Comparative Simulation of Solar Adsorption and Absorption Cooling Systems with Latent Heat Storage with Erythritol and MgCl₂·6H₂O

Abstract

The energy requirements for conditioning spaces have been increasing primarily due to population growth and climate change. This paper shows a comparison between an adsorption (ADC) and absorption cooling (ABC) systems to keep a building below the 25 °C set-point in dynamic conditions, utilizing a latent heat storage tank with MgCl₂·6H₂O and erythritol, and employing evacuated tube and parabolic trough collectors. The storage tank geometry is a plate heat exchanger. An auxiliary system was incorporated to control the temperature range of the solar cooling systems. The results showed that the coefficient of performance was kept around 0.40–0.60 and 0.70 for adsorption and absorption cooling, respectively. The latent heat storage tank with erythritol captured more solar energy than MgCl₂·6H₂O. A maximum solar fraction of 0.96 was obtained with MgCl₂·6H₂O, a thickness of 0.15 m, 20 m² of parabolic trough collector area, and absorption cooling, while the energy supply was fully satisfied with a solar collector with erythritol, a thickness of 0.1 m, 13 m² of parabolic trough area, and absorption cooling. In general, erythritol obtained better results of solar collector fractions than MCHH; however, it has less thermal stability than MgCl₂·6H₂O, and the cost is higher.

1. Introduction

The use of air conditioners and refrigeration accounts for approximately 15% of the world’s total electricity production [1]. This consumption exhibits an increasing trend due to demographic growth, as well as the impact of global warming. Solar thermal-driven refrigeration and cooling technology [2] is a promising renewable and sustainable technology for reducing greenhouse gas emissions.

Solar energy can be utilized to meet the heat requirements for sorption cooling systems in a sustainable manner, such as absorption cooling (ABC) and adsorption cooling (ADC) technologies. One of its limitations is its intermittent nature; therefore, the development of energy storage systems has become crucial [2]. A latent heat storage tank (LHST) has an advantage due to its high energy storage density and the ability to operate at a constant or near-constant temperature compared to storage tanks without phase change.

Studies related to LHST-coupled absorption systems in dynamic conditions are presented. Fan et al. [3] analyzed a shell-and-tube heat exchanger with hydroquinone as PCM in a solar-driven H₂O/LiBr double-effect absorption system. The enthalpy method was used to model the phase change. The results showed that natural convection is an essential parameter in the solidification process. Additionally, 100 kW of cooling capacity was fulfilled with a 12.55 m³ storage tank.

Pintaldi et al. [4] compared sensible (water and oil) and latent heat (KNO₃/NaNO₃ and AlSn) storage systems in a vertical shell-and-tube configuration, coupled to a triple-effect chiller and a parabolic trough collector (PT) at 200 °C. The results showed that the LHST obtained higher storage efficiencies than sensible heat storage systems. On the other hand, the sensible heat storage tank obtained better results with solar collectors than the LHST.

Zhou et al. [5] analyzed a single and a double hybrid effect absorption cooling device coupled with a linear Fresnel solar system. A typical shell-and-tube configuration was used as a storage tank to store molten salt. The COP varied from 0.73 to 1.09 for the single and double effect absorption systems. The optimized solar collector area (A_SC) and thermal storage capacity obtained values from 900 to 1100 m² and 5.0 to 8.5 m³, respectively.

Cerezo et al. [6] analyzed a solar absorption cooling system using a flat plate heat exchanger as LHST with MgCl₂·6H₂O (MCHH) in dynamic conditions with a PT, alongside water and synthetic organic fluid as the heating fluid using TRNSYS 17 software. Three configurations were evaluated: (a) sensible storage tank, (b) LHST, and (c) LHST with a tempering valve. The result showed that the sensible heat (configuration 1) did not fulfill the energy demand. Configurations (b) and (c) satisfy the demand using an LHST of 0.5 m³ with 30 m² and 20 m² of A_SC, respectively, and utilizing water as the thermal fluid. Water obtained better results than synthetic organic fluid, primarily due to its higher thermal conductivity.

Mehmood et al. [7] analyzed a solar absorption cooling to optimize the size of essential components, comparing conventional sensible heat and LHST in a configuration of a vertical cylinder with Rubitherm SP90 as PCM heated by evacuated tube collectors (ET) using TRNSYS software. The results showed that a 1.5 m²/kW_th latent heat storage tank, 30 L/m², and an insulation below 0.8 W/m² were selected to optimize the solar fraction (SF). The use of LHST increases the SF by 4% more than a conventional tank.

Migla et al. [8] presented an optimization of an absorption cooling system coupled to a vacuum solar collector and an LHST with different geometries (cylinders, spheres, and plates) using RT90HC as PCM, simulated in TRNSYS software. The results showed a 6.2% reduction in auxiliary energy.

Hosseini [9] designed and analyzed an ABC coupled to an LHST. The system consists of a cylindrical tank with multiple inner tubes through which the heating fluid (HF) passes using erythritol (ERY) as PCM, which is located between the inner tubes and the outer cylinder. The behavior of the system during charging and discharging was studied by varying the radius of the tubes, the mass of PCM, and the number of tubes. It was determined that, as the number of tubes and their radius increase, the efficiency of the LHST increases and the charging time decreases. The discharge process was studied by integrating the LHST with the maximum energy demand of the generator. It was concluded that the system provides more than 10 h of hot water at the necessary temperature to operate the cooling system.

Poshtiri and Jafari [10] analyzed an ADC in dynamic conditions to provide air conditioning during 24 h with stearic acid as PCM in a shell-and-tube heat exchanger, heated by flat solar collectors. The results showed that the air changes per hour and the fresh air ratio decreased the SF and increased auxiliary energy consumption. The reduction of the fresh air ratio from 100% to 20% increased the operation time of the chiller without PCM by approximately 2.5 h.

Studies on sorption systems coupled to an LHST are still very few, primarily for adsorption systems. The advantage of the ABC is its higher coefficient of performance (COP) compared to the ADC. On the other hand, the temperature of activation for adsorption is lower than that of absorption cooling, so a simple solar collector, such as a flat plate or ET, can be used. The use of PCM to store thermal energy is of great importance for the sorption cooling systems due to their high capacity to store heat in a small volume. This work shows an analysis of adsorption (ADC) and absorption cooling (ABC) systems heated by PT and ET using an LHST with MCHH and ERY as PCM to air-condition a building. Essential parameters of PCMs will be analyzed based mainly on PCM thermal properties and solar fraction. Both PCMs have similar melting temperatures; however, the latent heat of ERY is almost twice that of MCHH, while MCHH has a better thermal conductivity in the liquid phase than ERY.

2. Materials and Methods

2.1. Adsorption and Absorption Cycles

Absorption (ABC) and adsorption (ADC) cooling systems are based on the liquid and solid sorption processes, respectively. Sorption is a process that involves the absorption and desorption of a substance, known as a refrigerant, by a sorbent. The ABC cycle is composed of four components: generator, condenser, evaporator, and absorber, see Figure 1a. The thermodynamic cycle operates at two levels of pressure and usually three levels of temperature. The process serves as follows. Heat is supplied to the generator to vaporize the liquid refrigerant from the poor solution (low concentration). The refrigerant in the vapor phase is directed to the condenser device; in this part of the cycle, it releases heat (Q_CO). Then, the refrigerant in the liquid phase moves into a throttling valve to decrease its pressure and its temperature. This liquid is heated (Q_EV) in the evaporator until it reaches the vapor phase, and it is blended in the absorber device to create the rich solution (high concentration) from the generator. This physical/chemical sorption delivers heat (Q_AB), which is transferred to the environment. To maintain the steady-state cycle, the poor solution is pumped back to the generator. Typically, a single- or twice-cooling tower was used to absorb the heat from the absorber and condenser.

The ADC system is constructed from components such as ABC [11]; however, heat transfer can be achieved in two phases (Figure 1b). The first phase is the adsorption process: the refrigerant enters the evaporator, removing heat (Q_EV). After that, the refrigerant in the vapor phase enters the adsorber, where it is adsorbed into an adsorber bed and heat is dissipated (Q_AD). The second phase is the desorption process: the desorber (or generator) is heated (Q_DE) to evaporate the refrigerant and enters the condenser, dissipating a quantity of energy (Q_CO). It then moves through the throttling valve into the evaporator. To ensure the continuous operation of the adsorption chiller, two beds are used, in which the adsorption and desorption processes occur simultaneously as the chambers switch.

2.2. Description of the Solar Cooling System

Figure 2 illustrates a diagram of the solar absorption cycle, comprising three primary circuits: the solar collector (represented by two red lines), the storage tank (represented by two green lines), and the chilled water (represented by two blue lines). In the operation with a solar configuration, an HF is pumped (P₁) from the LHST into the solar collector and returned to the tank. In the storage tank configuration, the HF is pumped (P₂) from the LHST to transfer energy at the generator (GE) of the ABC. When the temperature of the LHST falls outside the operational range of the solar collector, a heating system or a heat dissipater is activated, controlling the generator temperature range (T_GE, as shown in

Table 1. Range of operation of the sorption cooling systems.

Component	TSET (°C)	TCHILLED (°C)	TCOOLING (°C)	TGE (°C)
ABC	6.7	5.5–10.0	26.6–32.2	108.9–116.1
ADC	6.7	5.0–12.0	10.0–35.0	65.0–95.0

). Finally, in the chilled water circuit, chilled water is pumped (P₃) from the evaporator (EV) of the ABC to the building and returns to it. The heat dissipated by the condenser (CO) and absorber (AB) is sent to a cooling tower.

The pump P₁ is activated by a differential controller (C₁). The controller opens or closes the fluid circuit when the output temperature is higher than the input temperature of the fluid in the storage tank or when the surface temperature of the PCM (placed at the outlet of the heating fluid) reaches 1 °C above the melting point (117 °C for MCHH and 119 °C for ERY, respectively). A thermostat (C₂) is used to control the building’s temperature at 25 °C, which in turn activates P₂, P₃, and P₄. A differential controller (C₃) regulates the flow rate entering the LHST. When the temperature of the surface PCM (placed at the same point as control C1) is one degree centigrade lower than its melting point, the flow rate is directed directly to the heating system; otherwise, it enters the LHST.

The components of the solar sorption system are simulated in TRNSYS software [12] using different types, described as follows.

Type 56 (Building) is a multi-zone building [13]. This component is simulated in a single zone with four windows. It is a simple house of 187 m³. Three occupants are in a resting position; more details can be found in [6]. The only difference in this work is that the building’s ceiling was isolated with a 0.10 m-thick layer of polyurethane to reduce energy losses.

Type 909 (Adsorption Chiller) and Type 107 (Single-Effect Absorption Chiller) utilize normalized catalog data taken to calculate the COP ratios as a function of the inlet temperature of chilled, cooling, and hot water [14]. The working fluids usually are water as a refrigerant and lithium bromide and silica gel as absorbent and adsorbent solutions. Table 1 presents the external input temperature range (T_CHILLED) of the evaporator and the condenser and absorber (T_COOLING) components of the sorption cooling systems, as well as the set point temperature (T_SET).

Table 2. Description of the type used in TRNSYS software.

Component	Type	Description
Parabolic trough collector	1288	This subroutine models a concentrating solar collector. The efficiency parameters for the PT are F′(τα) = 0.611, C1 = 1.42, C2 = 0.021, C3 = 0.0, C4 = 0.0, C5 = 6.653, C6 = 0.0 [16].
Evacuated tube collector	71	Intercept efficiency = 0.418; Negative first-order efficiency coefficient = 4.212 kJ/h m2 K [17].
Weather data	15-2	The weather data processor reads data at regular intervals from an external weather data file, interpolates it at time steps of less than one hour, and provides the processed data to other TRNSYS components. Meteorological data were used for Temixco, Mexico.
Heating system	6	This component increases the temperature of a flow stream using either internal or external control. Data provided: Overall heat coefficient = 2.58 kJ/h m2 K, efficiency = 0.98.
Heat dissipater	92	This component reduces the temperature of a flow stream using either internal or external control. Data provided: Overall heat coefficient = 2.58 kJ/h m2 K, efficiency = 0.98.
Cooling tower	510	“This type models a closed-circuit cooling tower. A closed-loop evaporative cooler that removes heat from a liquid stream by evaporating water over coils, with the working fluid fully isolated from air and water contact”.
C1 and C3	2b	This component is an on/off differential controller that switches off or on when it has a value of 0 or 1, depending on the upper and lower temperature differences, as well as the dead-band temperature differences.
C2	1503	This models three ON/OFF control functions to control a fluid cooling system.
P1, P2, P3, P4	3d	This component computes a mass flow rate using a variable control function. Data provided: Conversion coefficient = 0.05.

provides a brief description of the types used and the mathematical references for the solar cooling system components [12,15].

2.3. Mathematical Model of the Latent Heat Storage Tank

Type 66d was used to simulate the LHST programmed in Engineering Equation Solver [18] from a TRNSYS simulation. The storage tank is a rectangular heat exchanger consisting of three flat containers filled with PCM, with the heating fluid flowing through four channels to extract or supply energy [6]. The mathematical model was developed based on the following assumptions:

• Isothermal condition is considered in the phase change.
• The PCM thermophysical properties are independent of temperature.
• PCM is homogeneous and isotropic.
• The thermal resistance of the metal wall in the plates is not taken into account.
• Input and output transport properties of the HF are considered the same.

The mathematical model was solved using a finite difference scheme. In this work, each node was discretized for two dimensions. The m and n subscripts are the nodes in the x and y directions. The following equations represent the energy balance in transient conditions, expressed using the finite difference formulation and solved via the implicit method.

Heating fluid

$$\begin{gathered} m_{HF}{Cp}_{HF}\left(T_{m-1,n}^{i+1}-T_{m,n}^{i+1}\right)+m_{HF}{Cp}_{HF}\left(T_{m+1,n}^{i+1}-T_{m,n}^{i+1}\right) \\ \hspace{1em}\hspace{1em}\hspace{1em}\hspace{1em}\hspace{1em}\hspace{1em}\hspace{1em}\hspace{1em}\hspace{1em}\hspace{1em}\hspace{1em}\hspace{1em}\hspace{1em}\hspace{1em}\hspace{1em}+{htc}_{HF}^{i+1}∆A_y\left(T_{m,n+1}^{i+1}-T_{m,n}^{i+1}\right)+{htc}_{HF}^{i+1}∆A_y\left(T_{m,n-1}^{i+1}-T_{m,n}^{i+1}\right) \\ \hspace{1em}\hspace{1em}\hspace{1em}={\displaystyle \frac{{\rho }_{HF}{∆V}_{HF}\left(h_{HF}^{i+1}-h_{HF}^i\right)}{∆t}} \end{gathered}$$

(1)

where m, T, Cp, htc, and $p$ are the flow rate (kg/s), temperature (°C), heat capacity (kJ/kg °C), convection heat transfer coefficient of the heating fluid (kW/m² °C), and density (kg/m³), respectively; ΔV and ΔAy are volume element and the area in the y direction, respectively. $h_{m,n}^i$ and $h_{m,n}^{i+1}$ are the enthalpies of node m, n at times t_i = iΔt and t_i+1 = (i + 1)Δt, respectively. Δt is the step time.

PCM block in the sensible heat zone

$$\begin{gathered} {\displaystyle \frac{k_{PCM}∆A_x\left(T_{m-1,n}^{i+1}-T_{m,n}^{i+1}\right)}{∆x}}+{\displaystyle \frac{k_{PCM}{∆A}_x\left(T_{m+1,n}^{i+1}-T_{m,n}^{i+1}\right)}{∆x}} \\ \hspace{1em}\hspace{1em}\hspace{1em}\hspace{1em}\hspace{1em}\hspace{1em}\hspace{1em}\hspace{1em}\hspace{1em}\hspace{1em}\hspace{1em}\hspace{1em}\hspace{1em}\hspace{1em}\hspace{1em}+{\displaystyle \frac{k_{PCM}{∆A}_y\left(T_{m,n+1}^{i+1}-T_{m,n}^{i+1}\right)}{∆y}}+{\displaystyle \frac{k_{PCM}{∆A}_y\left(T_{m,n-1}^{i+1}-T_{m,n}^{i+1}\right)}{∆y}} \\ \hspace{1em}\hspace{1em}\hspace{1em}\hspace{1em}={\displaystyle \frac{{\rho }_{PCM}{∆V}_{PCM}\left(h^{i+1}-h^i\right)}{∆t}} \end{gathered}$$

(2)

PCM block in the latent heat zone

$$\begin{gathered} {\displaystyle \frac{k_{EFF,PCM}∆A_x\left(T_{m-1,n}^{i+1}-T_{m,n}^{i+1}\right)}{∆x}}+{\displaystyle \frac{k_{EFF,PCM}{∆A}_x\left(T_{m+1,n}^{i+1}-T_{m,n}^{i+1}\right)}{∆x}} \\ \hspace{1em}\hspace{1em}+{\displaystyle \frac{k_{EFF,PCM}{∆A}_y\left(T_{m,n+1}^{i+1}-T_{m,n}^{i+1}\right)}{∆y}} \\ \hspace{1em}\hspace{1em}\hspace{1em}\hspace{1em}\hspace{1em}\hspace{1em}\hspace{1em}\hspace{1em}\hspace{1em}\hspace{1em}\hspace{1em}\hspace{1em}+{\displaystyle \frac{k_{EFF,PCM}{∆A}_y\left(T_{m,n-1}^{i+1}-T_{m,n}^{i+1}\right)}{∆y}}={\displaystyle \frac{{\rho }_{PCM}{∆V}_{PCM}\left(h^{i+1}-h^i\right)}{∆t}} \end{gathered}$$

(3)

where Δx and Δy represent the spacing between nodal points in the rectangular mesh, h is the enthalpy, k is the thermal conductivity (kW/m °C), and k_EFF is the effective conductivity and takes a value of k_L × 4. More details and validation can be found in [6].

Table 3. Properties used for PCM.

Component	Tmelting (°C)	Heat Fusion (kJ/kg)	Cp (kJ/kg °C)	k (W/m K)	$\mathit{p}$ (kg/m3)	Thermal Diffusivity (m2/s, 1 × 104)	Energy Density (kJ/m3)
ERY	118.1	337	cpS = 1.34 (20 °C) cpL = 2.87 (150 °C)	kS = 0.89 (20 °C) kL = 0.33 (140 °C)	$p$S = 1440.4 (20 °C) $p$L = 1289.1	αS = 4.61 αL = 0.89	edS = 0.234 edL = 0.261
MCHH	115.1–117.4	166.9	cpS = 1.83 (100 °C) cpL = 2.57 (120 °C)	kS = 0.70 (110 °C) kL = 0.63 (120 °C)	$p$S = 1595.5 (20 °C) $p$L = 1455.7	αS = 2.39 αL = 1.6	edS = 0.105 edL = 0.115

shows the property values of ERY and MCHH [19] used in the calculations. An average temperature of 116.25 °C was considered the melting point of MCHH. MCHH has the advantage of having higher thermal diffusivity in the liquid phase than ERY. The cost of the MCHH is USD 174/kg (>99.0% wt.); in contrast, the ERY has a cost of USD 7755/kg [20]. On the other hand, ERY has almost double the value of thermal diffusivity in the solid phase. Additionally, the energy density suggests that it has a higher energy storage capacity. El-Sebaii et al. [21] reported that the MCHH has a thermal stability of 1000 cycles, as analyzed for solar cooking, while Agyenim et al. [22] reported a 7.5% reduction in latent heat after 20 cycles.

Figure 3 shows the energy of LHST using ERY and MCHH, and

Table 4. Operating conditions and size of the LHST.

Parameter	Value
PCM Thickness (dth), m	0.05
Total flow rate of heating fluid, kg/s	0.55
Wide channel, m	8.99
Height channel, m	0.005
Longitude channel, m	0.16
Isolation, m	0.10

shows the dimensions and input operation conditions. Erythritol stores more energy (27,790 kJ) than MCHH (21,700 kJ). This energy corresponds only to one of the four blocks of the tank; however, the phase change takes longer, as MCHH takes 100 min, while ERY takes 180 min. It can be observed that the slope changes because the PCM heats up (so it not only stores latent heat but also sensible heat), and this decreases the temperature difference between the heating fluid and PCM and reduces the heat transfer.

3. Results

Three days of operation were selected (2069 to 2141 h) because they yielded good total horizontal radiation and corresponded to the hot season in Temixco, Morelos, Mexico, as shown in Figure 4a. Figure 4b shows the generator energy and COP of the sorption cooling systems at 20 m² of A_SC and 0.05 m of dth for MCHH. The COP remains almost constant for ABC at approximately 0.7 due to the short range of operation of the generator temperature, resulting in a generator energy of 5.50 × 10⁵ kJ. In contrast, the COP changes from 0.4 to 0.6 for the ADC, with an energy value of approximately 6.64 × 10⁵ kJ. The lowest value of COP for ADC occurs when the storage tank temperature reaches the minimum operating temperature (usually at the end of the day), and the heating system then turns on to maintain the minimum operational temperature (65 °C).

The initial conditions of the LHST were a temperature of 115 and 117 °C for MCHH and ERY, respectively, a solar collector tilt angle of 7° (optimized value), and Δt = 180 s for the simulation time step. The parametric analysis consisted of evaluating A_SC at 10 to 30 m² and 5 to 20 m², for ET and PT, respectively, and 0.05, 0.10, and 0.15 m of dth.

Figure 5 shows the behavior of the solar collector temperature (T_SC) at the central (T_PCM,C) and edge (T_PCM,E) locations of the PCMs, as well as the mass flow rate of the solar collector (m_SC) for the sorption systems, with an A_SC of 20 m². It can be observed that the solar collector system operates longer with ERY because the PCM retains more energy. Consequently, the LHST should be prevented from exceeding the maximum allowable temperature limits. The four systems illustrate how the PCM remains in a phase change state over the three days evaluated (with almost constant temperature), which prevents the reduction of supercooling during the night.

Figure 6 shows the solar collector energy (Q_SC) as a function of A_SC using evacuated tubes from 10 to 30 m² at different PCMs and sorption cooling systems. The minimum values of Q_SC were approximately 2.50 × 10⁵ kJ with an A_SC of 10 m² for all systems. The maximum values were around 6.81 × 10⁵ and 7.4 × 10⁵ kJ for MCHH and ERY, respectively, with a dth of 0.15 m and an A_SC of 30 m² in both sorption systems. This means that it is the maximum amount of solar energy that can be obtained by the ET.

Figure 7 shows the Q_SC as a function of solar area collector using a PT at different PCMs and sorption cooling. The system ADC-MCHH (Figure 7a) and ADC-ERY (Figure 7b) obtained Q_SC maximum values of 9.96 × 10⁵ and 10.60 × 10⁵ kJ, respectively, with a dth of 0.15 m with 20 m² of A_SC. ERY obtained higher Q_SC values than MCHH because it has sufficient capacity to absorb the heat supplied by the solar collector.

The ABC-MCHH (Figure 7c) system has Q_SC values from 4.88 × 10⁵ to 7.89 × 10⁵ and 5.41 × 10⁵ to 8.69 × 10⁵ kJ from 10 to 20 m² of A_SC with dth of 0.05 and 0.15 m, respectively. The ABC-ERY system (Figure 7d) is evaluated from 10 to 13 m² of A_SC, since a solar fraction above 0.96 is obtained (as seen later), the Q_SC values range from 5.40 × 10⁵ to 6.40 × 10⁵ for a dth of 0.05 m, and 5.40 × 10⁵ to 6.98 × 10⁵ kJ for dth of 0.10 and 0.15 m, respectively.

Figure 8 compares the solar fraction as a function of the solar collector area for ET with different dth and PCMs, as well as sorption systems. The solar fraction is defined as the ratio of solar energy-to-the total energy from both the solar and heating systems. The ADC-MCHH and ADC-ERY system (Figure 8a,b) obtained a similar value of SF of 0.25 with an A_SC of 10 m² for all the dth, while the maximum SF was 0.62 and 0.64 for MCHH and ERY, respectively, with a dth of 0.15 m and A_SC of 30 m².

The SF of the ABC systems (Figure 8c,d) had a maximum value of 0.61 and 0.79 for MCHH and ERY, respectively, with a dth of 0.05 m and an A_SC of 10 m². In contrast, the maximum SF was 0.90 and 0.93 for MCHH and ERY, respectively, with a dth of 0.15 m and an A_SC of 30 m². It can be observed that ERY obtained slightly better SF than MCHH.

Figure 9 shows the solar fraction as a function of the solar collector area for parabolic trough collectors at different thicknesses of PCMs and sorption systems. The ADC-MCHH (Figure 9a) obtained an SF value from 0.55 to 0.76 and 0.53 to 0.83 for a dth of 0.05 and 0.15 m, respectively, from A_SC of 10 to 20 m². The ADC-ERY (Figure 9b) obtained a value of 0.51 to 0.78, and 0.51 to 0.84, for dth with 0.05 and 0.15 m from 10 to 20 m² of A_SC, respectively, In the ABC-MCHH system (Figure 9c), it is shown that the SF does not reach 1 at values above 13 m² with dth of 0.10 and 0.15 m as A_SC is incremented, because despite the solar collector energy is increasing (Figure 7c), the heater system almost obtained similar values (from 1.04 × 10⁴ to 3.91 × 10⁴ kJ with A_SC from 12 to 20 m²), this means the heat dissipation in the PCM is very poor, and the energy is kept in the LHST, while the ABC-ERY system (Figure 9d) does reach a value of 1 with A_SC of 13 m² with 0.15 m.

Table 5. Maximum values of solar fraction for each system.

Case	SFMAX (dth = 0.05 m)	SFMAX (dth = 0.10 m)	SFMAX (dth = 0.15 m)
ADC-ET-MCHH	0.53	0.61	0.62
ADC-ET-ERY	0.60	0.64	0.64
ADC-PT-MCHH	0.76	0.83	0.83
ADC-PT-ERY	0.78	0.81	0.84
ABC-ET-MCHH	0.61	0.81	0.90
ABC-ET-ERY	0.79	0.92	0.93
ABC-PT-MCHH	0.89	0.94	0.96
ABC-PT-ERY	0.92	1.00	1.00

presents a summary of the maximum values of solar fraction at different dth with an A_SC of 20 and 30 m² for PT and ET, respectively, except for the ABC-PT-ERY system (A_SC of 13 m²). The minimum and maximum values of SF_MAX were for the ADC-ET-MCHH (0.53) and ABC-PT-ERY (1.0) systems, respectively. When the systems are compared under similar conditions, in general, the ABC obtained a better SF_MAX than the ADC system, ERY obtained a better SF_MAX than MCHH, and PT obtained better results than ET, despite using less A_SC.

Figure 10 shows the effectiveness of the LHST as a function of time for the ABC-ERY and ABC-MCHH systems with a dth of 0.10 m and an A_SC of 13 m². This parameter was estimated using instantaneous temperature, and it was defined as the ratio of the actual heat transfer rate-to-the maximum possible. The effectiveness is in discharge mode (ε_DISCH) and is calculated when only the pump of the storage tank is turned on and enters the LHST, and the pump of the ABC is turned off. The ε_DISCH obtained an average value of 0.56 in both systems.

4. Discussion

The results indicate that ERY exhibits a higher energy storage capacity, accumulating 27,790 kJ over 180 min, compared to MCHH, which stores 21,700 kJ in 100 min. This is primarily due to ERY’s higher heat of fusion (337.0 kJ/kg) compared to MCHH (166.9 kJ/kg). However, ERY’s lower thermal conductivity results in slower heat dissipation. This characteristic could limit the ERY in applications requiring rapid thermal response. Conversely, MCHH’s higher thermal conductivity (in liquid phase) facilitates a slightly faster phase change dynamic, making it more suitable for systems with high transient thermal demands. The selection of PCM should thus be aligned with operational requirements, the ERY for maximizing energy storage, and MCHH for applications prioritizing rapid heat transfer. The results obtained for this investigation showed similar effectiveness in both PCMs, and the performance of the ERY obtained good results in the SF.

Increasing the dth from 0.05 m to 0.15 m enhances the energy storage capacity but exacerbates heat transfer limitations in ERY due to its low thermal conductivity. This suggests that thicker PCM layers are beneficial for maximizing storage but require enhanced heat transfer mechanisms, such as fins or composite PCMs, to improve thermal response in ERY-based systems. The optimal PCM thickness should strike a balance between storage capacity and heat transfer efficiency to maximize system performance.

The ABC system maintains a stable COP of approximately 0.70, while the ADC system’s COP fluctuates between 0.40 and 0.60. This variability in ADC performance is attributed to its lower activation temperature range (65–95 °C) compared to ABC (108.9–116.1 °C), which necessitates auxiliary heating when storage tank temperatures drop below operational thresholds, particularly during periods of low solar input. The ABC system’s consistent COP reflects its suitability for stable cooling demands but requires high-temperature solar collectors, potentially increasing capital costs. In contrast, ADC systems benefit from compatibility with simpler collectors, such as ET, but their reliance on auxiliary systems reduces overall efficiency. This highlights a critical trade-off between operational simplicity and thermal efficiency, necessitating further analysis to optimize system design for specific applications.

5. Conclusions

This work shows an analysis of adsorption and ABC systems using parabolic trough and evacuated tube solar collectors coupled to a rectangular LHST with MCHH and ERY in dynamic conditions, as evaluated using TRNSYS software over three days. The following conclusions were drawn.

A simulation of the LHST was performed, yielding a total energy of 27,790 kJ in 180 min with ERY and 21,700 kJ in 100 min with MCHH, respectively. The energy consumed includes latent and sensible heat with a thickness of 0.05 m.

The operation time of the solar energy collector was higher for ADC than for ABC because the temperature of the LHST was lower due to ADC’s higher energy consumption, so the pump was turned on for a longer period.

The maximum Q_SC (with ET) was around 6.81 × 10⁵ and 7.4 × 10⁵ kJ for MCHH and ERY, respectively, with a dth of 0.15 m and an A_SC of 30 m².

The maximum Q_SC (with PT) was 9.96 × 10⁵ and 10.60 × 10⁵ kJ for ADC-MCHH and ADC-ERY systems with a dth of 0.15 m and A_SC of 20 m², respectively. The system ABC-MCHH obtained a maximum Q_SC of 8.69 × 10⁵ kJ with dth of 0.15 m and an A_SC of 20 m², and ABC-ERY obtained a maximum value of 6.98 × 10⁵ kJ with dth of 0.15 m and an A_SC of 13 m².

The maximum SF (with ET) was 0.62 and 0.64 for the ADC-MCHH and ADC-ERY systems, respectively, with a dth of 0.15 m and an A_SC of 30 m², while with absorption cooling, higher SF values of 0.90 and 0.93 for MCHH and ERY, respectively, were obtained at similar conditions.

The maximum SF (with PT) was 0.83 and 0.84 for ADC-MCHH and ADC-ERY systems, respectively, with a dth of 0.15 m and A_SC of 20 m². The ABC-MCHH system obtained values higher than 0.95 above A_SC of 13 m²; however, it did not reach a value of 1 at higher A_SC, because the heat dissipation in the PCM is very poor and the energy is retained in the LHST. The ABC-ERY system did reach a value of 1 with A_SC of 13 m² with 0.15 m.

The control system of the LHST limited the solar energy harvesting time due to the temperature difference between the center and the edge of the PCM, which could fall under supercooling conditions; therefore, the application of advanced heat transfer techniques is important.

The SF of the evacuated tube solar collectors would be improved by using another PCM with a lower melting point for the adsorption cooling.

Erythritol was able to meet the energy requirements only with solar collectors due to its high energy density; however, the cost per kilogram was 44 times higher than MCHH, and it exhibited a reduction in latent heat after a few thermal cycles. A deep study should be conducted for ERY, including an economic analysis.