wet-air channel. The dehumidified air coming out from the DCHE enters the dry air channel and releases the heat to the adjacent wet channel where a certain portion (20–35%) of the dehumidified air flows. The inner surface of the wet channel is always wetted by water film, and the evaporation takes place on the surface of the water film by absorbing the heat from itself, which causes the reduction of the temperature. It is noteworthy that the proposed system can handle all climate conditions successfully as well as it employs only waste heat for the regeneration of adsorbent, and the IEC uses only water as a refrigerant.

Figure 1.

The working principle of the decoupling cooling system and its thermodynamic process on a psychrometric chart [2, 14].

2.1 Desiccant-coated heat exchanger-based dehumidification

The desiccant-coated heat exchanger (DCHE) has been tested successfully for dehumidification application using only the low-temperature waste heat, as proven in the many case studies [2, 5, 15, 16, 17]. It could remove effectively moisture from humid air at a relative humidity ratio of 95% or a humidity ratio of above 20 g/kg using commercial adsorbents such as the silica gel, Zeolite, and MOF.

The DCHE was first introduced in order to improve the efficiency of a conventional solid desiccant dehumidifier, where solid desiccant was packed within a finned tube heat exchanger using a wire mesh. It was found that the packing method is inefficient due to poor heat transfer between solid desiccant and cooling water flowing through the heat exchanger. To coat the fins of the heat exchanger with the desiccant powder (ex. silica gel), hydroxyethylcellulose (HEC) is used as a binder since it exhibits the best binding force between the silica gel powder and the metal fins as well as it does not affect the adsorbate uptake capacity. During the dip-coating process, the heat exchanger is mounted onto a rotation machine to ensure a uniform coating layer. After that, the heat exchanger is placed into an oven for curing at 120°C for 12 h. The optimal coating thickness of the desiccant layer is found to be 0.1 mm in many studies [2, 14, 17, 18, 19, 20]. As depicted in Figure 2, the DCHE operates alternatively between two different modes, namely dehumidification (adsorption) mode and regeneration (desorption) mode. During the dehumidification mode, cooling water is supplied through the tubes in order to remove the adsorption heat, keeping the desiccant layer at a lower temperature and thus ensuring the highest uptake. On the other hand, hot water is supplied to release the water vapor that was adsorbed in the previous process during the regeneration mode. In this mode, each process continues until the uptake reaches the equilibrium uptake given the temperature and vapor pressure.

Figure 2.

The concept of the desiccant-coated heat exchanger (DCHE) for the dehumidification: (a) adsorption/dehumidification process and (b) desorption/regeneration process.

2.2 Dew-point evaporative cooler (DEC)

As one of evaporative cooling methods, the dew-point evaporative cooler (DEC) is an energy-efficient and environmentally friendly sensible cooling device that employs only water as a refrigerant [21, 22, 23, 24]. The key feature of the DEC is that it employs the process air as the working air by purging the process air at a specified rate (20–30%) to reduce the supply air temperature below the wet-bulb temperature. The working principles for three different evaporative coolers are represented in Figure 3.

Figure 3.

The working principles of evaporative cooling devices: (a) direct evaporative cooler, (b) indirect evaporative cooler, and (c) dew-point evaporative cooler.

The DEC differs from the conventional adiabatic cooling or commonly known as the cooling towers, swarm coolers, etc., where the air stream experience changes in both the temperature and absolute humidity. In the DEC, however, the product air (conditioned air for space cooling) is flowing in a dry channel while a small fraction (20–30%) of the product air is purged into a wet channel, where the purged air picks up the water vapor from a hydrophilic membrane that physically separates the evaporative moist air flowing in a counter-flow direction from the product air. The evaporative cooling in the wet channel cools the air flowing in the adjacent dry channel by heat transfer, and the processes are repeated in succession to achieve an overall cooling of primary air. At the end of a section, the saturated secondary air is purged while a part of the conditioned primary air (but at the original inlet humidity) is directed into the wet channel to repeat the evaporative cooling. Thermodynamically, it incurs two major energy losses to achieve cooling of the primary air stream: Firstly, energy is transferred to the air by a fan for maintaining the necessary air flows in all channels. Secondly, it continuously purges a fraction of the conditioned air of the non-wetted channel so as to reinitiate the evaporative cooling in successive wetted channels. The key feature here is that it continuously purges a fraction of the conditioned air so that the wet-bulb temperature of the conditioned stream can be lowered. The novelty of the decoupling cycle is the harnessing of evaporative potential of inlet air as well as waste heat, but the overall electricity incurred is a quantum lower than that of conventional chillers.

3. Numerical modeling of a decoupling cooling system

3.1 Latent cooling by DCHE

In order to investigate the heat and mass transfer for the DEC, a mathematical model was established. The differential control volume includes half the height of the dry and wet channels, the separating plate, and the water film. To simplify the heat and mass transfer processes, the following assumptions were made:

1. The air streams in both the dry and wet channels are at a steady state.

2. Air flow is laminar and fully developed in both channels.

3. The heat and mass transfer coefficients are constant.

4. Heat losses to the surroundings are negligible.

5. The water is continuously supplied to maintain the water film.

Mass transfer between the adsorbent (desiccant) and adsorbate (water vapor) can be expressed by LDF (linear driving force) model as follows:

$$\frac{\mathit{dq}}{\mathit{dt}}=k\left(q^{\ast }-q\right)$$E1

where dq/dt is the rate of the uptake of water vapor, q* is the equilibrium uptake. k is a reaction coefficient and can be obtained by

$$k=\frac{15D_{\mathit{so}}\mathit{exp}\left(-\frac{E_a}{R{T}_d}\right)}{R_p^2}$$E2

where D_so a diffusion coefficient and E_a is an activation energy, R is an universal gas constant, T_d is the desiccant temperature, R_p is an average radius of desiccant particles, respectively. These values are listed in Table 1. The equilibrium uptake, q*, can be approximated with Tóth model.

Operation parameters	Baseline	Values
Process air inlet dry-bulb temperature [°C]	35	27, 30, 40, 46
Process air inlet wet-bulb temperature [°C]	24	19, 22, 24, 26
Process/regeneration air mass flow rate [kg/h]	13
Cooling water inlet temperature [°C]	30	25, 30, 35
Hot water inlet temperature [°C]	80	50, 60, 70, 80
Cycle time [sec.]	150

Table 1.

Simulation conditions for DCHE.

$$\frac{q^{\ast }}{q_0}=\frac{K_T\mathit{\text{Pexp}}\left(\frac{Q_{\mathit{st}}}{\mathit{RT}}\right)}{{\left\{1+{\left[K_T\mathit{exp}\left(\frac{Q_{\mathit{st}}}{\mathit{RT}}\right)\right]}^t\right\}}^{1/t}}$$E3

where q₀ is the maximum uptake of the desiccant and can be obtained by the measured isotherm. Q_st is the isosteric heat of the desiccant, t is heterogeneity factor, KT is Tóth constant, P is the vapor partial pressure.

The energy conservation equation of DCHE during adsorption/desorption process can be expressed by the following equation:

$$\left[M_d{c}_{p,d}+M_{\mathit{hx}}{c}_{p,\mathit{hx}}+M_d{c}_{p,v}q\left(t\right)\right]\frac{d{T}_d}{\mathit{dt}}=\begin{array}{l}{M}_d\frac{\mathit{dq}}{\mathit{dt}}{Q}_{\mathit{st}}-{\dot{m}}_a{c}_{p,a}\left(T_{a,o}-T_{a,i}\right)\\ -{\dot{m}}_w{c}_{p,w}\left(T_{w,o}-T_{w,i}\right)\end{array}$$E4

The first term on the left-hand side represents the rate of the change of energy content of the desiccant. The first term on the right-hand side represents the rate of adsorption heat generation. The second and the third terms are the rates of energy of air and water streams, respectively. Mass conservation equation of adsorbed phase of water vapor in the desiccant can be expressed by the following equation:

$$M_d\frac{\mathit{dq}}{\mathit{dt}}={\dot{m}}_{v,\mathit{\text{in}}}-{\dot{m}}_{v,\mathit{\text{out}}}$$E5

The first term on the left-hand side represents the rate of the change of moisture content in the desiccant. ${\dot{m}}_{v,\mathit{\text{in}}}$ and ${\dot{m}}_{v,\mathit{\text{out}}}$ are the mass flow rate of the moisture in the air at the inlet and outlet of DCHE, respectively. The humidity ratio of the air at the outlet can be calculated by Eq. (6).

$${\omega }_o=\frac{{\dot{m}}_{v,o}}{{\dot{m}}_{\mathit{da}}}$$E6

where ${\dot{m}}_{\mathit{da}}$ is the mass flow rate of dry air. It should be noted that the mass flow rate of dry air remains constant during the dehumidification and regeneration processes.

The temperature of the process air at the outlet of DCHE is obtained by the following equation:

$$T_{a,o}=\left(T_d-T_{a,i}\right)\mathit{exp}\left(-\frac{A_{c,a}{h}_{c,a}}{{\dot{m}}_a{c}_{p,a}}\right)$$E7

$$T_{w,o}=\left(T_d-T_{w,i}\right)\mathit{exp}\left(-\frac{A_{c,w}{h}_{c,w}}{{\dot{m}}_w{c}_{p,w}}\right)$$E8

In order to develop the temperature and velocity profile along the channel, Nusselt number is approximated by the following equation [25].

$$\mathit{Nu}={\mathit{Nu}}_{\mathit{FD}}+\frac{0.0841}{0.002907+G{z}^{-0.6504}}$$E9

$$\mathit{Gz}=\frac{\mathit{\text{RePr}}{P}_e}{4L_w}$$E10

$$\mathit{Re}=\frac{\mathit{\rho u}{D}_h}{\mu }$$E11

$$\mathit{Pr}=\frac{\mu c_p}{k}$$E12

The convective heat transfer coefficient for air side is calculated by Dittus-Boelter correlation:

$$h_a=\frac{N_{u,a}{k}_a}{D_{h,a}}$$E13

The same correlation can be applied to predict the convective heat transfer coefficient for water side of the DCHE

$$h_w=\frac{N_{u,w}{k}_w}{D_{h,w}}$$E14

The cooling/heat water flow inside the tube at a mass flow rate of 4 LPM and thus turbulent flow is formed with Reynolds number greater than 2300. Therefore, the Nusslet number is calculated from an empirical relation,

$$N{u}_w=0.023{\mathit{Re}}^{0.8}{\mathit{Pr}}^n$$E15

The exponent n equals to 0.4 for heating the water bulk flow (regeneration) and 0.33 for cooling conditions (dehumidification).

3.2 Sensible cooling by DEC

The mathematical model was for a differential control volume including half the height of the dry and wet channels, the separating plate, and the water film. To simplify the heat and mass transfer processes, the following assumptions were made:

1. The air streams in both the dry and wet channels are at steady state.

2. The air streams are fully developed laminar flow in both channels.

3. The heat and mass transfer coefficients are constant.

4. Heat losses to the surroundings are negligible.

5. The water is continuously supplied to maintain the water film.

In the dry channel of DEC, only a sensible cooling effect takes place by forced convective heat transfer without changing the humidity ratio. The energy conservation equation for air flowing in the dry channel is expressed by

$$k_a\frac{d^2{T}_{\mathit{da}}}{d{x}^2}-u_{\mathit{da}}{\rho }_{\mathit{da}}{c}_{p,m}\frac{\partial T_{\mathit{da}}}{\mathit{dx}}=\frac{h_{\mathit{da}}}{\overline{H}}\left(T_{\mathit{da}}-T_p\right)$$E16

where c_p,m is the specific heat of air-water vapor mixture. $\overline{H}$ represents half the height of the dry channel. The first term on the left-hand side of the above equation represents the longitudinal heat transfer by conduction, and the second term denotes heat transfer due to advection. The term on the right-hand side of the equation indicates convective heat transfer between the air and the separator.

Unlike the dry channel, both sensible and latent cooling effects take place in the wet channel. That is, both heat and mass transfer mechanism are involved to create cooling effect between the air and water film layer, and it is expressed as follows:

$$\begin{array}{c}{k}_a\frac{d^2{T}_{\mathit{wa}}}{d{x}^2}+u_{\mathit{da}}{\rho }_{\mathit{da}}{c}_{p,m}\frac{d{T}_{\mathit{wa}}}{\mathit{dx}}=\frac{h_{\mathit{wa}}}{\overline{H}}\left(T_{\mathit{wa}}-T_f\right)+\frac{h_m{\rho }_{\mathit{da}}}{\overline{H}}\left[{\omega }_{\mathit{sat}}\left(T_f\right)\right.\\ -{\omega }_v{c}_{p,v}\left(T_{\mathit{wa}}-T_f\right)\end{array}$$E17

The first term on the left-hand side represents conduction heat transfer in the longitudinal direction, and the second term denotes advection heat transfer. The first term on the right-hand side refers to convective heat transfer, and the second term is heat transfer due to mass transfer between the air and the water film surface.

The mass transfer takes place only in the wet channel by the driving force of vapor partial pressure difference, and the mass conservation equation is written as

$${\rho }_{\mathit{da}}{D}_v\frac{d^2{\omega }_v}{d{x}^2}=-{\rho }_{\mathit{da}}{u}_{\mathit{wa}}\frac{d{\omega }_v}{\mathit{dx}}+h_m\overline{H}{\rho }_{\mathit{da}}\left({\omega }_{v,\mathit{sat}}\left(T_f\right)-{\omega }_v\right)$$E18

On the other hand, the energy balances for the water film and the separating plate are given as

$$k_f\frac{d^2{T}_f}{d{x}^2}=-\frac{h_w\left(T_{\mathit{wa}}-T_f\right)}{{\delta }_f}-k_f\frac{T_p-T_f}{{\delta }_f^2}+\frac{h_{\mathit{fg}}{h}_m{\rho }_{\mathit{da}}\left[{\omega }_{\mathit{sat}}\left(T_f\right)-\omega \right]}{{\delta }_f}$$E19

$$k_p\frac{d^2{T}_p}{d{x}^2}=-\frac{h_d\left(T_{\mathit{da}}-T_p\right)}{{\delta }_p}+\frac{k_p\left(T_p-T_f\right)}{{\delta }_p^2}$$E20

Given the geometry of the channels and the operating conditions, the air streams are laminar flow with the maximum Reynolds number of 1048. The convective heat transfer coefficients for the dry and wet channels are expressed by Eq. (21).

$$h_{\mathit{wa}}=\frac{N_{u,\mathit{wa}}{k}_{\mathit{wa}}}{D_{h,\mathit{wa}}}$$E21

The Nusselt number can be calculated by the correlation by Eq. (8) for fully developed laminar flow in the rectangular dry channel [26].

$$\mathit{Nu}=8.235\left(1-2.042\mathit{AR}+3.085{\mathit{AR}}^2-2.477A{R}^3+1.058A{R}^4-0.186A{R}^5\right)$$E22

where AR is the aspect ratio of the channel (i.e., the ratio of the minimum to the maximum dimensions). The Nusselt number in the wet channel, in which both heat and mass transfer processes are involved, can be calculated by the following Eq. (25).

$$\mathit{Nu}=0.10{\left(\frac{l_e}{l}\right)}^{0.12}R{e}^{0.8}{\mathit{Pr}}^{1/3}$$E23

where l_e is the characteristic length, l_e = V/As and, V is the volume occupied by the wick material or water film. As is the wetted surface area of the channel, l is the total thickness of channel plate including the water film.

The convective mass transfer coefficient for the wet channel is approximated using heat and mass transfer analogy and can be expressed by Lewis number (Table 2).

	Parameter	Values
DCHE	Height of heat exchanger (mm)	100
	Length of heat exchanger (mm)	25
	Width of heat exchanger (mm)	100
	Pitch of fins (mm)	1.8
	Thickness of the coating layer (mm)	0.1
DEC	Channel length (m)	1.2
	Channel width (m)	0.3
	Channel height (mm)	4
	Thickness of the separating plate(mm)	0.1
	Thickness of the water film(mm)	0.4

Table 2.

Design parameters of DCHE and DEC for numerical study.

$$\frac{h}{h_m}={\rho }_a{c}_{p}L{e}^{\frac{2}{3}}$$E24

4. Numerical method and performance indices

4.1 Numerical method

The heat and mass transfer analysis for the decoupled cooling system has been carried out using the developed mathematical models. Variable-order method (ode15s) was employed for the transient simulation of DCHE and fourth-order method (bvp4c) for the steady-state simulation of DEC in MATLAB platform.

At first, simulation has been carried out for DCHE and examined the output states of the process air, namely the dry-bulb temperature and humidity ratio. The cyclic average values are then calculated and applied to the inlet conditions for DEC simulation. Simulation conditions were determined on the basis of IISO 5151:2017, which defines outdoor temperatures for three different climates, namely moderate climate, cool climates, and hot climates [27]. As shown in Table 1, moderate climate conditions (T_db = 35°C T_wb = 24°C) were considered as the baseline, and the cool climates are adopted for the lower limit condition and hot climates for the upper limit conditions.

4.2 Performance indices

The performance of DCHE was examined by analyzing two different metrics, namely moisture removal capacity (MRC) and latent cooling capacity (Q_L) and coefficient of performance (COP_th). MRC is defined as the total amount of water vapor that is absorbed by the desiccant during the dehumidification process, and Q_L is defined as the cooling power to remove the moisture. They can be calculated by the following equation:

$$\mathit{MRC}=\dot{{\int }_{t_0}^t\mathit{dv}\mathit{dt}}$$E25

$$\mathit{dv}={\dot{m}}_{\mathit{da}}\left({\omega }_{\mathit{\text{in}}}-{\omega }_{\mathit{\text{out}}}\right)$$E26

$$Q_l=\dot{H_v\times \mathit{dv}}$$E27

COP_th is defined as the ratio of MRC to the total amount heat that is consumed by the desiccant during the regeneration process and can be calculated by the following equation:

$${\mathit{COP}}_{\mathit{th}}=\frac{Q_l}{Q_{\mathit{reg}}}$$E28

$$Q_{\mathit{reg}}={\dot{m}}_{\mathit{hw}}(T_{\mathit{hw},\mathit{\text{in}}}-T_{\mathit{hw},\mathit{\text{out}})}$$E29

For the evaluation of DEC, the temperature of product air (T_da,out), sensible cooling capacity (Q_S), wet-bulb effectiveness (ε_wb), and dew-point effectiveness (ε_dp) are analyzed judiciously, which can be calculated by the following formulas:

$$Q_S=\left(1-\gamma \right){\dot{m}}_{\mathit{da}}(T_{\mathit{da},\mathit{\text{in}}}-T_{\mathit{da},\mathit{\text{out}})}$$E30

$${\epsilon }_{\mathit{wb}}=\frac{T_{\mathit{da},\mathit{\text{in}}}-T_{\mathit{da},\mathit{\text{out}}}}{T_{\mathit{da},\mathit{\text{in}}}-T_{\mathit{wb}}\left(T_{\mathit{da},\mathit{\text{in}}}\right)}$$E31

$${\epsilon }_{\mathit{db}}=\frac{T_{\mathit{da},\mathit{\text{in}}}-T_{\mathit{da},\mathit{\text{out}}}}{T_{\mathit{da},\mathit{\text{in}}}-T_{\mathit{db}}\left(T_{\mathit{da},\mathit{\text{in}}}\right)}$$E32

5. Results and discussions

5.1 Validation

The developed numerical models for DCHE and DEC have been validated with our experimental results and the published experimental results from references [28]. To evaluate the accuracy of mathematical models, relative error percentage and root-mean-square error (RMSE) were used for the consistency between simulated results and experimental data, and they are expressed as [29]

$${\mathrm{E}}_{\mathrm{r}}=\left(\frac{X_{\mathit{sim},}-X_{\mathit{exp},i}}{X_{\mathit{exp},i}}\right)\times 100$$E33

$$\text{RMSE}=\sqrt{\frac{\sum {\left[\frac{X_{\mathit{sim},i}-X_{\mathit{exp},i}}{X_{\mathit{exp},i}}\right]}^2}{n}}$$E34

Figure 4 shows the comparison of the transient humidity variation with time during one cycle between the simulation results and the experimental results under the specific inlet air conditions of 32.6°C temperature with 14.0 g_v/kg_da humidity ratio. As shown in Figure 4a, the outlet humidity ratio obtained from the numerical analysis is quite similar to the experimental result. It is also observed from Figure 4b that the highest relative error is 8.6% with an RMSE of 0.0261.

Figure 4.

Validation of the developed the mathematical model for DCHE: (a) transient outlet humidity ratio, (b) fitting of outlet humidity ratios in numerical and experimental results.

Figure 5 shows the comparison of the results obtained from the numerical analysis with the experimental results obtained from reference [23]. It also shows a good agreement between the simulation and experimental data under different inlet temperatures. The maximum deviation of 3.2% is observed with an RMSE of 0.0159.

Figure 5.

Validation of the developed the mathematical model for DEC with experimental data from reference [28]: (a) temperature distributions along the channel under specific primary air conditions, (b) fitting of product air temperature in numerical and experimental results.

5.2 Performance of DCHE

The performance of the DCHE was numerically analyzed, and simulation results are judiciously compared through a series of runs under various operating conditions. The inlet air temperature is 35°C for both dehumidification and regeneration processes. The temperatures of cooling and hot water are set at 30°C and 50°C, respectively. During the entire process, the inlet humidity ratio is kept to 14 g/kg. Figure 6 shows the dynamic variation of the performance of DCHE during three cycles under the baseline conditions where. It is easily seen from Figure 6a that the equilibrium uptake reaches its saturated state at 0.33 during the dehumidification process and at 0.12 during the regeneration process in a given condition. It is also seen that the actual uptake follows the trajectory of the equilibrium uptake at a different rate. It can also be observed from Figure 6b that the outlet humidity ratio reaches its equilibrium state at a faster rate during the regeneration process. However, it is noteworthy that the amount of the vapor that is released and adsorbed by the desiccant is conserved during the cycle so that the system continues to operate at a cyclic steady state. The lowest humidity ratio of 0.0092 kg_v/kg_da was observed at 16 s after the dehumidification process starts, and then it increases gradually up to its saturated level. It is also worthy to note that the desorption rate is a function of the adsorbent’s surface temperature. Therefore, as the surface temperature rises when the supply of hot water flows through the heat exchanger, the more adsorbates (water vapor) are released at faster rate. Hence, the regeneration time becomes shorter at higher surface temperature. This explains the reason for supplying the cooling water to the heat exchanger to enable a cooler adsorbent that is capable of adsorbing more moisture than the desorbing process.

Figure 6.

Dynamic variations of the performance of the DCHE under the baseline conditions (T_db = 35°C, T_wb = 24°C,${\dot{m}}_a$ =13 kg/h, T_cw = 30°C, T_hw = 50°C, t_cycle = 150 s): (a) the variations of instantaneous uptake and the equilibrium uptake; (b) the variation of the air humidity ratio at the outlet of DCHE.

Figure 7 shows the effect of change in the dry-bulb temperature of the process air on the performance of DCHE, namely moisture removal capacity (MRC), COP_th, latent cooling capacity (Q_l), and regeneration energy (Q_reg). The dry-bulb temperature varied from 27 to 46°C while the other parameters were kept constant as in the baseline conditions. It is observed that MRC, Q_l, and Q_reg are markedly affected by process air inlet dry-bulb temperature, whereas COP_th remains constantly around 0.125. MRC drops by 23.6% from 1.867 (T_db = 35°C) to 1.42 (T_db = 46°C). This is mainly because the humidity ratio decreases from 0.0142 to 0.0096, which implies the air carries less water vapor at a higher temperature when wet-bulb temperate is fixed at constant.

Figure 7.

Effect of change in the process air dry-bulb temperature on the performance of DCHE: (a) moisture removal capacity (MRC) and COP_th; (b) latent cooling capacity (Q_l) and regeneration energy (Q_reg).

Figure 8 shows the effect of change in the wet-bulb temperature of the process air on the performance of DCHE. Unlike the previous results, the performance indices are raised when increasing the wet-bulb temperature except for COP_th. The reason for this is that the water vapor partial pressure increases as the dry-bulb temperature rises when the wet-blub temperature is fixed. Hence, the equilibrium uptake of the desiccant increases, which is defined as the ratio of the vapor partial pressure to the saturation pressure at the desiccant temperature. As a result, more water vapors can be absorbed by the desiccant. The highest values for MRC, Q_l and Q_reg were 1.87 g/cycle and 30.1 W, respectively when T_wb = 24°C. It can be observed that MRC, Q_l, and Q_reg decrease slightly at T_wb = 26°C.

Figure 8.

Effect of change in the process air wet-bulb temperature on the performance of DCHE: (a) moisture removal capacity (MRC) and COP_th; (b) latent cooling capacity (Q_l) and regeneration energy (Q_reg).

5.3 Performance of DEC

Figure 9 depicts the temperature and humidity ratio profiles of the product air and the working air for DEC under a specified operating condition. The product air flows along the channel length and the working air flow reversely in a counter-flow manner. It is observed from Figure 9b that the product air is purged at the end of the channel and is flowed back to the inlet. The product air temperature drops sensibly from 26°C to 16.18°C while the working air temperature increases from 16.8 to 24°C. It can be found that the temperature drops significantly within a distance of 0.1 m from the end of the channel immediately after being purged. This is because that instantaneous evaporation takes place from the surface of the water film by the continuously purged air. As depicted n Figure 9b, the relative humidity surges rapidly from 0.4 to 0.98 in the immediate vicinity of the exit. In addition, once the actual humidity ratio reaches the saturation level, both humidity ratios increase gradually, which allows for water to evaporate continuously along the channel length.

Figure 9.

Temperature profile (a) and humidity ratio profile (b) of each component in DEC.

Figure 10 depicts the effect of change in the air temperature and humidity ratio on the performance of DEC while other parameters are kept constant as in the baseline conditions. The dry-bulb temperature varies from 26 to 38°C, and the humidity ratio varies from 0.005 kg_v/kg_da to 0.0187 kg_v/kg_da respectively. It can be found that the performance indices are almost linearly dependent on the air dry-bulb temperature and the humidity ratio. For the dew-point effectiveness, it varies between 0.44 and 0.57 when the dry-bulb temperature varies from 26 to 38°C. It can be inferred that the higher the air temperature is, the higher the dew-point effectiveness. It should be noted that the higher dew-point effectiveness and the wet-bulb effectiveness can be expected at a higher humidity ratio, whereas higher cooling capacity can be obtained at a lower humidity ratio. This is because both wet-bulb and dew-point temperatures increase in proportion to the humidity ratio when the dry-bulb temperature is fixed. Therefore, the potential for water evaporation is declined, and thus, the temperature drop in the dry channel gets closer to its maximum (i.e., 1).

Figure 10.

Effect of change in the air temperature and humidity ratio on the performance of DEC: (a) dew-point effectiveness, (b) wet-bulb effectiveness, and (c) cooling capacity.

6. Conclusions

This study presents an innovative decoupling cooling technology where latent cooling load and sensible cooling load are handled separately by a desiccant-coated heat exchanger (DCHE)-based dehumidifier and a dew-point evaporative cooler (DEC). The DCHE first removes the undesired moisture of humid outdoor air by adsorption process. Subsequently, the DEC sensibly cools down the dehumidified air up to the desired temperature, maintaining the moisture level. Their performances are investigated numerically by analyzing the heat and mass transfer. Simulation has been carried out for DCHE and examined the output states of the process air, namely the dry-bulb temperature and humidity ratio. The following findings can be inferred from this study:

1. The equilibrium uptake reaches its saturated state at 0.33 during the dehumidification process and at 0.12 during the regeneration process at a given condition. The actual uptake follows the trajectory of the equilibrium uptake at a different rate.

2. MRC and Q_l of the DCHE were markedly affected by varying the dry-bulb temperature and wet-bulb temperature, whereas COP_th remained constant around 0.125. The MRC drop by 23.6% at higher temperature was observed due to the decreases in the humidity ratio.

3. As the wet-bulb temperature of the process air increases, the improved MRC and Q_l of the DCHE were observed due to the increases in the humidity ratio.

4. For the DEC, dew-point effectiveness, wet-bulb effectiveness, and cooling capacity were linearly dependent on the dry-bulb temperature of the process air as well as the humidity ratio.

5. The higher dew-point effectiveness and the wet-bulb effectiveness can be expected at a higher humidity ratio, whereas higher cooling capacity can be obtained at a lower humidity ratio.

The proposed decoupling system has no moving parts and harmful refrigerant, rendering less maintenance compared with an existing cooling system. Furthermore, it is an energy-efficient means of latent and sensible cooling by adsorption process and water evaporation process with a waste heat source as compared with other conventional air-conditioning processes.

Acknowledgments

This work was supported by the National Research Council of Science & Technology (NST) grant by the Korea government (MSIT) (No. CPS21171-110).

Nomenclature

c_p

specific heat at constant pressure, J/(kg K)

D_v

diffusion coefficient, m²/s

h

heat transfer coefficient, W/(m² K)

h_fg

latent heat evaporation, J/kg

h_m

mass transfer coefficient, m/s

H̅

half height of channel, m

k

thermal conductivity, W/(m K)

L

channel length, m

ṁ

mass flow rate, kg/s

P

pressure, Pa

Q̇

heat transfer rate, W

r

working air ratio

T

temperature, °C

u

velocity, m/s

W

channel width, m

x

x-coordinate, m

Greek symbols

δ

thickness, mm

ρ

density, kg/m³

ω

humidity ratio, kg/kg dry air

Subscripts

a

air

da

dry air

dp

dew point

f

water film

i

inlet

m

water vapor-air mixture

p

plate

pa

primary air

s

supply

sat

saturated

v

water vapor

w

wet air