Experimental performance evaluation of desiccant coated heat exchangers from a combined first and second law of thermodynamics perspective

Energy Conversion and Management 207 (2020) 112518

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Experimental performance evaluation of desiccant coated heat exchangers from a combined first and second law of thermodynamics perspective

T

P. Vivekh, D.T. Bui, M.R. Islam, K. Zaw, K.J. Chua

⁎

Department of Mechanical Engineering, National University of Singapore, 9 Engineering Drive 1, Singapore 117576, Singapore

ARTICLE INFO

ABSTRACT

Keywords: Desiccant coated heat exchangers Experiments Second law of thermodynamics Energy analysis Exergy analysis Entropy generation

The electricity-driven vapor-compression air-conditioning system registers around 3–10% second law efficiency under tropical climates due to the coupling between sensible and latent cooling loads. Desiccant coated heat exchangers (DCHEs) are the next-generation technologies that can improve the efficiency of the air-conditioning process by decoupling sensible and latent loads. Existing literature on DCHEs is primarily directed towards material development and performance evaluation from the first law of thermodynamics standpoint. However, to have an insight into the operation of DCHEs concerning its process irreversibility, performing the second law of thermodynamics analysis is necessary. This analysis will pinpoint the causes of irreversibility in the DCHE system and evaluate the useful work destroyed during its cyclic operation. In this paper, we have developed a general steady-state thermodynamic framework to study the performance of DCHEs from a combined first and second law perspective. Experiments were carried out to obtain the thermodynamic state properties of air, water, and desiccant that are necessary to perform energy and exergy analyses. Further, fundamental parametric studies were conducted to understand the influence of different desiccant type, varying operating parameters, and ambient conditions. Key results revealed that the regeneration process contributed to significant entropy generation rates and by reducing the hot water temperature by 10 °C, the second law efficiency illustrates improvement of almost two times. Additionally, by selecting to coat the heat exchanger with a composite polymer material instead of silica gel, the second law efficiency of DCHEs is enhanced by around 2.6 times. Lastly, a hybrid air-conditioning system comprising vapor-compression chiller and DCHE records 50% savings in energy consumption and yields two-times higher second law efficiency.

1. Introduction Mechanical vapor compression (MVC) air-conditioners are the most commonly employed cooling technology to meet rising building energy demands [1]. Despite its advantages on cooling effectiveness and capacity, the air-conditioning system is energy-intensive and demonstrates around 3–10% second law efficiency operating in tropical climatic conditions [2]. Since the cooling energy demand is expected to rise at a much faster rate due to rapid economic growth and global warming [3], finding energy-efficient and clean alternatives to conventional air-conditioners is necessary. Desiccant coated heat exchangers (DCHEs) are one of the possible substitutes that promote the air-conditioning process efficiency by decoupling the sensible and latent cooling loads. DCHEs do not rely on mechanical compressors and are powered by solar/geothermal energy or by industrial waste heat. Due to their high potential in energy savings, DCHEs have been extensively studied in several applications such as air-conditioners [4,5],

⁎

heat pumps [6–8], adsorption chillers [9,10], and atmospheric water harvesters [11,12]. A comprehensive review on the recent developments in DCHEs is found in one of our previous works [13] and they are broadly classified into three categories, as illustrated in Fig. 1. In the material development and characterization purview, multiple studies highlighted the use of pure ceramic desiccants such as silica gel [14,15], zeolite [16], molecular sieves [17,18], and metal-organic frameworks [19,20]. Inherent limitations in the pore characteristics of these ceramic materials affected their adsorption capacity [21] and problems in bulk production of MOFs restricted their use in DCHEs [22]. Therefore, to improve the dehumidification performance, composite ceramic desiccants were developed by combining silica gel with lithium chloride [23,24], calcium chloride [25], potassium formate [26], and sodium acetate [27]. Further, composite polymer desiccants were developed by mixing equal concentrations of superabsorbent polymers and lithium chloride. This combination achieved 6–12 times improvement in sorption capacity

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

https://doi.org/10.1016/j.enconman.2020.112518 Received 21 October 2019; Received in revised form 14 January 2020; Accepted 17 January 2020 0196-8904/ © 2020 Elsevier Ltd. All rights reserved.

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Nomenclature cp Exdest f h ma mw n P Pel, Qa Qlost Qw R s sg Sgen T tcyc Wactual Wrev x y

ω ~

specific heat capacity at constant pressure, kJkg-1K−1 exergy destruction rate, kW function of derived parameters with respect to directly measured parameters specific enthalpy, kJ/kg air flow rate, kg/s or kg/h water flow rate, L/min total number of directly measured parameters pressure, kPa electrical power, kW cooling capacity of air, kW heat loss rate, kW total heat energy exchange of hot and cold water, kW gas constant, kJkg-1K−1 specific entropy, kJkg-1K−1 specific entropy generation, Jg-1K−1 overall entropy generation rate, W/K temperature, oC or K cycle time, min actual work input in the DCHE process, kW reversible work input for an ideal air-conditioning process, kW absolute uncertainty of the directly measured parameters absolute uncertainty of the derived parameters

Subscripts 0 a atm cw CT d f g hw in MVC out r rev sat v w

v

II

dead state air atmosphere cooling water cooling tower dehumidification saturated liquid state saturated vapor state hot water inlet mechanical vapor compression chiller outlet regeneration reversible saturated vapor liquid water

Abbreviations COPth DCHE MRC MRR MVC RH

Greek symbols Δ

humidity ratio, g/kg or kg/kg mole fraction ratio of vapor to air, mol/kmol relative humidity

difference specific flow exergy, kJ/kg second law efficiency specific volume, m3/kg

thermal coefficient of performance desiccant coated heat exchangers moisture removal capacity, g/kg moisture removal rate, g/h mechanical vapor compression relative humidity, %

Fig. 1. Classification of the recent advances in the field of desiccant coated heat exchanger technology.

when compared to silica gel based desiccants [28,29]. Dynamic performance evaluation studies were conducted by coating the developed desiccants on fin-tube heat exchangers, and their respective moisture removal capacity and energy efficiency were computed. Parametric studies were carried out to evaluate the DCHE’s performance under the controlled variation of (1) outdoor weather conditions such as air temperature [30,31] and air humidity content [32,33]; (2) operating parameters such as air flow rate [23], cycle time [15], and water temperature [35]; and (3) geometric configurations such as variations in fin-depth [34] and the use of microchannel heat exchangers [35]. Since obtaining experimental findings were timeconsuming and expensive, Ge et al. [36] first developed a lumped parameter model incorporating pre-determined heat and mass transfer correlations and predicted the transient behavior of DCHEs. Higashi et al. [37] utilized the model to calculate the temperature dependence of mass diffusivity on polymer coated heat exchangers. Li et al. [38]

adopted the heat and mass transfer resistance network principle and established new correlations for determining transfer coefficients. Kim et al. [39] developed a 2D analytical approach of an adsorption chiller comprising polymer coated heat exchanger and evaluated the heat and mass fluxes as explicit functions of dimensionless numbers. Vivekh et al. [40] developed a 3D numerical approach and predicted the flow field temperature and humidity distributions in different domains of DCHEs. The research on material synthesis and performance evaluations were conducted from the first law of thermodynamics standpoint. The first law analysis primarily provides information on the required regeneration energy as compared to the system cooling load. However, it fails to account the quality of the energy required, and no information on the irreversibility in the system can be obtained. A second law based analysis is therefore essential to develop a more in-depth understanding on the causes of irreversibility and quantify the amount of entropy generated by the DCHE system vis-à-vis an ideal air-conditioner. While 2

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combined first and second law of thermodynamics or entropy generation analyses have been carried out extensively for adsorption chillers, fixed-bed dehumidifiers, and desiccant wheel based systems, DCHEs have not been evaluated from the second law of thermodynamics outlook. A summary of the second law based studies employed for adsorption chillers, fixed-bed dehumidifiers, and desiccant wheels is presented in Table 1. Chua et al. [41] conducted an entropy generation analysis for twobed adsorption chillers. They assessed that the most substantial portion of entropy generation in a desiccant bed resulted from the adsorption/ desorption heat transfer and occurred during the switching phase. Li et al. [42] conducted an entropy generation analysis and studied the effect of cycle time and hot water temperature for an adsorption chiller with zeolite-water pair. The desorption bed recorded the highest entropy production rate, and a hot water temperature of 65 °C was necessary to minimize the entropy generation. Myat et al. [43] formulated an entropy generation minimization approach for multi-bed silica gel packed dehumidifiers. The system recorded the least specific entropy generation rate when the regeneration source temperature was at 73 °C. Enteria et al. [44] performed a first and second law of thermodynamic analysis on a solar energy powered desiccant wheel system and found that the solar collector destroyed around 55% exergy. In contrast, Jani et al. [45] identified that the desiccant wheel dehumidifier contributed to the highest percentage of exergy destruction in a hybrid solid desiccant air-conditioning system. The overall exergetic efficiency of the hybrid air-conditioning system could be improved to 11.45%. Hurdogan et al. [46] conducted an exergoeconomic analysis on a desiccant wheel air-conditioner and observed that the overall irreversibility of the hybrid system was around 70%. Caliskan et al. [47] conducted energy, exergy, and sustainability analyses to assess the performance of a cooling system comprising desiccant wheel, sensible heat wheel, and evaporative cooler and determined that desiccant wheels generated maximum entropy when compared with other components. These studies evaluated the thermodynamic efficiency of an overall desiccant cooling system, and identified the sources of maximum entropy generation. However, Bulck et al. [48] argued that conducting a second law analysis of individual components of the system would provide additional insight on irreversibility in the component’s operation. Accordingly, they developed a methodology to evaluate the second law efficiency of desiccant wheels as a function of its operating parameters. It is noteworthy that DCHEs offer high heat transfer efficiency and enhanced dehumidification capacity when compared to fixed-bed and desiccant wheel dehumidifiers [13]. Therefore, the rate of entropy production must be identified, and the second law efficiency of a DCHE

system must be evaluated, both of which have not been quantified yet in existing studies found in the literature. Moreover, information on the thermodynamic performance of DCHEs under the controlled variation of operating parameters and ambient conditions is also not available. In order to bridge these research gaps, this study attempts to provide an in-depth understanding of DCHEs from the second law of thermodynamics perspective. A judiciously crafted energy and exergy balance analyses are carried out on the experimental results of different DCHEs. The heat losses involved in dehumidification and regeneration processes and the specific flow exergy of moist air and liquid water are determined. The actual work during DCHE processes is computed and compared against the theoretical least work. Key second law performance indicators are judiciously defined, and their behavior under the influence of ambient conditions and operating parameters are investigated. Lastly, the impact of introducing a DCHE in a hybrid airconditioning system is comprehensively studied by computing the energy savings potential and determining the overall improvement in the second law efficiency. 2. Experimental methodology 2.1. Testing facility Fig. 2 shows the schematic and the photograph of the testing facility developed to evaluate the dynamic performance of DCHEs. The testing facility comprises (a) a DCHE chamber, (b) an air blower to regulate the air flow rate, (c) a humidity and temperature chamber to achieve desired climatic conditions, (d) two water baths to produce the hot and cold water requirements at a chosen temperature, and finally (e) a water valve mechanism to switch between hot and cold water. Two RTD temperature sensors and a velocity meter were placed before and after the testing chamber to measure the dry bulb, wet bulb temperature, and the velocity of the air passing through the heat exchanger. Additionally, to measure the surface temperature of the DCHE, two surface RTD temperature sensors with ± 0.5% full scale accuracy were mounted on the surface of the metallic fins. A variable-area flowmeter with ± 4% full-scale accuracy was used to record the water flow rate. LabVIEW data acquisition system was used for logging and recording the data obtained from the sensors. The uncertainty in the experiments is mainly attributed to the accuracy of the sensors and repeatability of the experiments. A summary of the measuring instruments used in the setup, along with their respective accuracies, is listed in Table 2. While the uncertainty of directly measured parameters were determined by the accuracies of the sensors, the error propagation method [49], as

Table 1 A summary of second law based studies on different types of desiccant dehumidifiers. Study

System type

Key features

Chua et al. [41]

Fixed-bed adsorption chiller

Li et al. [42]

Enteria et al. [44] Jani et al. [45] Hurdogan et al. [46]

Desiccant coated adsorption chiller Multi-bed desiccant dehumidifier Desiccant wheels Desiccant wheels Desiccant wheels

Caliskan et al. [47]

Desiccant wheels

Bulck et al. [48]

Desiccant wheels

Current Study

Desiccant coated heat exchangers

Transient entropy generation analysis of a two-bed adsorption chiller with silica gel-water pair. The effect of cycle time on entropy generation rate is discussed. Entropy generation analysis was conducted for zeolite-water paired adsorption chiller. The effect of variation of hot water temperature and cycle time was studied. Experimental and theoretical study on transient entropy generation was carried out for multi-bed silica gel packed desiccant dehumidifiers. First and second law based study for a solar assisted desiccant air-conditioning system. Energy and exergy analysis of a hybrid air-conditioning system was carried out. Exergoeconomic analysis (EXCEM) of a desiccant wheel based air-conditioning system was carried out. A method to evaluate the cost associated with exergy was developed and the economic value of the thermodynamic losses were identified. Energy, exergy, and sustainability analyses were carried out for a desiccant cooling system comprising desiccant wheel, sensible wheel, and regenerative evaporative cooler. The second law of thermodynamics was applied specifically to desiccant wheels with infinite and limited transfer coefficients. The effect of operating parameters on second law was discussed. Performance of a DCHE system was analyzed from a combined first and second law of thermodynamics framework. A steady-state exergy-balance approach was used to evaluate the entropy production rates during dehumidification and regeneration processes. Parametric studies were carried out to understand the effects of desiccant types and variation of operating parameters and ambient conditions.

Myat et al. [43]

3

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Fig. 2. (a) A schematic of the DCHE dynamic performance testing facility and (b) a photograph of the testing facility.

Table 2 Summary of specifications and accuracy of temperature and velocity sensors used in the testing facility. Parameter

Sensor/Instrument

Range

Accuracy

Temperature Surface temperature Air flow rate Water flow rate

RTD temperature sensor RTD temperature sensor Air velocity meter Variable area flowmeter

−29–100 °C −200–260 °C 0–10 m/s 1–7 L/min

1/10 DIN SRTD-1 ± 0.5% ± 2.5% (Full Scale) ± 4% (Full Scale)

4

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expressed in Eq. (1) was employed to compute the uncertainties of the derived parameters. n

y= i=1

f xi

an exergy balance approach to its reverse process i.e., reversible isothermal mixing of air and water vapor [51]. The exergy destroyed in the reversible isothermal mixing process is regarded as the minimum work required for reversible isothermal dehumidification process, as shown in Eq. (2).

2

( x i )2

(1)

where y is the absolute uncertainty of the derived parameters; x is the absolute uncertainty of the directly measured parameters; n is the total number of directly measured parameters; and f is the function of the independent variables.

Wrev,1

2

= ma

a,2

1

+

2

w,2

1000

(2)

a,1

where ma is the mass flow rate of dry air; is the humidity ratio of air; a is the specific flow exergy of moist air; and w is the specific flow exergy of liquid water. It is noteworthy that in this study, the moist air is assumed to behave as an ideal mixture of dry air and water vapor. The thermodynamic properties of liquid water are taken as those of saturated liquid at the desired temperature. Accordingly, the specific flow exergy of moist air and water is evaluated by employing Eqs. (3) and (4), respectively [52]. The thermo-mechanical components of specific flow exergy measure the magnitude with which the temperature and pressure of the system differ from the dead state. On the other hand, the chemical exergy signifies the change in the chemical composition of the system with respect to its dead state.

2.2. Operation description The humidity chamber and the air blower were precisely controlled to achieve the required inlet conditions. The air passed through the bypass line until steady-state conditions were reached. Subsequently, the air valves were adjusted such that the air interacted with the DCHE. The coated desiccant absorbed the moisture from the air, and the cooling water flowing through the tubes captured the sorption heat. When the specified cycle time for dehumidification reached its threshold, the cooling water flow was switched to the hot water flow. The absorbed water molecules were purged out of the DCHE as the hot water internally heated the coated desiccant, and the air carried away the released moisture. Since a single DCHE based chamber was employed in the testing facility, the dehumidification and regeneration cycles were switched alternatively.

a

= (cp, a + cp, v ) T0

T T0

1

ln

T T0

P + (1 + ~) Ra T0 ln P0 Mechanical component

Thermal component

1 + ~0 + Ra T0 ln 1+ ~

2.3. Variation of the operating conditions

~ 1+ ~ 0 + ~ ln ~ ~ 0 1 +

(3)

Chemical component

Table 3 specifies the list of operating parameters employed in this study. During both dehumidification and regeneration processes, the inlet air temperature and humidity ratio were maintained at the baseline conditions of 30 °C and 21.5 g/kg. The range of air flow rate is selected based on the commonly maintained desiccant-air contact time of 0.1–0.12 s. The water flow rate was maintained relatively stable between 2.5 and 2.7 L/min throughout the experimentation, and a constant cycle time (tcyc ) of 10 min was used for all the studies. In addition, experiments were carried out to understand the influence of different climatic conditions, which will be discussed in detail in section 3.3.

w

= h f (T )

hg (T0)

T0 (sf (T )

sg (T0 )) + (P

Thermal component

Psat (T )) vf (T )

Mechanical component

Rv T0 ln 0 Chemical component

(4)

where cp, a and cp, v are the respective specific heat capacities of dry air and water vapor at constant pressure; T is the absolute temperature; T0 is the dead state temperature; ~ is the mole fraction ratio of vapor to air with ~ = 1.608 ; Ra is the gas constant of dry air; hf (T ) and sf (T ) are the respective specific enthalpy and entropy values of water in the saturated liquid state evaluated at T; hg (T0 ) and sg (T0) are the respective specific enthalpy and entropy values of water in the saturated vapor state evaluated at T0; P is the pressure of air (taken as Patm = 101.325 kPa in this study); Psat is the saturation pressure of water evaluated at T; Rv is the gas constant for water vapor; and 0 is the ambient relative humidity ratio. The mathematical correlations for the thermodynamic state properties of water were directly adopted from Saul and Wagner [53]. In the second stage of the ideal air-conditioning process, the dehumidified air is sensibly cooled without any change in its chemical composition. The minimum work required for this process is evolved by considering a reversible Carnot refrigerator exchanging heat with the environment [51] and is represented as

3. Evaluation methodology In this section, the least work for an ideal air-conditioner is first determined, and a methodology is then developed to evaluate the entropy generated in the actual DCHE system. For any thermodynamic system, the heat energy input does not entirely convert to useful work. A parameter is therefore necessary to evaluate the quality of the energy. As a result, exergy analysis approach is employed since it measures the available/useful energy of a thermodynamic system with reference to a zero/dead state. The selection of the dead state impacts the magnitude of the exergy, and typically for the HVAC applications, the ambient state is accepted as the dead state [50]. 3.1. Evaluation of thermodynamic least work for air-conditioning process

Table 3 Operating conditions for performance testing of DCHEs.

Thermodynamics imposes a process-independent lower bound on the energy consumption of the air-conditioning process. This is the minimum energy input required for the idealized operation of an airconditioning system operating between the specified thermodynamic states. Fig. 3 shows the schematic of an ideal air-conditioner designed for evaluating the least work. A psychrometric chart highlighting the two stages of the corresponding reversible thermodynamic paths is also presented. In the first stage, the air is dehumidified in a quasi-steady reversible isothermal process. Its minimum work input is evaluated by employing

Parameters

Units

Baseline conditions

Change range

Cooling water temperature (Tcw,in) Hot water temperature (Thw,in) Air flow rate (ma ) Inlet air temperature (Ta,in) Inlet air humidity ratio ( a, in)

o

C C kg/h o C g/kg

25 50 27.5 30 21.5

15–30 40–70 17.5–32.5 30–36 –

Water flow rate (mw )

L/min

2.5–2.7

–

Cycle time (tcyc )

5

o

min

10

–

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Fig. 3. Schematic of (a) an ideal air-conditioner for calculating the thermodynamic least work; and (b) psychrometric chart highlighting the two stages of the thermodynamic paths involved in the ideal air-conditioner. T2

Wrev,2

2

=

T0 T

1 T2

Q = ma (cp, a +

2

1000

cp, v ) T0 ln

T2 T2

(T2

represent the respective mass flow rates of cooling water and hot water; cp, w is the specific heat capacity at constant pressure for liquid water; Tcw,in/Thw,in and Tcw,out/Thw,out are the respective cooling/hot water temperature at DCHE inlet and outlet; and Qlost , d and Qlost , r are the respective heat loss rates of DCHE during dehumidification and regeneration processes. The specific enthalpy of moist air (ha ) is defined as the total enthalpy of dry air and water vapor mixture per every kg of dry air [54] and is computed via

T2 ) (5)

Therefore, by adding Eqs. (2) and (5), the least work input for the air-conditioning process is determined as

Wrev,1

2

= Wrev,1

2

+ Wrev,2

(6)

2

3.2. Evaluation of the actual work in DCHEs

ha = cp, a T +

Fig. 4 shows the control volume of a DCHE. The first law of thermodynamics for a control volume in a steady-state is applied to obtained the energy balance of DCHE during dehumidification and regeneration processes, and is respectively shown in Eqs. (7) and (8).

ma, d (ha, d, in

ha, d, out ) + mcw cp, w (Tcw, in

Tcw, out )

Qlost , d = 0

(7)

ma, r (ha, r , in

ha, r , out ) + mhw cp, w (Thw, in

Thw, out )

Qlost , r = 0

(8)

(cp, v T + 2501)

(9)

where T is the air temperature and is the specific humidity of moisture in air. The general exergy balance associated with the operation of DCHE under dehumidification and regeneration processes is carried out via Eqs. (10) and (11), respectively. Exergy balance analysis combines both the first and second laws of thermodynamics to assess the quality of the corresponding energy transfer. Lior [55] concluded that the average exergy of the humid air stream dominates over the total exergy in the desiccant. Therefore, the contribution from the desiccant towards the exergy is taken as negligible.

where ma, d and ma, r are the respective mass flow rates of air during dehumidification and regeneration processes; ha, d, in, ha, d, out , ha, r , in and ha, r , out are the specific enthalpies of moist air at the inlet and outlet of dehumidification and regeneration processes, respectively; mcw and mhw

Fig. 4. Schematic of (a) a DCHE system undergoing dehumidification and regeneration processes; and (b) psychrometric chart highlighting the corresponding thermodynamic paths. 6

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Exdest , d = ma, d (

a, d, in

ma, d (

Exdest , r = ma, r (

a, d, in

a, r , in

+ ma, r (

a, r , out

a, d, out )

+ mcw (

cw, in

a, d, out ) ad, water

a, r , out ) + mhw (

hw, in

a, r , in ) des, water

Table 4 Ambient climatic conditions used for the combined first and second law analysis of the DCHE based hybrid air-conditioning system.

cw, out )

(

Qlost , d 1

T0 Td

)

(10)

hw, out )

Qlost , r 1

T0 Tr

(11)

where Exdest , d and Exdest , r are the exergy destruction rates associated with the DCHE’s operation during dehumidification and regeneration modes; and T0 is the absolute dead state temperature. The first two terms on the right hand side of Eqs. (10) and (11) represent the exergy associated with the flow of moist air and cooling water/hot water during dehumidification and regeneration processes. The third term indicates the exergy storage in the system as the result of sorption/ desorption of moisture. Further, the last term in both the equations represents the exergy associated with the heat loss during dehumidification/regeneration process. The absolute rate of entropy generation (Sgen ) is obtained by combining the exergy destroyed during dehumidification and regeneration process and is defined as

Sgen =

Exdest , d + Exdest , r T0

Climatic condition

Ta,in (oC)

RH (%)

ωa,in (g/kg)

1 2 3 4

High T and high RH Moderate T and high RH High T and moderate RH Moderate T and moderate RH

~36 ~30 ~38 ~30

85% 80% 40% 40%

~32.7 ~21.6 ~18.7 ~11.3

°C °C °C °C

3.3. Evaluation of entropy generated in a hybrid air-conditioning system A schematic of the hybrid system comprising a DCHE and a MVC chiller is shown in Fig. 5. The first and second law of thermodynamics analysis is carried for the hybrid system and then compared with a standalone MVC air-conditioner. A space cooling application with 1 refrigeration ton load is considered for the study, and the system is aimed to operate with 100% outdoor air. To assess the hybrid system’s performance under various climatic settings, four specific outdoor air conditions were selected with a range of 25–40 °C in temperature and 50–90% in RH. These conditions are listed in Table 4. The performance of DCHE is obtained by conducting relevant experiments under varying climatic conditions and corresponding reduction in the total cooling is then computed. Following assumptions have been made to simplify the analysis:

(12)

Further, by employing the Gouy Stodola theorem [56], the relationship between the actual work input required for DCHE’s operation and its thermodynamic lower limit is established as

Wactual = Wrev + T0 Sgen

Cases

1. The indoor air temperature and moisture requirements were set at 24 °C and 50% RH in accordance with the ASHRAE thermal comfort guidelines [57].

(13)

Fig. 5. Schematic of (a) a conventional MVC chiller and (b) a hybrid DCHE+MVC air-conditioner operating in full-outdoor air configuration. 7

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2. The average COP of the conventional MVC chiller unit was assumed as 3 [58]. 3. Cooling water requirements of DCHE were supplied using a cooling tower. Its efficiency and power requirement were selected as 75% and 0.04 kW/RT, respectively [59]. On the other hand, hot water temperature requirement of 40 °C was supplied by condenser/industrial waste heat. 4. The fan power consumption is governed by several factors such as the pressure increase across the fan, volumetric air flow rate, and the efficiency of fan shaft and motor. From the detailed analysis carried out by Stephens et al. [60], the fan electrical power input ranges between 18 and 22% of the input power required for operating the standalone MVC system. As a result, a factor of 1.2 is taken to account for the 20% fan power requirement while evaluating the total power consumption.

moisture that is removed during the dehumidification process. To facilitate a proper comparison due to the irreversibility alone, it is necessary to normalize Sgen with MRR. Accordingly, a new parameter termed the specific entropy generation (sg ) is introduced in Eq. (19). It is defined as the ratio of the entropy generated per unit mass of moisture removed from the air.

sg =

Sgen ma (

(19)

a, out )

a, in

In addition, another key indicator, second law efficiency ( II ) is introduced to quantify the thermodynamic reversibility of DCHEs. It is defined in Eq. (20) as the ratio of the least work of an ideal air-conditioning system to the actual work input of the DCHE system, both operating between the same thermodynamic states.

4. Performance parameters

II

=

Wrev Wactual

(20)

4.1. First law based performance parameters of DCHE 4.3. Performance parameters of the hybrid air-conditioning system

The dehumidification performance of a DCHE is measured by computing its moisture removal capacity (MRC) which is defined as the average amount of moisture removed per kg of the dry air flowing over the DCHE for a given cycle time, as shown in Eq. (14).

MRC =

1 tcyc

The total electrical power required (Pel, MVC ) for the conventional MVC chiller is computed as

tcyc

(

a, in

Pel, MVC =

a, out ) dt

(14)

0

Pel, MVC =

(15)

The first law energy efficiency of DCHEs is calculated in terms of the thermal coefficient of performance (COPth), which is defined as the ratio of the average cooling capacity of air (Qa ) to the total heat exchange rate of cold and hot water during respective dehumidification and regeneration processes (Qw ) ,

COPth =

(16)

Tcw, in) + mhw cp, w (Thw, in

Thw, out )

(23)

(24)

Pel, hybrid = Pel, MVC + Pel, CT

Likewise, Q w is computed is computed using eq.

Qw = mcw cp, w (Tcw, out

ha,1 )

where ha,1 and ha,1 are the respective inlet and outlet specific enthalpies of moist air passing through DCHE. The total electrical power required (Pel, hybrid ) for the hybrid DCHE +MVC chiller is computed as

(17)

ha, out )

(22)

where ha,1 and ha,2 are the respective inlet and outlet specific enthalpies of moist air passing through the chiller in the hybrid system. The electrical power needed for the cooling tower (Pel, CT ) is computed by Eq. (23).

Qa is calculated using Eq. (17) where ha, in and ha, out represent the specific enthalpy of process air at inlet and outlet, respectively, during the dehumidification process. Qa = ma (ha, in

1.2ma (ha,1 ha,2 ) COPch

Pel, CT = 0.0114ma (ha,1

Qa Qw

(21)

where ma is the mass flow rate of air in the chiller; ha,1 and ha,2 are the respective inlet and outlet specific enthalpies of moist air passing through the chiller; and COPch is the coefficient of performance of the chiller. The factor 1.2 accounts for the contribution of fan power in the total electrical power consumption. The total electrical power required to operate the MVC chiller in the hybrid system (P el, MVC ) is obtained by Eq. (22).

where a, in and a, out are the respective air humidity ratio values at inlet and outlet of the DCHE chamber. For the specific case of variation in air flow rate, the amount of moisture absorbed may vary depending on the overall effect caused by the variation of ma and MRC. Therefore, the moisture removal rate (MRR) of the DCHE is then used to compare the dehumidification performance and is evaluated as

MRR = (MRC) ma

1.2ma (ha,1 ha,2 ) COPch

The electrical power saved (EPS) by the hybrid air-conditioning system is computed via

(18)

where cp, w is the specific heat capacity of water at constant pressure; Tcw, in , Tcw, out , Thw, in , and Thw, out are the respective inlet and outlet temperatures of the cooling water and hot water flowing inside the tubes of heat exchanger.

EPS =

Pel, MVC

Pel, hybrid

Pel, MVC

× 100

(25)

To evaluate the second law efficiency of the system, the thermodynamic least work is computed using Eq. (6). The second law efficiency of the hybrid air-conditioning system vis-à-vis conventional MVC chiller is then evaluated via

4.2. Second law based performance parameters of DCHE

Sgen , computed using Eq. (12), considers the overall entropy production rate during dehumidification and regeneration processes. However, it does not account for the moisture removal rate of the DCHEs. When the performance comparison is carried out between different desiccants operating under similar parameters, Sgen changes primarily due to the irreversibility in the system and the amount of

II , MVC

II , hybrid

8

=

=

Wrev,1 2 Pel, MVC

Wrev,1 2 Pel, hybrid

(26)

(27)

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5. Results and discussion

quick moisture sorption in the layer proximate to the desiccant-air interface. However, when this layer reaches saturation, the absorbed moisture needs to diffuse into the bulk of the desiccant layer to allow the sorption of new water vapor molecules. Since the desiccant’s moisture diffusivity (Dv, d) is of the order ~10-13–10-9 m2/s and is around 4–6 orders of magnitude lower than the diffusivity of water vapor in air (around 10 5 m2/s) [28], the moisture sorption is gradual. A schematic explaining this phenomenon is illustrated in Fig. 7. The variation of the air temperature is shown in Fig. 6 (b) and (c). Consistent with the initial rapid decline of a, out , outlet air temperature (Ta, in) and water outlet temperature (Tw, out ) swiftly rise by around 3 °C at the beginning of dehumidification/regeneration processes. This behavior occurs due to the fact that the moisture sorption process is exothermic and facilitates the release of the heat of sorption. Since the sorption rate is faster in the initial stage, both air and water temperature rise rapidly. Thereafter, the variation of air and water temperature is stable until the specified tcyc is attained. A similar analogy holds for the behavior of both air and water temperature during the regeneration process. Further, the transient function of the average desiccant surface temperature (Tsurface) is shown in Fig. 6(d). We observe that Tsurface remains largely stable at around 27.5 °C and varies marginally by less

5.1. Transient performance of DCHEs SAP-LiCl (50w%) coated heat exchanger is mounted on the testing facility, and dynamic experiments were carried out to study its performance. The inlet supply conditions of air, water, and other operating parameters were maintained stable at the baseline values as listed in Table 3. Fig. 6(a) illustrates the transient results of inlet and outlet air humidity ratio. At the start of the dehumidification process (t > 0) , when the desiccant begins absorbing moisture from air, the outlet air humidity ratio ( a, out ) first drops rapidly by about 30% from 21 g/kg to 14.9 g/kg in 50 s. It increases gradually by around 20% and tapers at tcyc = 600 s) . 17.8 g/kg towards the end of the specified cycle time (t Similarly, a, out swiftly rises by 50% to 31.8 g/kg in 40 s at the onset of regeneration process (t > 600 s) . Subsequently, it follows a steady drop until it arrives at a value closer to the inlet humidity ratio ( a, in ) . The sorption/desorption process remains largely gradual for over 90% of the cycle. The initial rapid decrease/increase of a, out is attributed to higher diffusivity of water vapor molecules in air. When the desiccant is dry, higher driving force for moisture transfer facilitates

Fig. 6. Dynamic behavior of inlet and outlet (a) air humidity ratio; (b) air temperature; (c) water temperature; and (d) average desiccant surface temperature. 9

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Fig. 7. A schematic representing moisture sorption and diffusion between air and desiccant layer.

than 2.5% during dehumidification process. This indicates that despite the variation in the release of heat of sorption, the desiccant surface temperature remains predominantly consistent. This observation reflects the fact that the heat of sorption is spontaneously distributed to air and water and does not contribute to increasing the thermal mass of the heat exchanger. Likewise, during the regeneration process, Tsurface stays around 43.9 °C and fluctuates by 4.5%. This observation supports the proposition that the energy stored in the system during the process is negligible, and a steady state assumption is indeed valid while analyzing DCHEs from the first and second law perspectives. It is worthy to note that the steady-state assumption is reasonable as we are concerned with the simultaneous heat and mass transfer process at a time scale of about 10 min or higher. When tcyc is reduced to 2 min or lower, the transient response of the system is critical as any minor fluctuations at the inlet will significantly affect the outlet conditions. A higher cycle time reduces the frequency of switching between the processes and also mitigates the heat losses occurring due to the frequent mixing between cold/hot water sources [61].

0.004 W demonstrating higher reversibility. In contrast, around 24.42 W of exergy input is needed during the regeneration process through the supplied hot water and around 17.06 W of exergy is destroyed. The irreversibility is associated with a higher temperature difference between hot water and dead state. Further, the limited heat transfer coefficient between hot water and air lowers the amount of useful work done during regeneration. As a result, the regeneration process generates higher exergy than the dehumidification process. It is insightful to compare the performance of the DCHE system with an ideal reversible air-conditioner operating between the same thermodynamic states. By employing Eq. (6), the theoretical least work input is computed as 3.49 W, which marks the second law efficiency of the DCHE system as 17%. The second law efficiency suggests that the scope for DCHE performance improvement in utilizing energy usefully exist, and some of the possible ways will be explored in the following sections.

5.2. Total exergy flow in DCHEs

This section investigates the thermodynamic performance of DCHEs based on the type of coated desiccants. Four DCHEs coated with different desiccants are considered for this study; a ceramic silica gel desiccant and three composite polymer desiccants: PVA-LiCl (50w%), SAP-HCO2K (50w%), and SAP-LiCl (50w%). To ensure that the performance variation is purely due to the desiccant type employed, identical heat exchangers were used, and the amount of the desiccantcoated was 50 g. The operating conditions were maintained stable at the baseline conditions, as listed in Table 3. Fig. 9 portrays the influence of the desiccant type on the first law performance indicators, i.e. MRC and COPth. Besides, the effect on air temperature is also depicted. The composite polymer based desiccants record around three times improvement in MRC compared to silica gel. Among the polymer desiccants, SAP-LiCl (50w%) records the highest MRC, which is marginally higher than PVA-LiCl (50w%) and SAPHCO2K (50w%). The improvement in dehumidification capacity is attributed to the enhanced water sorption capacity of the composite polymer desiccants. Higher sorption capacity of composite polymer is due to the excellent water retention ability of the polymer and greater absorption capacity of the hygroscopic salts. However, the type of desiccant coated has a marginal effect on the outlet air temperature, as the inlet and outlet air temperature differ by 1 to 2 °C. In terms of thermal efficiency, PVA-LiCl (50w%) and SAP-LiCl (50w%) record over two times improvement in COPth compared to silica gel coated heat exchangers. In contrast, SAP-HCO2K (50w%) shows around 50% enhancement. The improvement is solely attributed to the higher

5.3. Effect of desiccant type

The transient performance of DCHEs has been discussed in the previous section. In the ensuing section, the useful conversion and destruction of the input exergy in the dehumidification and regeneration processes of DCHEs are judiciously investigated. Fig. 8(a) shows the time-averaged cyclic inlet and outlet states of air temperature, humidity ratio, water temperature, and desiccant surface temperature. Based on these experimentally measured values of temperature and humidity ratio, the magnitude of the total flow exergy is evaluated. Fig. 8(b) portrays the graphical flow of exergy in the simultaneous heat and mass transfer occurring in DCHEs. Since the ambient conditions are employed as the dead state, the inlet air supply exergy is 0 W. The air is in thermal, mechanical, and chemical equilibrium with the dead state. In one of our previous studies [28], we have determined that the pressure drop in the DCHEs marginally fluctuates between 10 and 15 Pa (less than 0.001% compared to P0 = 101.325 kPa) for the selected range of airflow rate. Therefore, the mechanical component of the specific flow exergy of air is considered negligible compared to thermal and chemical components. During the dehumidification process, 3.694 W of net exergy is supplied through the cooling water and is used to dehumidify the air. During the specified cycle time, both air temperature and humidity ratio are reduced. Consequently, the exergy of the outlet air rises to 1.157 W while the adsorbed water reduces the exergy in the system by 2.33 W. The dehumidification process registers an exergy destruction of 10

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Fig. 8. (a) Time-averaged cyclic inlet and outlet states of air temperature, air humidity, water temperature, and desiccant surface temperature; (b) exergy flow diagram in the simultaneous heat and mass transfer process of DCHEs.

dehumidification capacity of the polymer desiccant. Fig. 10 (a) shows that the silica gel coated heat exchanger generates the maximum specific entropy (sg ) , which is 35% higher than the composite polymer desiccants. This observation is due to the greater exergy destruction during the regeneration process and poor MRC of silica gel coated heat exchangers. Among the polymer desiccants, only SAP-HCO2K (50w%) records increased sg . This is due to higher heat of sorption and marginally lower dehumidification capacity when compared with PVA-LiCl (50w%) and SAP-LiCl (50w%) coated heat exchangers. It also indicates that a higher work input is needed for SAPHCO2K (50w%) to reduce the temperature/humidity of the moist air. The results of DCHE’s reversibility vis-à-vis ideal air-conditioning system operating between the same thermodynamic limit is shown in Fig. 10(b). SAP-LiCl (50w%) records the highest second law efficiency of 35%; translating to 2.6 times enhancement in contrast to silica gel. A commercial vapor compression air-conditioner with a COP = 3 and

operating between the same thermodynamic states as that of SAP-LiCl (50w%) would record about 7% second law efficiency and would need approximately six times more input work than SAP-LiCl (50w%). 5.4. Effect of inlet air temperature The effect of inlet air temperature (Ta, in) on the thermodynamic performance of DCHEs is evaluated in this section. Fig. 11(a) illustrates that MRC is marginally lowered when Ta, in is regulated from 30 to 36 °C. The slight decline in dehumidification performance is attributed to the reduced sorption capacity of the desiccant at higher air temperature. A lower RH at higher air temperature decreases the sorption capacity of the desiccant and reduces the moisture transfer rate. However, a reduced MRC does not translate to a smaller COPth due to the higher temperature gradient between air and cooling water. Tcw, in is maintained consistently at 30 °C while the air temperature is regulated from 11

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Fig. 9. Influence of changing desiccant type on (a) MRC, (b) Air temperature, and (c) COPth, evaluated at the baseline conditions as listed in Table 3.

30 to 36 °C. A higher temperature gradient promotes heat transfer rates and yields improved COPth by about 40%. Consequently, additional sensible cooling is provided by the DCHE system at higher Ta, in , and the energy supplied for operating the DCHE is effectively utilized to increase the cooling capacity of the system. Correspondingly, Fig. 11(b) shows that the entropy generation rates are lowered, thereby improving the second law efficiency of the system by 25% at higher Ta, in .

5.5. Effect of air flow rate Fig. 12 (a) depicts the dehumidification performance of DCHE under varying air flow rate (ma) . Increasing ma from 17.5 to 32.5 kg/h reduces the air-desiccant interaction time by 47%, i.e. from 0.17 s to 0.09 s. As a result, MRC drops by 24% indicating that lower air flow rates produce drier air. However, the rate of moisture removed by the DCHE system

Fig. 10. Influence of changing desiccant type on (a) sg ; and (b) 12

II

evaluated at the baseline conditions as listed in Table 3.

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Fig. 11. Effect of variation of inlet air temperature (Ta, in ) from 30 °C to 36 °C on (a) MRC, COPth, and

II ;

and (b) Sgen and sg

Fig. 12. Effect of variation of air flow rate (ma) from 17.5 kg/h to 32.5 kg/h on (a) MRC and MRR; (b) COPth; (c) Sgen and sg ; and (d) Wrev , Wactual , and

increases by 43% at higher ma . Since the heat exchanger dimensions are identical, a greater magnitude of MRR points towards improved dehumidification capacity, i.e., higher absolute moisture removal rate per desiccant-air contact area. Fig. 12(b) shows that the COPth of the system depreciates almost linearly by 37%. The drop in the first law efficiency is attributed to the rise in total energy consumption at higher air flow rates.

II

Further, as observed in Fig. 12(c), absolute and specific exergy destruction rates measured in terms of Sgen and sg rise by about 150% and 80%, respectively. This is due to the directly proportional relationship between the exergy destruction rate and the air flow rate, as observed in Eqs. (10) and (11). Additionally, a higher heating energy requirement to regenerate the desiccant further contributes to higher entropy production 13

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rates. Fig. 12(d) shows that the actual work input required for operating the DCHE system increases by about 110%. Comparatively, for an ideal air-conditioning system, where the dehumidification and cooling processes occur reversibly, the rise in input work is considered to be marginal. This disproportionate increase in Wactual lowers the second law efficiency of the DCHE system by as much as 40%.

diminishes the dehumidification capacity of the DCHE. However, the heat losses to the environment appreciate at a higher Thw, in , and more heating energy is required for regeneration. Correspondingly, COPth is trimmed by 40%. From the first law analysis alone, the selection of an appropriate Thw, in requires a compromise between improved dehumidification performance and energy efficiency. From the second law perspective, a higher Thw, in raises the temperature difference with the dead state and a reduced dehumidification capacity also affects the chemical component of exergy. These two factors restrict the potential of DCHE in doing useful work. Therefore, Sgen and sg respectively increase by six times and 300%, and II drops by 70%. It is worthy to note that over 75% of II deterioration occurs when Thw, in is regulated between 40 and 50 °C. This observation highlights that the exergy losses are more sensitive to the temperature increase spanning 40–50 °C and maintaining the regeneration temperature to 40 °C improves the DCHE system’s potential to do useful work.

5.6. Effect of cooling water temperature The combined first and second law performance of DCHEs based on varying cooling water temperature (Tcw, in ) is shown in Fig. 13. At 15 °C, both MRC and COPth are observed to be the highest. When Tcw, in is systematically regulated to 30 °C, MRC and COPth are observed to linearly depreciate by 76% and 31%, respectively. A decreasing Tcw, in effectively contains the heat added during the sorption process and facilitates improved dehumidification performance and first law efficiency. On the other hand, the production of cooling water at temperatures significantly below the wet bulb temperature of air involves a higher operating cost. As a result, investigating the effect of Tcw, in on the DCHE’s second law performance is imperative to realize an optimum cooling water temperature. Fig. 13(b) shows that Sgen markedly depreciates over five times when Tcw, in is raised from 15 to 30 °C; indicating a better utilization of energy sources at higher Tcw, in . This is because a rising temperature difference between Tcw, in and T0 imposes greater destruction to the thermal component of exergy and results in heat losses. In a similar vein, sg is significantly diminished by about 25% when Tcw, in is raised from 15 °C to 25 °C. Thereafter, it marginally gains by 6% when Tcw, in is further regulated to 30 °C. Despite Sgen falling by 50% between 25 and 30 °C, the reason for a slight gain in sg at 25 °C is attributed to the improved dehumidification and cooling performance. Since lower Sgen promotes second law efficiency, II increases marginally by 7% between 15 and 25 °C and then decreases by the same magnitude when Tcw, in is further raised to 30 °C.

5.8. Second law performance of hybrid air-conditioning system In this section, the first and the second law performance analysis is carried out for a hybrid air-conditioning system comprising SAP-LiCl (50w%) coated heat exchanger and a conventional vapor compression chiller. The hybrid system’s performance was evaluated under four typical outdoor air conditions as listed in Table 4 and compared subsequently with the standalone MVC system. Fig. 15 (a) illustrates that the DCHE is capable of reducing the total cooling load imposed on the chiller and contributes to 40–55% reduction in the total electrical power consumed. While the DCHE primarily functions as a dehumidifier, 10–30% higher EPS is recorded under moderate humidity conditions. This observation conveys the point that under moderate humidity conditions, DCHE reduces both sensible and latent cooling loads considerably and lesser total cooling load is imposed on the MVC chiller. Accordingly, the second law efficiency of the overall hybrid system improves between 70 and 120%, as highlighted in Fig. 15 (b). Wrev,1 2 remains constant because both standalone and hybrid systems operate under similar outdoor and indoor air conditions. The improvement in II , hybrid is attributed to the reduction in input power requirement of the hybrid system involving DCHE. As far as standalone systems are concerned, the cooling coil temperature is maintained around 20–25 °C lower than the dead state/environment temperature. It is necessary to reach to such low-temperature levels because the chiller needs to remove moisture by dew point condensation. The employment of DCHE permits the total cooling load imposed on the chiller to be markedly

5.7. Effect of hot water temperature Fig. 14 illustrates the changes of the performance parameters under the influence of varying hot water temperature (Thw, in) . MRC increases by about 18% from 4.3 g/kg to 5.1 g/kg when Thw, in is regulated between 40 and 70 °C. This is because lowering Thw, in results in moisture being trapped in the desiccant that eventually leads to an inefficient regeneration. Consequently, the moisture contained in the desiccant

Fig. 13. Effect of variation of cooling water temperature (Tcw, in) from 15 °C to 25 °C on (a) MRC, COPth, and

14

II ;

and (b) Sgen and sg

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Fig. 14. Effect of variation of hot water temperature (Thw, in) from 40 °C to 70 °C on (a) MRC, COPth, and

trimmed. Accordingly, the temperature of the cooling coil can be sustained at closer proximity to the dead state temperature.

II ;

and (b) Sgen and sg

utilization of energy supplied in DCHEs. 2. The composite polymer DCHEs demonstrate up to four times improvement in dehumidification capacity and achieve 2–3 times higher first and second law efficiency vis-à-vis traditional silica gel coated heat exchangers. 3. A lower air flow rate lowers the entropy production rate and improves both first and second law efficiency by 3 times and 70%, respectively. 4. Tcw, in should be maintained around 5–7 °C lower than the ambient temperature to achieve improved dehumidification performance and lower entropy production rates. 5. Both first and second law efficiency significantly deteriorate with increasing supply hot water temperature. In contrast, selecting a low regeneration temperature may hinder the moisture sorption ability of the desiccants. Therefore, there exists an appropriate trade-off between thermodynamic efficiency and dehumidification capacity. 6. A hybrid air-conditioning system incorporating a DCHE reduces the electrical power consumption by about 50% and registers 75–120% higher overall second law efficiency.

6. Conclusions In this paper, several dynamic performance experiments are carried out on a DCHE system. Experimental results are analyzed using the first and second laws of thermodynamics. A steady-state evaluation methodology is employed to compute the energy transfer and heat losses associated with dehumidification and regeneration processes. The specific flow exergy of moist air and water flow are computed at each thermodynamic state, and the exergy balance equation is used to determine the entropy generated by the system. Further, DCHE is judiciously compared against a thermodynamically ideal and reversible airconditioning system operating between the same limits. The second law efficiency of the system is established accordingly. Key results and observations that emerged from this study include 1. The regeneration process contributes to the most substantial amount of entropy production, which points towards improvement in

Fig. 15. Comparing (a) electrical power required for standalone MVC and hybrid system and electrical power saved (EPS) by the hybrid system; and (b) overall second law efficiency of both standalone MVC and hybrid systems

15

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CRediT authorship contribution statement [18]

P. Vivekh: Conceptualization, Methodology, Investigation, Software, Validation, Visualization, Writing - original draft, Writingreview & editing. D.T. Bui: Validation, Formal analysis, Writing - review & editing. M.R. Islam: Supervision, Funding acquisition, Writing review & editing. K. Zaw: Supervision. K.J. Chua: Funding acquisition, Conceptualization, Formal analysis, Supervision, Writing - review & editing.

[19]

[20] [21]

Declaration of Competing Interest The authors declare that they have no known competing financial interests or personal relationships that could have appeared to influence the work reported in this paper.

[22]

[23]

Acknowledgement

[24]

The authors gratefully acknowledge the generous funding from the National Research Foundation (NRF) Singapore under Central Gap Fund Funding Scheme (R-265-000-641-281) managed on behalf by NUS Industry Liaison Office (ILO).

[25] [26]

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