An experimental and analytical study of a hybrid air-conditioning system in buildings residing in tropics

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An experimental and analytical study of a hybrid air-conditioning system in buildings residing in tropics X. Cui a,b, M.R. Islam b,c, K.J. Chua b,c,∗ a

Institute of Building Environment and Sustainable Technology, School of Human Settlements and Civil Engineering, Xi’an Jiaotong University, Xi’an, Shaanxi 710049, China b Department of Mechanical Engineering, National University of Singapore, 9 Engineering Drive 1, Singapore 117576, Singapore c Engineering Science Programme, National University of Singapore, 9 Engineering Drive 1, Singapore 117575, Singapore

a r t i c l e

i n f o

Article history: Received 26 March 2019 Revised 26 May 2019 Accepted 11 June 2019 Available online xxx Keywords: Air-conditioning Heat recovery Indirect evaporative cooling Cooling load Energy consumption

a b s t r a c t A hybrid evaporative pre-cooled air-conditioning system has been proposed for operation in humid tropical climatic conditions. The indirect evaporative heat exchanger (IEHX) works in tandem with a conventional air handling unit (AHU). The IEHX operates as an evaporative-enhanced heat recovery device to treat the ambient intake air. An experimental study has been carried out to systematically investigate the performance of a lab-scale IEHX when combined with a chiller under steady-state conditions. In addition, a mathematical model is judiciously developed for both the IEHX and cooling coil. Key results have indicated that the IEHX is capable of pre-cooling and pre-dehumidifying the ambient intake air. In addition, the pre-cooling process facilitates the supply of a higher chilled water supply temperature to meet cooling needs particularly at higher outdoor air fractions. Finally, the hybrid evaporative pre-cooled air-conditioning system is evaluated in terms of its energy consumption performance based on a design day weather condition. © 2019 Elsevier B.V. All rights reserved.

1. Introduction Heating, ventilation, and air-conditioning systems constitute a significant percentage of the total energy consumption in buildings [1]. Considering the ongoing global energy and environmental issues, it is currently a key challenge to evolve new sustainable air-conditioning systems [2,3]. As one of the energy-efficient cooling techniques, the interest in indirect evaporative cooling (IEC) system is gaining traction [4]. The evaporative cooling system is able to achieve cooling effect by taking the advantage of water evaporation [5]. The key feature of IEC is based on the key principle that the water evaporation process absorbs a large amount of latent heat [6]. As a result, IEC systems are deemed to consume less energy compared to conventional mechanical compression systems [7]. Recently, researchers continue to pursue experimental and numerical studies on evaporative coolers [8–12]. Comino et al. [13] developed a simplified model for IEC systems by validating against experimental measurements. Tariq et al. [14] numerically investigated the impact of alumina nanoparticles on the improvement of cooling effectiveness. Comparisons were performance by

∗

Corresponding author. E-mail address: [email protected] (K.J. Chua).

considering the improvement on cooling effectiveness, cooling capacity, and energy efficiency ratio. Baakeem et al. [15] studied a counter-flow indirect evaporative heat exchanger (IEHX) based on the M-cycle air flow and reported that it was a convenient airconditioning approach applied in Arab Gulf Countries. Duan et al. [16] estimated the energy saving potential of a regenerative IEHX based on the climate data for several cities in China. They reported that the IEHX was able to achieve up to 58% reduction of the annual electrical consumption in the selected regions. Dizaji et al. [17] developed an analytical model for an IEHX with perforated regenerative configuration. Zhu et al. [18] studied the regenerative IEC based on a data-driven model. The artificial neural network algorithm was employed to calculate the air temperature variation. Pandelidis et al. [19,20] presented a numerical model for the comparative study on three types of IECs with different flow schemes. Simulation results showed that a higher cooling effectiveness was achieved by employing the IEC in tandem with a crossregenerative counter flow arrangement. Moshari and Heidarinejad [21] presented an analytical model to investigate the pressure drop of the IEC systems. They analyzed the impact of fin height on overall fan power consumption. Chen et al. [22] conducted experiments to investigate an IEC with the integration of hollow fiber. The evaporative cooling technique can be also applied together to other air-conditioning devices [23]. Boukhanouf et al. [24] com-

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Nomenclature

W

area [m2 ] thermal diffusivity [m2 /s] molar concentration [mol/m3 ] specific heat of moist air [kJ/(kg·°C)] diffusivity [m2 /s] convection heat transfer coefficient [kW/(m2 ·K)] mass transfer coefficient [kg/(m2 s)] specific latent heat of water evaporation [kJ/kg] enthalpy [kJ/kg] thermal conductivity [kW/(m·°C)] length of the channel [m] mass flow rate of air [kg/s] molar mass [kg/mol] Nusselt number pressure [kPa] Prandtl number heat flux [kW/m2 ] heat transfer rate [kW] ideal gas constant [J/(K·mol)] temperature [°C] velocity in x direction[m/s] velocity in y direction [m/s] humidity ratio [g moisture/kg dry air] mass transfer rate [kg/s]

Subscript 1 2 a dew in j l out s sat sur w

product air (primary air) working air (secondary air) air dew-point temperature inlet condensation latent outlet sensible saturated surface water

A

α c cpa D h hm hfg i k L m M Nu P Pr q Q R T u v

ω

Abbreviation RH Relative humidity IEC Indirect evaporative cooling IEHX Indirect evaporative heat exchanger PA Product air WA Working Air DBT Dry-bulb temperature WBT Wet-bulb temperature DPT Dew-point temperature

bined heat pipes with an IEHX comprising porous ceramic tubes. The experimental data indicated that the proposed IEHX design was able to achieve the wet-bulb cooling effectiveness of 0.8. To reduce the humidity of the air, a liquid desiccant unit was suggested to operate in tandem with the IEC unit in order to handle varying supply air volumes [25–27]. Ham and Jeong [28] evaluated the energy performance of a system integrated the IEC with a liquid-desiccant unit. The combined system was compared with a variable-air-volume system in summer conditions. The internally evaporative cooled air improved the dehumidification efficiency due to lower solution temperatures [29,30]. The desiccant-coated heat exchangers [31,32] and vacuum-based membrane [33] can be also employed to enhance the evaporative cooling performance [34]. In addition, the application of evaporative cooling to the con-

denser of a conventional vapor-compression was observed to enhance the coefficient of performance of chillers [35,36]. Since the wet-bulb temperature (WBT) is known to be the theoretical temperature limit for conventional IEHX, regions with hotarid climate have been considered as suitable places for the application of IEC systems [37]. In humid climates, however, a single IEHX is usually inadequate to obtain the desired supply air conditions due to the high WBT of the ambient air [38]. To further promote the application of IEC in humid climates, there are suggestions to employ the IEHX as a pre-cooling device before the conventional air handling unit (AHU) [39]. In addition, the heat recovery unit is able to reduce the cooling demand for the AHU [40]. When the working air of the IEHX employs the room exhaust air, the interface channel temperature can be lower than the dew point temperature (DPT) of the ambient intake air resulting in a possibility of condensation [41]. Therefore, the air conditioning process in the proposed system is deemed to be much more complex in humid operating conditions. The above literature review indicates that the focus of many previous studies is on the heat and mass transfer through the IEHX with various air flow arrangements. However, few attempts have been made to comprehensively evaluate the energy performance of the air conditioning process in a hybrid air conditioning system with an evaporative-enhanced heat recovery unit for the application in humid tropical climates. The present study aims to address this issue by investigating the thermal performance of a hybrid evaporative pre-cooled air-conditioning system. In addition, the present study investigates the influence of the IEHX on the chilled water temperature and the energy saving potential. The advantages of this hybrid system include: (1) capturing the cooling potential from the exhaust air; (2) achieving a higher driving force for the evaporative cooling; (3) pre-cooling the outdoor humid air in order to lower the cooling load for the conventional chiller; and (4) extending the application of evaporative cooling technique in humid climates. In this paper, we will first introduce the design of the hybrid air conditioning system with an evaporative-enhanced heat recovery unit. This is followed by the numerical modelling for the combined IEHX chilled-water-based cooling coil. The validated mathematical model is then adopted to investigate the pre-cooling performance of the evaporative-cooled assisted heat recovery and its impact on the AHU. Finally, the energy saving potential is estimated for the proposed hybrid air conditioning system.

2. Description of the hybrid evaporative pre-cooled air-conditioning system Fig. 1 presents the schematic of the proposed hybrid airconditioning system. The IEHX operates in tandem with the conventional AHU. As shown in Fig. 1, the pre-cooling IEHX consists of alternatively arranged dry channels and wet channels. In the wet channels, the working air directly interacts with the water film at the wet channel surface. In the alternative dry channels, the product air is able to transfer heat to the adjacent wet channels due to the evaporation process occurred in the working stream. For the proposed hybrid system, the IEHX is viewed from the perspective of an evaporative-enhanced heat recovery unit. In humid tropical climates, the DBT and RH of the outdoor air are much higher than the exhaust air from the conditioned room. The evaporative-enhanced heat recovery unit provides an efficient method for pre-cooling the intake air and recovering the cooling potential of the exhaust air. As a result of the large temperature difference between the outdoor air and the room exhaust air in the IEHX, the humid outdoor air may condense in the product channel especially when the interface DBT is lower than its inlet DPT.

Please cite this article as: X. Cui, M.R. Islam and K.J. Chua, An experimental and analytical study of a hybrid air-conditioning system in buildings residing in tropics, Energy & Buildings, https://doi.org/10.1016/j.enbuild.2019.06.028

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Fig. 1. Schematic of the hybrid evaporative pre-cooled air-conditioning system.

Fig. 2. Schematic diagram of the experimental setup for the IEHX. Table 1 The measured parameters and the specification of measuring instruments. Parameter

Instrument

Uncertainty

Temperature Relative humidity Velocity

Thermistor probe (TJ80-44,033–1/8–6) Humidity sensor (CLIMOMASTER)

± 0.1 °C 2% RH to 80% RH: ± 2.0% 80% RH to 98% RH: ± 3.0% ± 2.0% of reading

Hot-wire anemometer (CLIMOMASTER)

3. Experimental systems 3.1. Indirect evaporative cooler A lab-scale experimental setup has been developed for investigating the cooling performance of the IEHX. Fig. 2 shows the schematic diagram of the experimental setup. The intake air velocity was adjusted by a variable speed blower, while a heater was used to control the inlet air temperature. The measuring instruments were positioned at varying locations to collect the experimental data including the air temperature, humidity, and velocity. The specification of sensors are shown in Table 1.

Table 2 Specifications for the counter-flow regenerative IEHX. Dimension

Value

Unit

Product channel gap Working channel gap Channel length Channel width Wall thickness

10 6 750 300 0.3

mm mm mm mm mm

The experimental study was carried out for a prototype of a counter-flow regenerative IEHX. Table 2 indicates the dimensions for the prototype. To investigate the steady state of the IEHX, the

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Fig. 3. Experimental setup for the chiller.

Fig. 4. Schematic of the computation element of the IEHX.

experimental procedure is described as follows. Firstly, the wet channel of the IEHX was supplied with water until the wicking material was fully saturated. Secondly, the inlet air was adjusted to a desired condition. Thirdly, the measuring sensors started to record the experimental data after the expected operation condition was stabilized for 15 min. After the measured data had been recorded for a specific steady-state, the inlet air can be adjusted to another operation condition. 3.2. Chiller An experimental setup was also developed to study the impact of chilled water supply temperature (Tw,in ) on the performance of a water-cooled chiller. Fig. 3 shows the experimental setup. The major components included a water-cooled scroll chiller, pumps, refrigerant flow meters, pressure transducers, temperature sensors, water flow meters, and water mixing chambers [42]. The required inlet water conditions were produced by mixing the water from the evaporator, the condenser and the cooling tower in the water mixing chambers. The experimental data was collected for each steady state condition. The coefficient of performance (COP) of the setup was evaluated under varying Tw,in . 4. Mathematical formulation 4.1. Evaporative pre-cooling process A mathematical model is established for the IEHX to study the evaporative pre-cooling process. The computational domain is schematically shown in Fig. 4.

The key governing equations of the moist air stream under steady state conditions are given as

∂ ua ∂ va + =0 ∂x ∂y ua

1 dp ∂ ua ∂ ua ∂ 2 ua + va =− + va ∂x ∂y ρa dx ∂ y2

∂ ∂ ∂ 2 Ta (ua Ta ) + (va Ta ) = αa 2 ∂x ∂y ∂y

(1)

(2)

(3)

The diffusion equation for water vapor is expressed below:

ua

∂ ca ∂ ca ∂ 2 ca + va = Da ∂x ∂y ∂ y2

(4)

where Da and ca are the diffusivity and the molar concentration of water vapor, respectively. In the working wet channel, the water evaporation process enables the working air to absorb heat. The driving force for evaporation is in proportion to the gradient of water vapor concentration. At the water film interface, the thin layer of the moist air is assumed to be saturated so that its concentration can be computed as

cw =

Psat (Tw ) RTw

(5)

where Psat (Tw ) is the saturated vapor pressure at the absolute DBT of air-water interface. The interfacial condition at the working wet channel surface:

ua,2 = 0,

va,2 = 0

(6)

Please cite this article as: X. Cui, M.R. Islam and K.J. Chua, An experimental and analytical study of a hybrid air-conditioning system in buildings residing in tropics, Energy & Buildings, https://doi.org/10.1016/j.enbuild.2019.06.028

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 Q(i, j ) = ma ia(i, j ) − ia(i+1, j )

(15)

In addition, the total heat transfer can be further evaluated by considering the convective heat transfer inside/outside the cooling coil:

 Q(i, j ) = hi Ai Ts,m(i, j ) − Tw,m(i, j )

(16)

 Q(i, j ) = hoAoηs Ta,m(i, j ) − Ts,m(i, j )  + h f g hm Aoηs ωa,m(i, j ) − ωs,m(i, j )

(17) (17)

Since the Lewis factor is usually assumed to be unity for airwater mixtures, Eq. (17) can be expressed as



ηs

Q(i, j ) =

c pa

hoAo ia,m(i, j ) − is,m(i, j )



(18)

Fig. 5. Computational element for the cooling coil.

−kw

dTw dTa,2 = −ka − MH2 O h f g Da dy dy



∂ ca,2 ∂y

The convective heat transfer coefficient for the inner surface of the coil (hi ) is obtained based on the following correlation [46]:

 (7) w

In the product channel, the condensation process of the PA occurs depending on the surface temperature. When the product channel interface DBT is higher than the product air DPT, the air can be sensibly cooled without the issue of moisture condensation. As a result, the product channel interface temperature can be gradually reduced. When the interface temperature decreases below DPT, the condensation will occur in this region. Therefore, the interface water vapor concentration in the product channel can be defined as



csur = f (Tsur ) =

ca,1,in , Psat (Tsur ) , RTsur

Tsur ≥ Ta,1,dew Tsur < Ta, 1,dew

(8)

The condensation rate, mc , is given as



mc =

0 , Tsur  ≥ Ta,1,dew ∂ ca,1 ∂y

−MH2 O D

sur

, Tsur < Ta,1,dew

(9)

The interfacial boundary condition at the product channel surface is written as

ua,1 = 0, −ksur

va,1 = 0

dTsur = −ka dy



dTa,1 dy

1 + 12.7

8

The convective heat transfer coefficient for the outer surface of the coil (ho ) is calculated by the Colburn j-factor analogy [47]:

ho = jGa c pa P r −2/3

(20)

The arithmetic mean values between adjacent grid points (Tw, m(i, j) , Ts, m(i, j) , Ta, m(i, j) , is, m(i, j) , and ia, m(i, j) ) are employed to facilitate the calculation. For example, Tw, m(i, j) is given as

Tw,m(i, j ) =

Tw(i, j ) + Tw(i, j+1) 2

 is,m(i, j ) =





ia(i, j ) + c pa Ts,m(i, j ) − Ta(i, j ) , Ts,m(i, j ) > Tdew(i, j ) 2 10.76 + 1.4Ts,m(i, j ) + 0.046Ts,m (i, j ) , Ts,m(i, j ) ≤ Tdew(i, j )

(11)

Finally, by considering all the above equation, the calculation of the parameters for the next grid is given as

Tw(i, j+1) = Tw(i, j ) −

Q(i, j )

(12)

(13)

Ta(i+1, j ) =

(23)

mw c pw

Uo 1 − NT 2

NT Uo · Ta(i, j ) + · Ts,m(i, j ) Uo Uo 1 + NT 1 + NT 2 2

(24)

where

4.2. Cooling coil unit After the treatment of the pre-cooling IEHX, the supply air is further conditioned through the cooling coil in the conventional AHU. A mathematical model for the cooling coil was developed “row-by-row” [44,45]. Fig. 5 illustrates the computational domain for the calculation. The total heat transfer within a computational element for the chilled water and the air can be obtained as

Q(i, j ) = mw c pw Tw(i, j+1) − Tw(i, j )

(21)

The cooling coil surface condition (dry condition or wet condition) depends on the coil surface DBT. Therefore, the interface air enthalpy is determined as:

The outlet boundary conditions are expressed as

∂ ua ∂ va ∂ Ta ∂ ca = 0, = 0, = 0, = 0, Pa = 0 ∂x ∂y ∂x ∂x

(19)

P rw − 1

sur

va = 0, Ta = Ta,in , ca = ca,in



 D  23  (Rew − 10 0 0)Prw i   1 +  f 0 . 5 2 L 3

(22)

The inlet boundary conditions are given as

ua = ua,in ,

8

(10)

 + mc · h f g

f

N ui =



(14)

NT Uo =

η s h o A o

(25)

ma c pm

ia(i+1, j ) = ia(i, j ) −

ωa(i+1, j ) =

Q(i, j )

(26)

ma

ia(i+1, j ) − c pa · Ta(i+1, j )

2501 + 1.89 · Ta(i+1, j ) · 10 0 0

(27)

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Fig. 7. Correlation between the chilled water supply temperature and the COP of the chiller.

Fig. 6. Flow-chart for the computational process.

4.3. Computational process Fig. 6 illustrates a flow-chart depicting the computational process for both pre-cooling IEHX and cooling coil. Firstly, the governing equations for the pre-cooling IEHX are established and then solved by employing the COMSOL Multiphysics platform [43]. Two physical modules, namely Conjugate Heat Transfer module and the Transport of Diluted Species module, are coupled to investigate the fluid flow and the heat and mass transfer of the air streams. Secondly, the mathematical formulation for the cooling coil is solved in the MATLAB environment. The two-step computational process is able to determine the desired pre-cooling and supply air conditions. 5. Results and discussion 5.1. Experimental results from IEHX and chiller The performance of the regenerative IEHX prototype has been experimentally investigated under varying inlet conditions. Table 3 presents the experimental data. It can be seen that a larger temperature reduction of the PA was obtained for a higher inlet air DBT. For example, when the product air inlet DBT increases from 23.9 °C to 27.8 °C, the outlet DBT spans 19.4- 21.5 °C. In other words, the temperature reduction of PA increases from 4.5 °C to

6.3 °C. It is attributed to the fact that the air with a higher DBT is able to absorb more moisture before the dry air becomes saturated. According to the experimental data, the wet-bulb effectiveness of the IEHX was further calculated in Table 3. It is observed that the wet-bulb effectiveness presents a small variation under a specific intake air velocity. In addition, the average wet-bulb effectiveness decreases for a higher inlet velocity. For example, the average wet-bulb effectiveness of the IEHX is reduced from 0.70 to 0.52 by varying the velocity of PA from 1.5 m/s to 3.0 m/s. It is probably due to the reason that a lower intake air flow rate results in a longer contact duration between the two air streams. The influence of the chilled water temperature on the performance of the chiller has been experimentally investigated. Fig. 7 shows the relationship between the Tw,in and the COP. It can be seen that the elevated Tw,in results in an enhanced chiller efficiency. For example, the average COP is 3.39 for the Tw,in of 7 °C, while the average COP increases to 3.84 for the Tw,in of 11 °C. The experimental data pointed out the possibility to achieve energy savings by raising Tw, in the hybrid air-conditioning system. 5.2. Validation of mathematical modelling The validation of mathematical modelling is conceived in three parts: (1) air treatment in a conventional regenerative IEHX; (2) condensation process of the moist air; (3) performance of the cooling coil. The model is first validated based on the experimental data on the counter-flow regenerative IEHX. Fig. 8 shows the comparison between the calculated outlet DBT and the related experimental data. The maximum discrepancy is about 8% indicating the high precision of the model to estimate the cooling performance of conventional IEHX without condensation. The validation process is performed by considering the condensation process of the moist air. Previous studies have investigated the correlation between the moisture mass fraction difference and the heat flux. The present numerical model is validated against experimental results [48] and previous numerical simulation results [49]. Fig. 9 compares the simulated heat flux with the data acquired from literatures [48,49]. Several essential parameters (qj , qs , and W) are employed to illustrate the results. In Fig. 9, the definition of qj is the water vapor condensation heat flux to the wall surface, qs is the sensible heat flux, and W represents the water vapor mass fraction difference (W = Wa,in − Wsur ). The inlet air

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Table 3 Experimental data on the regenerative IEHX. Test

Product air velocity (m/s)

Inlet air humidity ratio (g/kg)

Inlet air DBT ( °C)

Product air outlet DBT ( °C)

Web-bulb effectiveness

1 2 3 4 5 6 7 8 9

1.50 1.50 1.50 2.00 2.00 2.00 3.00 3.00 3.00

10.00 10.00 10.00 10.00 10.00 10.00 10.00 10.00 10.00

23.9 25.7 27.8 22.8 26.4 28.7 24.1 28.3 36.2

19.4 20.5 21.5 18.9 21.4 22.9 20.5 23.8 28.6

0.72 0.69 0.70 0.69 0.62 0.61 0.57 0.48 0.51

Fig. 10. Temperature and humidity ratio profiles in the IEHX. Fig. 8. Validation 1: comparison of the outlet air DBT of a counter-flow regenerative IEHX.

DBT with the measured data. It indicates that the computational model is able to accurately predict the cooling coil’s performance with a discrepancy of less than ±9.2%. In sum, the validation of these three parts provide a solid foundation and justification to employ the model to further investigate the performance of the hybrid evaporative pre-cooled airconditioning system. 5.3. Pre-cooling performance of IEHX

Fig. 9. Validation 2: comparison of the moist condensation performance in different works including present numerical model, previous numerical study [49], and experimental data [48].

DBT was 313.15 K, and the wall surface DBT was 294.15 K. Clearly, as shown in Fig. 9, the present model demonstrates good agreement with previous works. Third, the model is further validated in terms of the ability for evaluating the air conditioning process through the cooling coil. An experimental study in literature [50] is employed to conduct the validation. The cooling coil performance is investigated under varying inlet conditions. Table 4 compares the calculated outlet air

The IEHX is used as an evaporative enhanced heat recovery device to pre-cool the ambient air. The developed numerical model is first adopted to study the air conditioning process in the IEHX. Fig. 10 shows the variation of air stream conditions, namely, the DBT and humidity ratio profiles. The plate temperature represents the interface DBT at the surface of the channel wall. In addition, the plate temperature can be considered as the water film temperature at the channel surface. In this study case, a representative outdoor condition (T = 33 °C, RH = 80%) is selected as the inlet condition for the product air (PA), while the inlet condition for the working air (WA) is kept at room condition (T = 24 °C, RH = 60%). As illustrated in Fig. 10, the PA temperature reduces gradually as the air flows along the passages. The product channel involves two regions according the dry/wet conditions. In the first region, the DBT of the product channel interface is higher than the DPT of the PA. As a result, the product channel surface in this region is dry indicating that the PA is not condensed. In the second region, it is possible that the PA condenses since the DBT of the product channel interface is lower than the DPT. In addition, for the product channel interface, the reduced temperature results in a decrease of the saturated humidity ratio. For example, in Fig. 10, the product channel plate temperature decreases to its DPT at the location of

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Row number

ma (kg/s)

mw (kg/s)

Ta,in ( °C)

RHa,in (%)

Calculated Ta,out ( °C)

Experiment Ta,out ( °C)

Discrepancy

1 2 3 4 5 6 7 8

4 4 4 4 8 8 8 8

1.20 1.20 1.16 0.72 1.27 1.04 1.13 1.27

0.46 0.46 0.45 0.45 0.44 0.52 0.43 0.24

25.99 25.99 28.68 28.68 24.59 27.29 23.94 24.59

70.22 54.13 55.60 55.60 55.02 51.19 70.80 55.02

16.91 15.38 17.58 14.34 14.02 14.23 14.46 16.43

17.17 15.27 17.44 14.69 12.84 13.47 13.24 15.83

−1.51% 0.72% 0.80% −2.38% 9.19% 5.64% 9.21% 3.79%

Table 5 The specification of cooling coil. Parameter

Value

Unit

Number of rows Transverse pitch Longitudinal pitch Coil Length Outer diameter of tube Inner diameter of tube

6 0.038 0.033 1.2 0.0127 0.0119

– m m m m m

The water condensation from the PA requires latent heat transfer from the PA to the WA. The latent cooling load for PA with a higher inlet RH accounts for a larger proportion of the total cooling load. As a result, for a specific inlet temperature of PA, the outlet temperature increases with the increase of its RH as shown in Fig. 11(a). In addition, the calculated outlet humidity ratio demonstrates the ability of the IEHX to pre-dehumidify the PA and fulfil part of the latent cooling duty of the ambient intake air. For example, in Fig. 11(b), the humidity ratio of PA reduces from 36.8 g/kg to 24.2 g/kg for a specific inlet condition T = 37 °C and RH = 90%.

5.4. Impact of pre-cooling on the performance of air handling unit

Fig. 11. Pre-cooling performance of the IEHX. (a) Outlet temperature of PA; (b) Outlet humidity ratio of PA.

x/L = 0.25. Thereafter, the humidity ratio of the PA decreases along the product channel so that the PA condenses in this region. In the wet working channel, the WA is more capable of absorbing water at the entrance of the working channel, resulting in a sudden temperature drop. Thereafter, since the heat is steadily transferred from the PA to the WA, the temperature of WA first decreases slightly and then increases along the flowing passages. Fig. 11 shows the outlet conditions of PA under varying ambient intake air conditions. In the simulation, six varying outdoor air temperatures (27 °C, 29 °C, 31 °C, 33 °C, 35 °C, and 37 °C,) and three different relative humidity (70%, 80%, and 90%) levels have been considered. It is apparent that the IEHX is able to effectively precool the ambient air by taking the advantages of the evaporative cooling.

After being sensibly cooled by the IEHX, the pre-cooled air is still unable to satisfy the sensible or latent comfort needs in humid tropical climates. Therefore, the intake air needs to be further conditioned by the cooling coil to achieve a desired supply air condition. Fig. 12 presents the influence of the pre-cooling IEHX on the performance of the cooling coil. In the simulation process, the ambient air is taken at a representative tropical climate condition with DBT of 33 °C and RH of 80%. Fig. 12(a) shows the outlet temperature obtained by employing the cooling coil without the usage of pre-cooling IEHX. The outlet air temperature is calculated for different outdoor air intake fraction and Tw,in . The specification of the cooling coil is illustrated in Table 5. The humid outdoor air requires a significant cooling capacity. Therefore, the outlet air DBT appreciates for a higher outdoor air intake fraction. In addition, the Tw,in markedly influences the cooling performance. It is observed that the outlet temperature of the PA increases about 0.7 °C with a rise in the Tw,in of 1 °C. To compare the outlet temperature of PA with the assistance of IEHX, Fig. 12(b) indicates the cooling performance by using the proposed hybrid evaporative pre-cooled air-conditioning system. In the hybrid system, the IEHX is able to provide the pre-conditioned air for the cooling coil. In this example illustration, the ambient air is first treated by the IEHX to a product air condition with 25.4 °C DBT and 98% RH. It can be inferred from Fig. 12(b) that the precooling process is able to raise the required Tw,in especially under a higher fraction of the outdoor air intake. For example, to obtain a supply air with an outlet DBT around 15 °C, the Tw,in can be raised from 7 °C to 10 °C under an outdoor air fraction of 50%. Since the COP of the chiller is generally enhanced by around 3.7% for 1 °C

Please cite this article as: X. Cui, M.R. Islam and K.J. Chua, An experimental and analytical study of a hybrid air-conditioning system in buildings residing in tropics, Energy & Buildings, https://doi.org/10.1016/j.enbuild.2019.06.028

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Fig. 12. Impact of pre-cooling unit on the performance of cooling coil. (a) Without pre-cooling IEHX; (b) With pre-cooling IEHX.

increase of the Tw,in [42], accordingly, the energy consumption can be reduced with a rising Tw,in . 5.5. Psychrometric illustration of the hybrid air cooling process Fig. 13 illustrates an example of the calculated air treatment conditions for a conventional cooling coil and a hybrid evaporative pre-cooled air-conditioning system. On the psychrometric chart, the point O represents the selected outdoor air condition (T = 33 °C, RH = 80%), and the point R shows the assumed indoor air condition (T = 24 °C, RH = 60%). In this example, the outdoor air fraction is designated to be 50%. In Fig. 13(a), the ambient air at point O is first mixed with the room return air (R). Thereafter, the mixed air at point M is further conditioned through the cooling coil. The Tw,in of 7 °C is able to obtain a supply air with a state at point S (T = 14.6 °C, ω = 9.8 g/kg). In Fig. 13(b), the outdoor air (O) and the exhaust air (R) are first treated in the IEHX. Due to the combined effect of the heat recovery and the evaporative cooling process, the PA in the IEHX is cooled and dehumidified as the condition is varied from point O to point P. The exhaust air absorbs heat and moist in the wet channel resulting in a final condition at point W. The PA and the return air are then mixed to a condition at point M. It is observed that the enthalpy at point M’ in Fig. 13(b) is lower than the enthalpy at the mixed point M in Fig. 13(a). Therefore, the chiller’s cooling capacity

9

Fig. 13. Description of air treatment conditions on psychrometric chart for (a) conventional cooling coil, and (b) hybrid evaporative pre-cooled air-conditioning system.

is reduced due to the pre-cooling process in the IEHX. In addition, the hybrid system has the capability to increase the Tw,in to 10 °C for achieving a similar supply air condition at point S in Fig. 13(b). 5.6. Energy saving potential by employing indirect evaporative pre-cooling To evaluate the energy saving contribution of the pre-cooling IEHX under tropical climates, simulations are also conducted for a hypothetical office building equipped with the proposed hybrid system. The occupancy of 0.3 person/m2 is maintained as the design value for the office room. The internal load is set as a design value of 18.3 W/m2 for indoor lighting and equipment. The total area for the conditioned space is 30 0 0 m2 . In tropical climates, the outdoor air condition consistently presents a high temperature and a high relative humidity throughout the year. The intake of outdoor humid outdoor air has a significant impact on the ventilation load. An hour-by-hour evaluation of the proposed system has been carried out under the weather condition on a design day of Singapore that is a representative city featuring tropical rainforest climates. Fig. 14 shows the hourly ambient temperature variations on a design day. It is readily apparent that the hourly weather condition varies marginally with a highest DBT of 33.2 °C at 2 pm on

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6. Conclusions

Fig. 14. Temperature variation for the ambient air and the pre-cooled product air on design day.

The performance and potential of a hybrid evaporative precooled air-conditioning system has been studied based on a systematic experimental and analytical approach. Respective mathematical models for both IEHX and cooling coil have been judiciously developed. The validation of the numerical model is performed by comparing the predicted values against experimental data in terms of the temperature variations in IEHX, the moisture mass fraction difference during moist condensation process, and the outlet temperature of the cooling coil. Simulation studies are conducted to investigate the air conditioning process in the IEHX and the impact of the evaporative-enhanced heat recovery process on the performance of the AHU under humid tropical climatic conditions. The evaporative-enhanced heat recovery IEHX is an energy-efficient device with an achievable COP up to 14.2. The chiller is able to handle a lower cooling load since the IEHX has shown its capability to fulfil about 32% of the total cooling rate. Accordingly, the chilled water for the chiller can be supplied with a higher temperature resulting in an improvement on chiller’s COP. Acknowledgements This work was supported by the Fundamental Research Funds for the Central Universities (xjj2018074), and China Postdoctoral Science Foundation (2018M631153). Author declaration We wish to confirm that there are no known conflicts of interest associated with this publication and there has been no significant financial support for this work that could have influenced its outcome. References

Fig. 15. Variation of cooling rate for the IEC and the cooling coil on design day.

the design day. In addition, Fig. 14 also illustrates the air temperature after IEHX pre-cooling process. As shown in Fig. 14, the IEHX is capable of reducing the ambient air DBT with a highest temperature change of 10 °C on the design day. The enthalpy of the intake air is reduced in an energy efficient approach through the IEHX. As a result, the pre-cooling IEHX shows the ability to undertake a part of the cooling load. The desired ventilation rate for a typical office building is recommended to supply 10 L/s per person of ambient air [51]. Based on the pre-cooling performance of the IEHX, the respective cooling rate for the IEHX and the cooling coil have been calculated. Fig. 15 indicates the hourly cooling rate per air-conditioned floor area on the design day. The calculated result shows that the pre-cooling process in the IEHX can fulfill around 32% of the total cooling rate. The energy saving potential of the hybrid system can be achieved due to following three reasons. Firstly, the energy-efficient IEHX has been reported to reach an achievable COP up to 14.2 [52]. Secondly, the cooling capacity of the chiller is reduced due to the fact that a part of the ventilation load is handled by the pre-cooling IEHX. Thirdly, the pre-cooling process allows a higher Tw,in resulting in an improvement on chiller’s COP. Therefore, it is estimated that the proposed hybrid system obtains an energy saving up to 26% when it is employed in tropical climates.

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