Modelado y análisis del sistema de refrigeración solar con eyector para aplicaciones de climatización

1 Equation Chapter 1 Section 1 Trabajo Fin de Máster Sistemas de Energía Térmica Modelado y análisis del Sistema de Refrigeración Solar con Eyector para aplicaciones de climatización Dpto. Ingeniería Energética. Escuela Técnica Superior de Ingeniería Universidad de Sevilla Autor: Miguel Ávila Gutiérrez Tutor: José Manuel Salmerón Lissén Tutor: Bernardo Peris Pérez Sevilla, 2020 2 3 3 Modelado y análisis del Sistema de Refrigeración Solar con Eyector para aplicaciones de climatización Trabajo Fin de Máster Sistemas de Energía Térmica Modelado y análisis del Sistema de Refrigeración Solar con Eyector para aplicaciones de climatización Autor: Miguel Ávila Gutiérrez Tutor: José Manuel Salmerón Lissén Tutor: Bernardo Peris Pérez Dpto. Ingeniería energética Escuela Técnica Superior de Ingeniería Universidad de Sevilla Sevilla, 2020 4 5 5 Modelado y análisis del Sistema de Refrigeración Solar con Eyector para aplicaciones de climatización Trabajo Fin de Máster: Modelado y análisis del Sistema de Refrigeración Solar con Eyector para aplicaciones de climatización Autor: Miguel Ávila Gutiérrez Tutor: José Manuel Salmeron Lissén Tutor: Bernardo Peris Pérez El tribunal nombrado para juzgar el Proyecto arriba indicado, compuesto por los siguientes miembros: Presidente: Vocales: Secretario: Acuerdan otorgarle la calificación de: 6 Sevilla, 2020 El Secretario del Tribunal 7 7 Modelado y análisis del Sistema de Refrigeración Solar con Eyector para aplicaciones de climatización A mi familia A mis maestros 8 9 9 Modelado y análisis del Sistema de Refrigeración Solar con Eyector para aplicaciones de climatización Agradecimientos Agradecer a mis tutores por haberme dejado formar parte de lo que ha sido mi primer contacto con la investigación. Difícil y fructuoso a partes iguales, me quedo con todo lo aprendido y más aún con lo que queda por aprender. Aunque sea un año extraño para acabar los estudios, afronto con ganas lo que viene ahora. Por último, no podría despedirme de mi etapa como estudiante sin darle las gracias a mi familia, y a Cristina, por aguantarme y sufrirme cuando más lo he necesitado. Miguel Ávila Gutiérrez Sevilla, 2020 16 ÍNDICE DE TABLAS Tabla 1-1. Traditional and low GWP refrigerants (fluorinated and natural) proposed for the SERS use. 22 Tabla 4-1. Summary of ejector and ERS performance of theoretical results calculated with the present model, compared with the experimental data from Shestopalov et al. (2015) 12 Tabla 5-1. Ejector geomtries 14 Tabla 5-2. Optimal values for the three multi-ejectors using R600 19 17 17 Modelado y análisis del Sistema de Refrigeración Solar con Eyector para aplicaciones de climatización 18 ÍNDICE DE FIGURAS Fig. 2-1. Application of SERS as pre-cooling system for Air Handling Unit. 24 Fig. 2-2. SERS: (a) cycle architecture; (b) logP-h diagram using R600. 24 Fig. 3-1. Operation modes of an ejector: (a) Primary pressure remains constant; (b) Condensing back pressure remains constant. 25 Fig. 3-2. Schematic diagram of ejector and the corresponding pressure and velocity profile. 27 Fig. 3-3. Calculation flowchart of the present model. 33 Fig. 4-1. (a) Critical entrainment ratio and (b) critical condensing pressure comparison of the present model with the experimental results of Huang et al. (1999) 11 Fig. 4-2.(a) Critical entrainment ratio and (b) critical condensing pressure comparison of the present model with the experimental results of Li et al. (2018) 13 Fig. 4-3. Comparison of the ejector subcritical performance of the present model and the experimental data provides by Li et al. (2018) for different levels of secondary temperatures (Te) and primary temperatures (Tg) (a) Pp=23.66 bar; (b) Pp=26.35 bar. Both inlet conditions primary and secondary are saturated. 13 Fig. 5-1. Critical entrainment ratio and (b) critical condensing temperature as a function of generator temperature at different area ratios with R134a and Te=10°C. 15 Fig. 5-2. (a) Critical entrainment ratio and (b) critical condensing temperature as a function of condensing temperature at different evaporator temperatures with R134a and Te=90°C. 15 Fig. 5-3. (a) Critical entrainment ratio; (b) critical COP and (c) critical condensing temperature as a function of generator temperature using different working fluids with Ar=5.04 and Te=10°C. 17 Fig. 5-4. Critical entrainment ratio; (b) critical COP and (c) critical condensing temperature as a function of generator temperature using different working fluids with Ar=5.04 and Te=28°C. 18 Fig. 5-5. Performance hot season-based of each optimized ejector and the multiejector in warm climate with Tg=120 using R600: (a) COP; (b) COPel 19 Fig. 5-6. Performance hot season-based of three different multi-ejector in warm climate for different generator temperatures using R600: (a) COP; (b) COPel 20 Fig. 5-7. (a) Seasonal COP; (b) Seasonal COPel compared with COPVCC as a function of generator temperature using different refrigerants. 21 19 19 Modelado y análisis del Sistema de Refrigeración Solar con Eyector para aplicaciones de climatización 20 1 INTRODUCTION uildings are responsible for about 40% of the EU energy consumption and 36% of the greenhouse gas emissions (European Commission, 2020). This is mainly because 35% of the EU buildings are over 50 years old, and almost 75% of the building stock is considered energy inefficient. To boost the improvement of the buildings energy performance a legislative framework was established, including the Energy Performance of Buildings Directive (EPBD) (2010/31/EU, 2010) and the Energy Efficiency directive (2012/27/EU, 2012). In particular, the EPBD establishes that all new buildings from 2021 must be Nearly Zero Energy Buildings (NZEB), which means that a very low amount of energy should be demanded, and a significant extent of renewable energy sources should be considered in the generation. In this regard, countries are committed to minimizing their primary energy dependence by including technologies to benefit from renewable energy sources for heating, cooling, and electricity production, besides other purposes (Kong et al., 2020). For instance, Spain has recently updated the legislation of buildings energy efficiency (Royal Decree- Lay 732/2019, 2019), stipulating that over 70% of the Domestic Hot Water (DHW) demand of new buildings must be supplied by renewable energies. Considering that Spain has become a solar energy-intensive country, most of the future DHW demand of NZEB would come from a solar origin by using solar thermal collectors (Pablo-Romero et al., 2013). Commonly, solar thermal energy fields designed in a warm climate to provide hot water for a district heating network, or DHW for a building, are oversized during summer months (Tian et al., 2017). Then, solar collectors require being cooled down by rejecting the waste heat to the ambient, while increasing the air conditioning demand of the building. An efficient alternative is to produce solar cooling by recovering the wasted heat (Pranesh et al., 2019). It has to be mentioned that the cooling demand in buildings is expected to grow until the triple by 2050 (IEA, 2018), requiring energy efficiency measures to be NZEB. Indeed, in 2019, about 1800 solar cooling systems were installed worldwide using absorption chillers (Weiss and Spörk-Dür, 2019). Among solar cooling technologies, the Solar-driven Ejector Refrigeration System (SERS) is an emerging technology in comparison to the mature absorption chillers. Some advantages of the SERS are its structural simplicity, lower purchasing costs and maintenance (Varga et al., 2017). Moreover, the SERS allows a flexible selection of working fluids, being able to design more environmentally friendly by using natural and fluorinated refrigerants with low Global Warming Potential (GWP). On the other hand, the major drawback of the SERS is the lower performance reachable compared to absorption systems. Nonetheless, recently, Braimakis (2021) argued that under certain circumstances, especially low scales, the SERS may be costcompetitive if their lower equipment costs compensate the capital expenditure of the higher solar field required. Similarly, numerous investigations focused on the advantages of the SERS for solar cooling as a future competitor of ammonia systems. Roman and Hernandez (2011) evaluated the SERS performance using R290, R600, R600a, R152 and R134a. The analysis was conducted considering evaporating temperatures ranging between 5 and 15 ºC, and generation temperatures from 70ºC to 95ºC. Galindo et al. (2020) focused on the use of the low GWP refrigerants R1234yf, R1234ze, and R600a. The analysis was conducted assuming evaporating temperature of 13ºC, condensing temperature of 40ºC, and generator temperatures up to 125 ºC, identifying R1234yf as the best candidate in terms of overall system efficiency. Guo and Shen (2009) proposed the SERS adoption in an office building for air-conditioning. The refrigerant proposed was R134a, for 8ºC of evaporating temperature, 65ºC of generator temperature, and condensing conditions from 27 ºC to 34ºC. Thereby, researchers concluded that electricity energy savings of 75% may be reached by using the SERS. Pridasawas and Lundqvist (2004) assessed the SERS using R600 at 10 ºC of evaporating temperature, 37ºC of condensing temperature, and 90 ºC in the generator, reaching a COP of 2.5. However, authors outlined that if the thermal energy is obtained for free, such as rejected waste heat, the system performance could be B 21 21 Modelado y análisis del Sistema de Refrigeración Solar con Eyector para aplicaciones de climatización defined as the ratio between cooling capacity and electrical consumption (COPel), reaching a value of 32. With respect to the SERS performance enhancement, two streams outline in the literature. First, the use of more efficient and sustainable working fluids. And second, the ejector performance improvement by multiple ejector modules or variable-geometry devices. Regarding the use of low GWP refrigerants, numerous alternatives were investigated in the literature to replace traditional working fluids. Fang et al. (2017) proposed hydro-fluoro olefin (HFO) refrigerants R1234yf and R1234ze(E) as a drop-in replacement for R134a in the ejector refrigeration cycle. Zhang and Mohamed (2015) demonstrated that R290, R600a, R601, R1234yf, R1234ze and R744 exhibit inherent advantages for cooling applications to replace R134a. Mwesigye and Dworkin (2018) proposed R600, R1234ze(Z), R1233zd(E), RE245fa2 and R1336mzz(Z) as replacements for the commonly used R141b and R245fa, demonstrating that greater values of entrainment ratio, cooling capacity and COP are reachable through a suitable refrigerant selection. A non-exhaustive review of proposed low GWP alternatives, including fluorinated and natural working fluids, to traditional refrigerants is collected in Table 1. Among thermophysical properties, the critical temperature is commonly used for screening the working fluid. This temperature sets the maximum evaporating temperature in the generator when subcritical cycles are implemented, which are preferable when R744 is not adopted (Besagni et al., 2016). Moreover, the fluid slope classifies the refrigerant according to the saturated vapor line as isentropic, dry and wet. Specifically, the slope has been outlined as a crucial parameter for the ejector model selection. Thus, since the first 1-D model developed by Keenan et al. (1950), ideal gasbased models were extensively used for the study of air, steam, but also dry slope refrigerants. However, more recent investigations demonstrate that a two-phase state may occur inside the nozzle of the ejector, being still more probable in wet fluids, requiring real fluid ejector models to improve the accuracy of the results (Chen et al., 2017). As mentioned, the SERS performance directly depends on the ejector efficiency, which, in turn, is strongly affected by its internal geometry and the operating conditions for a specific refrigerant (Tashtoush and Nayfeh, 2020). Despite the simplicity of using a single fixed ejector, current investigations to improve the ejector efficiency propose the adoption of multi-ejector modules as a suitable solution. In particular, it consists of using multiple ejectors connected in a parallel arrangement, enabling or disabling modules to modify the system capacity and performance. In this way, Aligolzadeh and Hakkaki-Fard (2019) demonstrated that by means of three modules a COP improvement of 70% is achievable. A comprehensive review of the latest developments in multi-ejectors was conducted by Gullo et al. (2019). In addition to the use of multiple modules, the ejector geometry can also be adapted by introducing moving elements. Thus, Pereira et al. (2014) tested the influence of the Area Ratio (AR) on the ejector performance by introducing a spindle inside the nozzle, reporting COP improvements up to 85% compared to the fixed geometry ejector under specific operating conditions. Van Nguyen et al. (2020) analyzed the AR and the nozzel exit position (NXP) using high-precission stepper motors. The results evidenced that variable geometry ejectors penalize the advantages of simplicity and low maintenance associated with SERS, but compensate for its low COP by maximizing the ejector performance according to the operating conditions. The reviewed literature manifests an increasing interest in the SERS investigation as a slightly less efficient but simpler solar cooling alternative to absorption systems. The possibility of using novel low GWP refrigerants and to achieve a flexible operation by means of multi-ejectors and variable geometry ejectors turns the SERS a feasible solution to be considered in future NZEB. In this regard, the SERS may benefit from the waste heat rejected by solar thermal energy fields during warmer seasons, when the DHW demand drops as the air conditioning demand increases. However, notice that the SERS performance is constrained when low to medium temperature solar thermal collectors, frequently installed for DHW production, are used in warm climates to produce cooling water at equivalent temperatures to chillers. In light of this, this paper proposes a novel application that reformulates the SERS use as a fresh air pre-cooling system for air handling units (AHUs). Thereby, greater evaporating temperatures that enable the SERS adoption are obtained and, consequently, a reduction on the heat load of the building would be achieved by means of the pre-cooling of the fresh air supply. In such way, since the heat load, which the chiller is working against, is reduced, the chiller will be working for more time at partial load and the total energy consumption along the year would be also reduced. To the best of authors knowledge, none of the current papers have analyzed the SERS seasonal performance considering the severe operating conditions of the proposed application, which correspond to warm climate temperatures in both evaporator and condenser, as well as low to medium temperatures in the generator. 22 Tabla 1-1. Traditional and low GWP refrigerants (fluorinated and natural) proposed for the SERS use. No. Type Mol. Mass (kg/kmol) Boiling T (ºC) Critical T (ºC) Critical P (bar) ASHRAE 34 Fluid slope ODP GWP Traditional R134a HFC 102.03 -26.1 101.1 40.6 A1 wet 0 1430 R141b HCFC 116.95 32 204.4 42.1 A2 dry 0 725 R245fa HFC 134.05 15 153.9 36.5 B1 dry 0 1030 Fluorinated R1234yf HFO 114.04 -29.5 94.7 33.8 A2L isentropic 0 4 R1234ze(E) HFO 114.04 -19 109.4 36.3 A2L isentropic 0 6 R1243zf HFO 96.05 -25.4 103.8 35.2 A2L isentropic 0 <1 Natural R290 HC 44.10 -42.1 96.7 42.5 A3 wet 0 20 R600a HC 58.12 -11.7 134.7 36.3 A3 isentropic 0 4 R600 HC 58.12 -0.5 152 38 A3 isentropic 0 4 R717 NH3 17.03 -33.3 132.3 113.3 B2L wet 0 0 For this purpose, the rest of the paper is organized as follows. Section 2 presents the proposed application, presenting a schematic of the facility, the thermodynamic cycle of the SERS, and the parameters considered in the performance analysis. Section 3 describes the numerical model developed, with special emphasis on the two-phase flow ejector and its geometry, and the performance analysis methodology. Section 4 validates the model of the ejector considering the most used traditional refrigerants. Section 5 conducts a performance analysis considering low GWP refrigerants, the geometry of single and multi-ejector modules, and a seasonalbased simulation. And, finally, section 6 collects the main conclusions of the study. 23 2 APPLICATION DESCRIPTION HU is used to condition and to distribute the air within the building as part of the Heating, Ventilation, and Air-Conditioning (HVAC) system. The unit is commonly installed on the roof of the building to take benefit from solar thermal energy. There are numerous possibilities of AHU setups, including rooftop packaged units, all fresh air handling units, or heating and cooling units, among others. In addition, a specific setup may include a ducted plate heat exchanger, free-cooling valves to mix fresh and exhaust air, an adiabatic cooling system, besides a fresh air pre-cooling system. Accordingly, an AHU setup in which the SERS can be included is presented in Fig. 1. In particular, an fullfresh air handling unit, which often includes a ducted plate heat exchanger for exhaust air energy recovery (Han et al., 2018), is illustrated. First, the cooling coil of the AHU is connected to a chiller, which may produce cooling water by means of a vapor compression cycle or an absorption system, which is preferable for a NZEBs. Second, a fresh air precooling system by adopting the SERS is considered, which uses indirect heat transfer loop to accomplish safety criteria of not using direct expansion evaporators with flammable or toxic refrigerants in HVAC applications. Thereby, the fresh air temperature that enters into the ducted plate heat exchanger is decreased. It has to be clarified that the reduction of the recuperator inlet temperature, directly affects to the heat exchanger effectiveness, but produces a positive effect for the application in the form of fresh air outlet temperature reduction. Notice that the fresh air pre-cooling SERS may also be included in many other AHU arrangements. Accordingly, this study focuses on the application described as guidelines able to be extrapolated to different application typologies. The architecture of the SERS is represented in Fig. 2.a. As observed, this configuration does not necessarily require a compressor to operate. Indeed, the primary stream is impelled by the pump, and it is evaporated in the generator by means of the solar thermal energy recovered. On the other hand, the low-pressure vapor enters the suction chamber of the ejector owing to the Venturi effect and leaves the diffuser mixed with the primary stream at an intermediate pressure, which corresponds to the condensing pressure of the cycle. This effect is depicted in Fig. 2.b in a logP-h diagram using R600 as working fluid. Furthermore, this figure illustrates a particularity of the application proposed in this study A Application description 24 Fig. 2-1. Application of SERS as pre-cooling system for Air Handling Unit. Specifically, there is a heat transfer between outdoor fresh air and the refrigerant of both, condenser and evaporator. The condenser uses the outdoor air to reject the waste heat of the cycle, while the evaporator allows cooling down the stream of fresh air that enters the AHU. The difference between condensing and outdoor temperatures is assumed 10 K, according to commercial chillers design criteria. This temperature difference may be kept constant by controlling the discharge pressure of the pump. On the other hand, the difference between outdoor and evaporating temperatures may vary according to the supplied air flow rate to the AHU. Considering that an full fresh AHU with a constant volume control is used, this temperature difference may be set constant as 15 K, which must be this high due to the intermediate water loop, in order to achieve a fresh air temperature reduction. Nevertheless, the implications of using variable air volume controls, e.g. including a damper and an exhaust air recirculation, are also considered and discussed below. (a) (b) Fig. 2-2. SERS: (a) cycle architecture; (b) logP-h diagram using R600. NZEB Chiller SERS Pre-cooling coil AHU Fresh air Ducted Plate Heat Exchanger Cooling coil Generator Condenser Evaporator Ejector Expansion valve Pump 1 2 3 4 5 6 3 4 1=s0 2=c 6=p0 5 py sy m 1 5 50 150 250 350 450 550 650 750 850 Log P (bar) h (kJ/kg) Toutdoor +10 K -15 K 25 3 SYSTEM MODELING 3.1 Two-phase flow ejector Ejectors can be classified into two main types by the mixing concept: constant-pressure mixing (CPM) ejectors and constant-area mixing (CAM) ejectors (Pianthong et al., 2007). Generally, CPM has better performance than CAM ejectors, according to Besagni et al. (2016). Thus, the theoretical model of CPM ejector is gaining the attention of several scientists and it is widely studied for being more suitable to use in SERS. For that reason, the CPM ejector is the type chosen for this study. The main components of the configuration, the velocity and pressure profiles through the ejector are represented in Fig. 4. The ejector modeling is based on equations gathered by Chen et al. (2017) to determine performance in the critical mode of operation. The subcritical mode is governed by the constant-pressure disturbing proposed by Li et al. (2017). Overall modes are adapted to the two-phase speed of sound calculations. The following hypotheses are also considered: • One-dimensional and steady flow inside the ejector. • Kinematic energy at suction and discharge ports is negligible, as well as the heat transfer between the ejector and the environment. • Frictional and mixing losses are considered through empirical coefficients. • The pressure remains constant in the mixing chamber for both, under the critical and subcritical mode of operation, from the section y-y and section x-x to the chocking phenomenon, respectively. • Before the hypothetical throat, the two streams are independent flows. When the ejector operate at the critical mode, the second stream reaches the speed of sound and begin to mix with the primary flow at section y-y with a uniform pressure Two key operational points defined the whole operating range of the supersonic ejector, as depicted the Fig. (3), based on the study provides by Besagni et al. (2016). In the critical mode, the entrainment ratio remains constant and equal to the entrainment ratio at the critical point since the two flow streams are chocked. This condition is met while P_c≤P_cc. In the subcritical mode, only the primary flow is chocked and the entrainment ratio decrease in linear proportion to zero with the backpressure. The ejector work in this range as while as P_cc<P_c<P_cb. In the back-flow mode, the secondary flow is reversed resulting in ejector malfunction, while P_cb≤P_c. (a) (b) Fig. 3-1. Operation modes of an ejector: (a) Primary pressure remains constant; (b) Condensing back pressure remains constant. It is should be noted that at the critical mode, the secondary flow achieves the maximum flow rate and it is Back Pressure (Condensing pressure, Pc) Pbc Critical mode Subcritical mode Back-flow mode Critical point Breakdown point Pcc Entrainment ratio (ω) ω=constant ω<0 Primary Pressure (Pp) Critical mode Subcritical mode Back-flow mode Primary flow Mass flow rate Secondary flow System modeling 32 3.1.3 Speed of sound Considering that the working fluids are entering the primary nozzle in saturation conditions, the sonic velocity may occur inside the liquid-vapor region, especially in the case of wet refrigerants. The critical velocity at chocked flow conditions has a non-linear dependency of the mixture quality, the two-phase flow pattern, and the slip between liquid and gas phase velocities (Petrovic et al, 2016). Angielczyk et al. (2010) suggested the Homogeneous Relaxation Model (HRM) to predict the speed of sound in the two-phase region, which resulted much more consistent than the Homogeneous Equilibrium Model (HEM). Therefore, the model proposed by Lund and Flatten (2010) is used. 𝑎−2 = 𝑎𝑤 −2+𝜌 𝑇·𝐶𝑝,𝑣 ·𝐶𝑝,𝑙(𝜁𝑙−𝜁𝑣)2 𝐶𝑝,𝑣+𝐶𝑝,𝑙 (39) 𝑎𝑤 −2 = 𝜌· ( 𝜖𝑣 𝜌𝑣·𝑎𝑣 2−𝜖𝑙 𝜌𝑙·𝑎𝑙2) (40) 𝜁𝑘=(𝜕𝑇 𝜕𝜌)=𝑇·𝛽𝑘·𝑣𝑘 𝑐𝑝,𝑘 (41) 𝐶𝑝,𝑘 =𝜌𝑘·𝜖𝑘·𝑐𝑝,𝑘 (42) 3.2 Two-phase flow ejector A flow chart of the calculation procedure for the SERS is presented in Fig. (5). Once the point of operation of the ejector is known, the COP of the system and the thermal powers of each heat-exchanger are calculated as follows. The total heat rejected to the environment in the condenser is calculated as: 𝑄󰇗𝑐𝑑 =(𝑚𝑝+𝑚𝑠)·(ℎ𝑐𝑑𝑖 −ℎ𝑐𝑑𝑜) (43) The work of the pump is given by W󰇗𝑝=𝑚𝑝·(ℎ𝑝𝑜−ℎ𝑝𝑖) (44) To ensure the energy equilibrium of the system, the next Eq. must be satisfied 𝑄󰇗𝑐𝑑 =𝑄󰇗𝑔+𝑄󰇗𝑒+W󰇗𝑝 (45) 33 Modelado y análisis del Sistema de Refrigeración Solar con Eyector para aplicaciones de climatización Fig. 3-3. Calculation flowchart of the present model. The COP of the ejector refrigeration system is defined as follow 𝐶𝑂𝑃= 𝑄󰇗𝑒 𝑄󰇗𝑔+W󰇗𝑝 (46) and using the Eqs. 42, 43 y 45 the Eq. 47 can be reformulated. 𝐶𝑂𝑃=𝜔·ℎ𝑒𝑜−ℎ𝑒𝑖 ℎ𝑔𝑜 −ℎ𝑓𝑜 (47) Additionally, two more Coefficients of Performances have been defined. The electrical COP, COPel, may be used to evaluate the electricity improvement of the SERS compared to a VCC (Braimakis, 2021). The seasonal COP of the SERS, COPseasonal, is defined as the ratio between the total cooling load and the total energy consumed over a year. In the present study, the seasonal COP is defined over a warm season instead of a year. 𝐶𝑂𝑃𝑒𝑙 =𝑄󰇗𝑒 W󰇗𝑝 (48) Thermodynamic inputs , Ejector geometry Nozzle throat Eqs. (1~6) Nozzle exit Eqs. (7~8) Adjust Until (Eq. 7) = f( , ) Guess Adjust Until (Eq. 1) = (Eq. 2) Suction inlet Eqs. (11~14) Guess Adjust Until (Eq. 11) = (Eq. 12) Hypothetical throat Eqs. (9~10) Secondary mass flow Eqs. (15~16) Mixing at constant section Eqs. (17~18) Shock wave Eqs. (19~21) Guess Adjust Until (Eq. 19) = f( , ) Guess Difusser Eqs. (22~25) ≥Yes No Constant pressure disturbing Eqs. (27~33) + Shock wave Eqs. (19~21) Guess Adjust Until (Eq. 19) = f( , ) Difusser Breakdown point Eqs. (34~37) No + Yes Yes No System modeling 34 𝑆𝑒𝑎𝑠𝑜𝑛𝑎𝑙 𝐶𝑂𝑃=(𝜔·𝛥ℎ𝑒 𝛥ℎ𝑔 )𝑤𝑎𝑟𝑚 𝑠𝑒𝑎𝑠𝑜𝑛 (49) Some assumptions are also considered in the thermodynamic cycle and the multi-ejector design. The ambient temperature at which cooling is required is in the range of 25-43°C. As the amount of hours that ambient temperature exceeds 41°C in the climate considered is negligible, the real range taken into account to decide at which points the ejectors are being designed to operate at optimal conditions is 25-41°C. According to Aligolzadeh and Hakkaki-Fard (2019), three ejectors could provide the optimum benefit in terms of COP. Then, with this range of ambient temperatures (ΔToutdoor=16) and three ejectors, the temperatures of design are 29, 33 and 37°C. The generator temperature is assumed to be constant and three differents values are considered. Thus, for each generator temperature, a multi-ejector is designed, with one fixed geometry ejector optimized for each outdoor design temperature. The optimization calculation is based on the method of the conjugate direction using EES. Furthermore, it is assumed that both the primary and the secondary inlets of the ejector are saturated without any kind of reheat. The loss coefficients of the ejector model are widely used in the literature, both constant and variable as a function of the geometry of the ejector and the operating conditions (Chen et al, 2017), (Li et al, 2017b). Due to the purpose of this study in which many fluids of different types are evaluated, the loss coefficients of the present ejector model are established in such a way that they do not depend on the type of working fluid, paying particular attention that the grade of accuracy achieved is acceptable 11 4 EJECTOR MODEL VALIDATION This section aims to check the validity of the model by comparing it with experimental results from different types of fluids. Except for the momentum transfer efficiency at the subcritical mode, 𝛷𝑚𝑏, which value change according to Eq. (33), the rest of the isentropic efficiencies and losses coefficients are adopted as constant (𝜂𝑡=0.92, 𝜂𝑠=0.85, 𝜂𝑝𝑦 =0.8, 𝜂𝑑=0.95,𝛷𝑚=0.95). Therefore, all these parameters do not depend on the type of working fluid so the model can be universally applied with any refrigerant. 4.1 R141b As presented in Figure (4-1), the present model and the experiment set carried out by Huang et al. (1999) are compared. Huang et al. tested R141b ejector performance for several combinations of nozzle throat areas, nozzle exit areas, and mixing chamber areas. The area ratio range from 6.44 to 10.44 and two different values of aspect area ratio were taken (2.905, 3.27). Operating conditions include saturated generator vapor temperature from 78ºC to 95ºC, and evaporator temperature from 8 to 12ºC. Considering no superheating is applied, all of these temperatures correspond to the saturated vapor pressures. (a) (b) Fig. 4-1. (a) Critical entrainment ratio and (b) critical condensing pressure comparison of the present model with the experimental results of Huang et al. (1999) Although the maximum error of the critical entrainment ratio is 24.2%, only in this case the error is higher than 20%. Additionally, for the 39 cases of the study, practically 50% of them do not exceed an error of 8%. In the case of the critical condensing pressure, the maximum error is 12.8%, however, the error range of almost 90% points of back pressures are within 10%. 4.2 R245fa Table (4-1) shows the comparison between the present model and the data from the R245fa experimental study by Shestopalov et al. (2015). Furthermore, in this case, the COP and the 𝑄󰇗𝑒 are also included in the analysis. It can be seen that the theoretical results of all the parameters coincide fairly well with the measured 0.0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1.0 0.0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1.0 Present model ωcc Experimental ωcc -24% +8% 0.5 0.6 0.7 0.8 0.9 1.0 1.1 1.2 1.3 1.4 1.5 0.5 0.6 0.7 0.8 0.9 1.0 1.1 1.2 1.3 1.4 1.5 Present model Pcc (bar) Experimental Pcc (bar) -13% +4% T Ejector model validation 12 12 data. For critical entrainment ratios, the maximum absolute error is practically 5%. Most of the critical condensing pressures are within ±3% deviation compared with the experimental data and the maximum error is about 4%. All the estimated values of COP do not exceed a derivation of ±6% and the maximum cooling effect error, 𝑄󰇗𝑒, is 8.3% Tabla 4-1. Summary of ejector and ERS performance of theoretical results calculated with the present model, compared with the experimental data from Shestopalov et al. (2015) Critical condensing Pressure Pcc (bar) Critical entrainment ratio ωcc COP Qe (kW) Percentage error (theoretical compared with experimental) Ejector type Area ratio Tp (°C) Ts (°C) Exp. Present study Exp. Present study Exp. Present study Exp. Present study Pcc ωcc COP Qe 1A 7.25 90 8 2.098 2.044 0.241 0.233 0.186 0.175 3.1 2.8 -2.57 -3.34 -5.74 -8.30 1B 8.32 90 8 1.965 1.914 0.318 0.315 0.243 0.237 4.1 3.8 -2.61 -0.91 -2.41 -6.22 1C 9.63 90 8 1.793 1.784 0.402 0.416 0.309 0.313 5.2 5.1 -0.52 3.49 1.33 -2.37 2A 8.33 95 12 2.261 2.183 0.345 0.342 0.265 0.259 4.4 4.1 -3.45 -0.84 -2.46 -5.86 2B 9.55 95 12 2.062 2.049 0.423 0.441 0.323 0.333 5.4 5.3 -0.64 4.23 3.13 -1.15 2C 11.06 95 12 1.864 1.906 0.536 0.562 0.411 0.425 7 6.8 2.20 4.90 3.37 -2.75 3A 9.71 100 16 2.307 2.311 0.471 0.484 0.358 0.367 5.7 5.7 0.19 2.71 2.54 0.29 3B 11.14 100 16 2.163 2.157 0.575 0.604 0.440 0.458 7 7.1 -0.28 5.04 4.17 1.95 3C 12.89 100 16 1.944 2.019 0.744 0.752 0.570 0.570 9.2 8.9 3.83 1.03 0.08 -3.45 4.3 R245fa The experimental investigation carried out by Li et al. (2018) with R134a is also used to validate the present model, since this refrigerant is classified as wet fluid and is far from obeying ideal gas law. Furthermore, in this case, the validation is verified under both the double-chock and single shock conditions. For the critical entrainment ratios, 42% of the total of 24 cases compared have an error below 6%. The maximum discrepancy between theoretical and experimental entrainment ratios causes an error of -17%, as shown in Fig. (4-2). In the case of the critical condensing pressure, it can be observed an acceptable error range (+7%,-2.5%). 13 Modelado y análisis del Sistema de Refrigeración Solar con Eyector para aplicaciones de climatización (a) (b) Fig. 4-2.- . (a) Critical entrainment ratio and (b) critical condensing pressure comparison of the present model with the experimental results of Li et al. (2018) Fig. (4-3) shows the comparison results of the entrainment ratio between experimental and theoretical data for the entire operational range. For three different values of secondary pressure and two saturated temperatures of the generator, the performance of the present model is quite similar to the experimental results. The present model could predict the entrainment ratio and the pressure changes under subcritical mode by the influence of the backpressure for other working fluids. Fig. 4-3. Comparison of the ejector subcritical performance of the present model and the experimental data provides by Li et al. (2018) for different levels of secondary temperatures (Te) and primary temperatures (Tg) (a) Pp=23.66 bar; (b) Pp=26.35 bar. Both inlet conditions primary and secondary are saturated. 0.0 0.1 0.2 0.3 0.4 0.5 0.0 0.1 0.2 0.3 0.4 0.5 Present model ωcc Experimental ωcc -17% +14% 6.0 7.0 8.0 9.0 10.0 11.0 12.0 6.0 7.0 8.0 9.0 10.0 11.0 12.0 Present model Pcc (bar) Experimental Pcc (bar) -2.5% +7% (a) (b) 0.00 0.05 0.10 0.15 0.20 0.25 0.30 28 30 32 34 36 38 40 42 ω T c (°C) T g=75°C Ar =3.32 Experiment Present model Te=10°C Te=12.5°C Te=15°C 0.00 0.05 0.10 0.15 0.20 0.25 30 32 34 36 38 40 42 44 ω T c (°C) Te=10°C Te=12.5°C Te=15°C Experiment Present model T g=80°C Ar=3.32 Results and discussion 14 14 5 RESULTS AND DISCUSSION 5.1 Analysis of the ERS performance using R134a In this chapter, the present model is employed to analyze the ejector refrigeration system using R134a as the working fluid. The purpose of this section is to find out the key features of the SERS under the proposed air precooling working conditions to guide the study of the low GWP-ODP refrigerants in the most accurate way possible. Table (5-1) presents the different ejector geometries used to carry out the analysis of this chapter. Tabla 5-1. Ejector geomtries Ejector dt (mm) dp1(mm) dm (mm) Ar BB [Li et al.(2018)] 2.09 2.70 3.81 3.32 1_1 [Shestopalov et al. (2015)] 2.75 4.50 5.80 4.45 AC [Li et al.(2018)] 2.50 3.30 5.60 5.02 AA [Huang et al.(1999)] 2.64 4.50 6.70 6.44 EC [Huang et al.(1999)] 2.82 5.10 7.60 7.26 ED [Huang et al.(1999)] 2.82 5.10 8.10 8.25 EF [Huang et al.(1999)] 2.82 5.10 8.84 9.83 EH [Huang et al.(1999)] 2.82 5.10 9.20 10.64 3C [Shestopalov et al. (2015)] 3.90 6.41 14.01 12.9 4_2 [Shestopalov et al. (2015)] 3.25 6.50 12.50 14.79 As depicted in Fig. (5-1a), as the generator temperature increases the critical entrainment ratio reduces. At a given area ratio, while increasing the saturation inlet temperature, and consequently, the primary pressure as well, the primary flow rate increases. As a result, the hypothetical throat fills a larger section of the constant area mixing section, which means the secondary flow has less section available to cross. Therefore, if the primary pressure continues to rise it reaches a point the hypothetical throat totally blocks the pass of the secondary flow or even appears the back-flow effect. This threshold where the entrainment ratio is zero can be observed for the case Ar=3.32 and Tg=95°C in Fig. (5-1a). By applying similar reasoning, at a given primary inlet temperature, as the area ratio increases the constant section does too, causing an increase in the available cross-section and so an increase in the amount of the entrained secondary flow. A higher proportion of area does not necessarily mean a higher primary flow rate, but usually, a larger area ratio relates to a bigger nozzle throat diameter, which is directly proportional to the primary mass flow rate. In these cases in which the primary flow also increases, the increase in the secondary flow still being more important so the net effect is an increase of the entrainment ratio. Fig (5-1a) portrayed this trend in which as the area ratio increase, so does the critical entrainment ratio. The critical entrainments ratios presented in Fig (5-1a) are calculated forcing a condenser temperature lower enough that allows the chock conditions in both primary and secondary flows, which means the ejector operates at critical mode. However, in practice, since the condenser temperature is generally imposed by the ambient temperature, the calculation of the critical condensing temperature is crucial to ensure the correct 15 Modelado y análisis del Sistema de Refrigeración Solar con Eyector para aplicaciones de climatización operation of the ejector. As shown in Fig. (5-1b), for a fixed area ratio, the higher is the generator temperature the higher is the critical condensing temperature and, for a fixed generator temperature as the area ratio increases the critical condensing temperature drops. It can be noticed that these relations are opposite to those between Tg and Ar with the critical entrainment ratio. (a) (b) Fig. 5-1. Critical entrainment ratio and (b) critical condensing temperature as a function of generator temperature at different area ratios with R134a and Te=10°C. The effect of including systems that allow a variable air volume control is also portrayed in Fig. (5-2). As described above, the difference between outdoor and evaporating temperatures is set as constant. Nevertheless, for a certain size of the SERS, by increasing the flow of fresh air through the evaporator the fresh air stream is less cooled down, which means that the evaporator temperature increase since the pinch point remains approximately constant. As Fig (5-2) shows, both the entrainment ratio (and therefore the cooling effect does too) and the COP, such as the critical condensing temperature are grater for higher evaporator temperatures. Although the increase in the performance opposes the effect caused by the increased fresh air flow, the system probably balances out in a new situation, where a higher amount of airflow is cooled slightly less. (a) (b) Fig. 5-2. (a) Critical entrainment ratio and (b) critical condensing temperature as a function of condensing temperature at different evaporator temperatures with R134a and Te=90°C. 5.2 Selection of working fluids Figs. (5-3) and (5-4) show the critical entrainment ratio, the COP, and the critical condensing temperature of 0 0.5 1 1.5 2 2.5 70 80 90 100 110 ωcc Tg(°C) 3.32 4.45 5.02 6.44 7.26 8.25 9.83 10.64 12.90 14.79 Ar Te=10°C 15 25 35 45 55 70 80 90 100 110 Tcc Tg(°C) 3.32 4.45 5.02 6.44 7.26 8.25 9.83 10.64 12.90 14.79 Ar Te=10°C 0 0.05 0.1 0.15 0.2 0.25 0.3 25 35 45 ω Tc(°C) 0°C 2.5°C 5°C 7.5°C 10°C Te Tg=90°C Ar=5.02 0 0.05 0.1 0.15 0.2 0.25 0.3 25 35 45 COP Tc(°C) 0°C 2.5°C 5°C 7.5°C 10°C Te Tg=90°C Ar=5.02 Results and discussion 16 16 several refrigerants for the two borderline cases of the application range. The minimum outdoor temperature considered is 25°C as from this point the building spaces may require cooling, and the maximum temperature is set to the higher temperature recorded in Seville, i.e. 43 °C. As argued above, due to the characteristics of the system, both the evaporator and condenser temperature are assumed to depend on the outdoor temperature. The considered generator temperature is in the range of 75-120 °C because these are the common limits of solar applications. As shown in Figs. (5-3) and (5-4) R1234yf and R290 have the highest critical entrainment ratio and COP. However, the low backpressure of these fluids is not enough to operate at critical mode, even in the most favorable case, due to their critical temperature is lower than the minimum needed, as can be seen in Fig. (1-c). As commented above, by decreasing the area ratio a higher critical condensing temperature can be achieved. This fact must be taken into account but for these cases is not enough to reach the desire conditions. There are 3 refrigerants with low ODP and GWP that met the requirements of high critical condensing temperature and acceptable entrainment ratio and COP, namely, R600, R600a and R717. From Figs. (11b) and (12b), it can be seen that R717 has better performance of those fluids, while R600 can achieve higher critical condensing temperatures. Nevertheless, R600 and R717 have practical risks since they are inflammable and toxic, respectively. Although R1234ze(E) is slightly inflammable, it is also included in the study as it would be more appropriate than the R717 and R600 to operate in an ATU, taking into account safety. 17 Modelado y análisis del Sistema de Refrigeración Solar con Eyector para aplicaciones de climatización Fig. 5-3. (a) Critical entrainment ratio; (b) critical COP and (c) critical condensing temperature as a function of generator temperature using different working fluids with Ar=5.04 and Te=10°C. (a) (b) (c) 0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 70 80 90 100 110 120 130 ωcc Tg( ° C) R134a R141b R245fa R717 R600a R600 R290 R1234ze(E) R1243zf R1234yf Te=10°C A r=5.02 0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 70 80 90 100 110 120 130 COPcc Tg( ° C ) R134a R141b R245fa R717 R600a R600 R290 R1234ze(E) R1243zf R1234yf Te=10°C A r=5.02 20 25 30 35 40 45 50 55 70 80 90 100 110 120 130 Tcc Tg( ° C) R134a R141b R245fa R717 R600a R600 R290 R1234ze(E) R1243zf R1234yf Tc Te=10°C A r=5.02 Referencias 24 24 Li, F., Tian, Q., Wu, C., Wang, X., Lee, J.M., 2017b. Ejector performance prediction at critical and subcritical operational modes. Appl. Therm. Eng. 115, 444–454. https://doi.org/10.1016/j.applthermaleng.2016.12.116 Li, F., Li, R., Li, X., Tian, Q., 2018. Experimental investigation on a R134a ejector refrigeration system under overall modes. Appl. Therm. Eng. 137, 784–791. https://doi.org/10.1016/j.applthermaleng.2018.03.102 Lund H, Flatten T. 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Energy Rev. 22, 198–208. https://doi.org/10.1016/j.rser.2013.01.034 Pereira, P.R., Varga, S., Soares, J., Oliveira, A.C., Lopes, A.M., De Almeida, F.G., Carneiro, J.F., 2014. Experimental results with a variable geometry ejector using R600a as working fluid. Int. J. Refrig. 46, 77–85. https://doi.org/10.1016/j.ijrefrig.2014.06.016 Petrovic, M.M., Stevanovic, V.D.. Two-Component Two-Phase Critical Flow. FME Transactions (2016) 44, 109-114 - 109. Pianthong, K., Seehanam, W., Behnia, M., Sriveerakul, T., Aphornratana, S., 2007. Investigation and improvement of ejector refrigeration system using computational fluid dynamics technique. Energy Convers. Manag. 48, 2556–2564. https://doi.org/10.1016/j.enconman.2007.03.021 Pranesh, V., Velraj, R., Christopher, S., Kumaresan, V., 2019. A 50 year review of basic and applied research in compound parabolic concentrating solar thermal collector for domestic and industrial applications. Sol. 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Energy and economic analysis of a variable-geometry ejector in solar cooling systems for residential buildings. Journal of Energy Storage 27, 101061. https://doi.org/10.1016/j.est.2019.101061. Tian, Z., Perers, B., Furbo, S., Fan, J., 2017. Annual measured and simulated thermal performance analysis of a hybrid solar district heating plant with flat plate collectors and parabolic trough collectors in series. Appl. Energy 205, 417–427. https://doi.org/10.1016/j.apenergy.2017.07.139 Van Nguyen, V., Varga, S., Soares, J., Dvorak, V., Oliveira, A.C., 2020. Applying a variable geometry ejector in a solar ejector refrigeration system. Int. J. Refrig. 113, 187–195. https://doi.org/10.1016/j.ijrefrig.2020.01.018 Varga, S., Oliveira, A.C., Palmero-Marrero, A., Vrba, J., 2017. Preliminary experimental results with a solar driven ejector air conditioner in Portugal. Renew. Energy 109, 83–92. https://doi.org/10.1016/j.renene.2017.03.016 25 Modelado y análisis del Sistema de Refrigeración Solar con Eyector para aplicaciones de climatización Weiss, W., Spörk-Dür, M., 2019. Global Market Development and Trends in 2018. Sol. Heat Worldw. Rep. 1, 86. Zhang T, Mohamed S. Conceptual design and analysis of hydrocarbon-based solar thermal power and ejector cooling systems in hot climates. J Sol Energy Eng 2014:137. Índice de conceptos 26 26 ÍNDICE DE CONCEPTOS Acronyms a Speed of sound (m·s-1) A Area (m2) AHU Air Handling Unit AR Area ratio CAM Constant Area-Mixing COP Coefficient of Performance Cp Specific Heat at Constant Pressure ([kJ·kg-1·K-1) CPM Constant Pressure-Mixing D Diameter (m) DHW Domestic Hot Water EPBD Energy Performance Buildings Directive EU Europe Union GWP Global Warming Potential h Enthalpy (kJ·kg-1) HCFC Hydrochlorofluorocarbon HFC Hydrofluorocarbon HFO Hydrofluoroolefin HEM Homogeneous Equilibrium Model HRM Homogeneous Relaxation Model (HVAC) Heating, Ventilation, and Air-Conditioning i Inlet IEA International Energy Agency m󰇗 Mass flow rate (kg·s-1) NZEB Nearly Zero Energy Buildings NXP Nozzel exit position O Outlet ODP Ozone Depletion Potential P Pressure (bar) Q Thermal power (kW) s Entropy (kJ·kg-1·K-1) SERS Solar-driven Ejector Refrigeration System T Temperature (K) u Velocity (m·s-1) V󰇗 Volumetric flow rate (m3·s-1) 27 Modelado y análisis del Sistema de Refrigeración Solar con Eyector para aplicaciones de climatización VCC Vapor Compression Cycle W Electric power absorbed (kW) x Vapor quality Greek symbols β Thermal expansion coefficient (K-1) ϵ Void fraction η Efficiency ρ Density (kg·m-3) ϕ Efficiency due to frictional loss ω Entrainment ratio Subscripts as After shock wave b Breakdown c Condenser cc Condensation critical point d Diffuser e Evaporator el Electrical g Generator l Liquid m Mixed flow p Primary fluid p0 Primary flow at inlet of ejector p1 Primary nozzle exit r Ratio s Secondary fluid s0 Secondary flow at inlet of ejector s’ Isentropic process t Throat of the nozzle v Vapor phase VCC Vapor Compresion Cycle y Position of the hypothetical throat 3 Constant area sectio Índice de conceptos 28 28 29 Modelado y análisis del Sistema de Refrigeración Solar con Eyector para aplicaciones de climatización APÉNDICE A Table A1. Optimal values for the two multi-ejectors using R1234ze(E) Working conditions Optimal values Seasonal Performance Multiejector Tg (°C) Te (°C) Tc (°C) dt (mm) dp1(mm) Ar ωcc Tcc (°C) COPcc (COPel)cc COP COPel B 100 14 39 3.00 5.25 6.62 0.278 39.05 0.232 10.62 0.20 9.08 100 18 43 3.00 4.98 5.66 0.249 43 0.210 9.49 100 22 47 3.00 5.26 4.89 0.227 47.05 0.194 8.64 C 80 14 39 4.00 5.06 3.03 0.127 39 0.102 9.89 0.07 7.30 80 18 43 2.95 3.58 2.54 0.089 43.02 0.073 7.21 80 22 47 2.50 2.74 2.34 0.057 47.05 0.047 4.79 Table A2. Optimal values for the two multi-ejectors using R717 Working conditions Optimal values Seasonal Performance Multiejector Tg (°C) Te (°C) Tc (°C) dt (mm) dp1(mm) Ar ωcc Tcc (°C) COPcc (COPel)cc COP COPel A 120 14 39 3.75 7.13 8.97 0.362 39.05 0.41 19.10 0.37 17.07 120 18 43 4.00 7.61 7.77 0.344 43 0.39 18.10 120 22 47 2.95 5.78 6.77 0.328 47.05 0.37 16.99 B 100 14 39 4.00 7.79 5.63 0.281 39.05 0.29 23.68 0.25 20.55 100 18 43 4.00 7.79 4.83 0.255 43 0.27 21.76 100 22 47 3.00 5.84 4.15 0.228 47 0.24 19.77 C 80 14 39 4.00 6.87 3.03 0.135 39 0.13 20.45 0.10 15.70 80 18 43 3.02 4.28 2.54 0.099 43.05 0.10 15.76 80 22 47 3.65 3.65 2.00 0.048 48.25 0.04 8.32 Apéndice A 30 30 31 Modelado y análisis del Sistema de Refrigeración Solar con Eyector para aplicaciones de climatización APÉNDICE B: RESUMEN EN ESPAÑOL Proyecto Fin de Máster Sistemas de Energía Térmica RESUMEN EN ESPAÑOL Modelado y análisis del Sistema de Refrigeración Solar con Eyector para aplicaciones de climatización Autor: Miguel Ávila Gutiérrez Tutor: José Manuel Salmerón Lissén Tutor: Bernardo Peris Pérez Dpto. Ingeniería energética Escuela Técnica Superior de Ingeniería Universidad de Sevilla Sevilla, 2020 Apéndice B: resumen en español 32 32 1.-Introducción Los edificios son responsables de alrededor del 40% del consumo de energía de la UE y del 36% de las emisiones de gases de efecto invernadero (Comisión Europea, 2020). Esto se debe principalmente a que el 35% de los edificios de la UE tienen más de 50 años de antigüedad, y casi el 75% de los edificios se consideran ineficientes desde el punto de vista energético. Para impulsar la mejora del rendimiento energético de los edificios se estableció un marco legislativo, que incluye la Directiva sobre el rendimiento energético de los edificios (EPBD) (2010/31/UE, 2010) y la Directiva sobre la eficiencia energética (2012/27/UE, 2012). En particular, la DEEE establece que todos los edificios nuevos a partir de 2021 deben ser edificios de energía casi nula (NZEB), lo que significa que se debe exigir una cantidad muy baja de energía, y se debe considerar un grado significativo de fuentes de energía renovable en la generación. A este respecto, los países se han comprometido a reducir al mínimo su dependencia de la energía primaria mediante la inclusión de tecnologías para aprovechar las fuentes de energía renovable para la calefacción, la refrigeración y la producción de electricidad, además de otros fines (Kong y otros, 2020). Por ejemplo, España ha actualizado recientemente la legislación sobre eficiencia energética de los edificios (Real Decreto-Ley 732/2019, 2019), estipulando que más del 70% de la demanda de agua caliente sanitaria (ACS) de los nuevos edificios debe ser suministrada por energías renovables. Teniendo en cuenta que España se ha convertido en un país intensivo en energía solar, la mayor parte de la futura demanda de ACS de los NZEB procedería de un origen solar mediante el uso de colectores solares térmicos (Pablo-Romero et al., 2013). Comúnmente, los campos de energía solar térmica diseñados en un clima cálido para proporcionar agua caliente para una red de calefacción de distrito, o el agua caliente sanitaria para un edificio, están sobredimensionados durante los meses de verano (Tian et al., 2017). Entonces, los colectores solares requieren ser enfriados rechazando el calor de desecho al ambiente, mientras que se aumenta la demanda de aire acondicionado del edificio. Una alternativa eficiente es producir enfriamiento solar recuperando el calor desperdiciado (Pranesh et al., 2019). Hay que mencionar que se espera que la demanda de refrigeración en los edificios crezca hasta triplicarse para 2050 (AIE, 2018), lo que requiere que las medidas de eficiencia energética sean NZEB. De hecho, en 2019 se instalaron en todo el mundo unos 1800 sistemas de refrigeración solar que utilizan refrigeradores de absorción (Weiss y Spörk-Dür, 2019). Entre las tecnologías de refrigeración solar, el Sistema de Refrigeración por Eyectores Accionados por Energía Solar (SERS) es una tecnología emergente en comparación con los enfriadores de absorción maduros. Algunas ventajas del SERS son su simplicidad estructural, menores costes de compra y mantenimiento (Varga et al., 2017). Además, el SERS permite una selección flexible de los fluidos de trabajo, pudiendo diseñar de forma más respetuosa con el medio ambiente mediante el uso de refrigerantes naturales y fluorados con un bajo potencial de calentamiento global (GWP). Por otro lado, el mayor inconveniente de la SERS es el menor rendimiento alcanzable en comparación con los sistemas de absorción. No obstante, recientemente Braimakis (2021) sostuvo que, en determinadas circunstancias, especialmente en escalas bajas, el SERS puede ser competitivo en cuanto a costos si sus menores costos de equipo compensan los gastos de capital del campo solar más elevado que se requiere. Del mismo modo, numerosas investigaciones se centraron en las ventajas de la SERS para la refrigeración solar como futuro competidor de los sistemas de amoníaco. Roman y Hernández (2011) evaluaron el rendimiento de la SERS utilizando R290, R600, R600a, R152 y R134a. El análisis se llevó a cabo considerando temperaturas de evaporación entre 5 y 15 ºC, y temperaturas de generación entre 70 y 95ºC. Galindo y otros (2020) se centraron en el uso de los refrigerantes de bajo PCG R1234yf, R1234ze y R600a. El análisis se llevó a cabo asumiendo una temperatura de evaporación de 13ºC, una temperatura de condensación de 40ºC, y temperaturas de generación de hasta 125 ºC, identificando al R1234yf como el mejor candidato en términos de eficiencia global del sistema. Guo y Shen (2009) propusieron la adopción del SERS en un edificio de oficinas para el aire acondicionado. El refrigerante propuesto fue el R134a, para 8ºC de temperatura de evaporación, 65ºC de temperatura del generador, y condiciones de condensación de 27 ºC a 34ºC. De esta manera, los investigadores concluyeron que se puede alcanzar un ahorro de energía eléctrica del 75% utilizando el SERS. Pridasawas y Lundqvist (2004) evaluaron el SERS usando el R600 a 10 ºC de temperatura de evaporación, 37ºC de temperatura de condensación, y 90 ºC en el generador, alcanzando un COP de 2,5. Sin embargo, los autores señalaron que si la energía térmica se obtiene gratuitamente, como el calor residual rechazado, el rendimiento del sistema podría definirse como la relación entre la capacidad de refrigeración y el 33 Modelado y análisis del Sistema de Refrigeración Solar con Eyector para aplicaciones de climatización consumo eléctrico (COPel), alcanzando un valor de 32. Con respecto a la mejora del rendimiento del SERS, dos corrientes se perfilan en la literatura. Primero, el uso de fluidos de trabajo más eficientes y sostenibles. Y en segundo lugar, la mejora del rendimiento del eyector mediante múltiples módulos eyectores o dispositivos de geometría variable. Con respecto al uso de refrigerantes de bajo GWP, se han investigado numerosas alternativas en la literatura para reemplazar los fluidos de trabajo tradicionales. Fang y otros (2017) propusieron los refrigerantes de hidrofluroolefina (HFO) R1234yf y R1234ze(E) como un sustituto del R134a en el ciclo de refrigeración del eyector. Zhang y Mohamed (2015) demostraron que el R290, R600a, R601, R1234yf, R1234ze y R744 presentan ventajas inherentes para las aplicaciones de refrigeración para reemplazar al R134a. Mwesigye y Dworkin (2018) propusieron R600, R1234ze(Z), R1233zd(E), RE245fa2 y R1336mzz(Z) como reemplazos del R141b y R245fa comúnmente utilizados, demostrando que se pueden alcanzar mayores valores de relación de arrastre, capacidad de enfriamiento y COP a través de una selección adecuada de refrigerantes. En la Tabla 1 se recoge un examen no exhaustivo de las alternativas propuestas de bajo PCG, incluidos los fluidos de trabajo fluorados y naturales, a los refrigerantes tradicionales. Entre las propiedades termofísicas, la temperatura crítica se utiliza comúnmente para la selección del fluido de trabajo. Esta temperatura establece la máxima temperatura de evaporación en el generador cuando se implementan ciclos subcríticos, que son preferibles cuando no se adopta el R744 (Besagni et al., 2016). Además, la pendiente del fluido clasifica el refrigerante según la línea de vapor saturado como isoentrópico, seco y húmedo. Específicamente, la pendiente se ha esbozado como un parámetro crucial para la selección del modelo de eyector. Así, desde el primer modelo 1-D desarrollado por Keenan y otros (1950), los modelos ideales basados en el gas se utilizaron ampliamente para el estudio del aire, el vapor, pero también los refrigerantes de pendiente seca. Sin embargo, investigaciones más recientes demuestran que puede producirse un estado bifásico dentro de la boquilla del eyector, siendo aún más probable en fluidos húmedos, lo que requiere modelos reales de eyectores de fluidos para mejorar la precisión de los resultados (Chen et al., 2017) Table 1. Refrigerantes tradicionales y de bajo potencial de calentamiento global (fluorados y naturales) propuestos para trabajar con los sistemas SERS. No. Type Mol. Mass (kg/kmol) Boiling T (ºC) Critical T (ºC) Critical P (bar) ASHRAE 34 Fluid slope ODP GWP Traditional R134a HFC 102.03 -26.1 101.1 40.6 A1 wet 0 1430 R141b HCFC 116.95 32 204.4 42.1 A2 dry 0 725 R245fa HFC 134.05 15 153.9 36.5 B1 dry 0 1030 Fluorinated R1234yf HFO 114.04 -29.5 94.7 33.8 A2L isentropic 0 4 R1234ze(E) HFO 114.04 -19 109.4 36.3 A2L isentropic 0 6 R1243zf HFO 96.05 -25.4 103.8 35.2 A2L isentropic 0 <1 Natural R290 HC 44.10 -42.1 96.7 42.5 A3 wet 0 20 R600a HC 58.12 -11.7 134.7 36.3 A3 isentropic 0 4 R600 HC 58.12 -0.5 152 38 A3 isentropic 0 4 R717 NH3 17.03 -33.3 132.3 113.3 B2L wet 0 0 Como se ha mencionado, el rendimiento del SERS depende directamente de la eficiencia del eyector, que, a su vez, se ve fuertemente afectado por su geometría interna y las condiciones de funcionamiento de un refrigerante específico (Tashtoush y Nayfeh, 2020). A pesar de la simplicidad de utilizar un solo eyector fijo, Apéndice B: resumen en español 40 40 3.-Resultados 3.1.-Estudio previo del comportamiento del SERS usando R134a En este capítulo, el presente modelo se emplea para analizar el sistema de refrigeración del eyector utilizando R134a como fluido de trabajo. El propósito de esta sección es descubrir las características clave del SERS bajo las condiciones de trabajo de preenfriamiento de aire propuestas para guiar el estudio de los refrigerantes de bajo GWP-ODP de la manera más precisa posible. La tabla (3) presenta las diferentes geometrías de los eyectores utilizadas para llevar a cabo el análisis de este capítulo Table 3. Geometrías de los eyectores usados en el análisis previo Ejector dt (mm) dp1(mm) dm (mm) Ar BB [Li et al.(2018)] 2.09 2.70 3.81 3.32 1_1 [Shestopalov et al. (2015)] 2.75 4.50 5.80 4.45 AC [Li et al.(2018)] 2.50 3.30 5.60 5.02 AA [Huang et al.(1999)] 2.64 4.50 6.70 6.44 EC [Huang et al.(1999)] 2.82 5.10 7.60 7.26 ED [Huang et al.(1999)] 2.82 5.10 8.10 8.25 EF [Huang et al.(1999)] 2.82 5.10 8.84 9.83 EH [Huang et al.(1999)] 2.82 5.10 9.20 10.64 3C [Shestopalov et al. (2015)] 3.90 6.41 14.01 12.9 4_2 [Shestopalov et al. (2015)] 3.25 6.50 12.50 14.79 Como se muestra en la Fig. (7a), a medida que la temperatura del generador aumenta, la relación de arrastre crítica se reduce. En una relación de área determinada, al tiempo que aumenta la temperatura de entrada de saturación, y por consiguiente, también la presión primaria, aumenta la tasa de flujo primario. Como resultado, la hipotética garganta llena una sección mayor de la sección de mezcla de área constante, lo que significa que el flujo secundario tiene menos sección disponible para cruzar. Por lo tanto, si la presión primaria sigue aumentando, llega a un punto en que la garganta hipotética bloquea totalmente el paso del flujo secundario o incluso aparece el efecto de reflujo. Este umbral en el que el índice de arrastre es cero puede observarse para el caso Ar=3,32 y Tg=95°C en la Fig. (7a). Aplicando un razonamiento similar, a una temperatura de entrada primaria dada, a medida que la relación de área aumenta la sección constante también lo hace, causando un aumento de la sección transversal disponible y, por lo tanto, un aumento de la cantidad de flujo secundario arrastrado. Una mayor proporción de área no significa necesariamente una mayor tasa de flujo primario, pero, por lo general, una mayor relación de área se relaciona con un mayor diámetro de la garganta de la boquilla, que es directamente proporcional a la tasa de flujo másico primario. En estos casos en que el flujo primario también aumenta, el aumento del flujo secundario sigue siendo más importante, por lo que el efecto neto es un aumento del índice de arrastre. En la figura (7a) se representa esta tendencia en la que, a medida que aumenta la relación de superficie, también lo hace la relación de arrastre crítica. 41 Modelado y análisis del Sistema de Refrigeración Solar con Eyector para aplicaciones de climatización Los coeficientes de arrastre críticos presentados en la figura (7a) se calculan forzando una temperatura del condensador lo suficientemente baja como para permitir las condiciones de calce tanto en los flujos primarios como en los secundarios, lo que significa que el eyector funciona en modo crítico. Sin embargo, en la práctica, como la temperatura del condensador viene generalmente impuesta por la temperatura ambiente, el cálculo de la temperatura crítica de condensación es crucial para asegurar el funcionamiento correcto del expulsor. Como se muestra en la figura (7b), para una relación de área fija, cuanto más alta es la temperatura del generador, más alta es la temperatura crítica de condensación y, para una temperatura fija del generador, a medida que la relación de área aumenta, la temperatura crítica de condensación disminuye. Puede observarse que estas relaciones son opuestas a las que existen entre Tg y Ar con la relación de arrastre crítica. (a) (b) Fig. 7. (a) Ratio crítico de arrastre y (b) temperatura crítica de condensación en función de la temperatura de generación para diferentes geometrías, usando R134a y Te=10°C 3.2.-Selección de los fluidos de trabajo Las figuras (8) y (9) muestran la relación de arrastre crítica, el COP y la temperatura de condensación crítica de varios refrigerantes para los dos casos límite del rango de aplicación. La temperatura mínima exterior considerada es de 25°C, ya que a partir de este punto los espacios de los edificios pueden necesitar refrigeración, y la temperatura máxima se fija en la temperatura más alta registrada en Sevilla, es decir, 43 °C. Como se ha argumentado anteriormente, debido a las características del sistema, se supone que tanto la temperatura del evaporador como la del condensador dependen de la temperatura exterior. La temperatura considerada del generador se encuentra en el rango de 75-120 °C porque estos son los límites comunes de las aplicaciones solares. Como se muestra en las Figs. (8) y (9), el R1234yf y el R290 tienen el coeficiente de arrastre crítico y el COP más altos. Sin embargo, la baja contrapresión de estos fluidos no es suficiente para operar en modo crítico, incluso en el caso más favorable, debido a que su temperatura crítica es inferior a la mínima necesaria, como se puede ver en la Fig. (8-c). Como se ha comentado anteriormente, al disminuir la relación de área se puede lograr una mayor temperatura crítica de condensación. Este hecho debe tenerse en cuenta, pero para estos casos no es suficiente para alcanzar las condiciones deseadas. Hay 3 refrigerantes con bajo ODP y GWP que cumplen los requisitos de alta temperatura crítica de condensación y relación de arrastre y COP aceptables, a saber, R600, R600a y R717. En las Figs. (8b) y (9b), se puede ver que el R717 tiene un mejor rendimiento de esos fluidos, mientras que el R600 puede alcanzar temperaturas de condensación críticas más altas. No obstante, el R600 y el R717 tienen riesgos prácticos ya que son inflamables y tóxicos, respectivamente. Aunque el R1234ze(E) es ligeramente inflamable, también se incluye en el estudio ya que sería más apropiado que el R717 y el R600 operar en una UTA, teniendo en cuenta la seguridad. . 0 0.5 1 1.5 2 2.5 70 80 90 100 110 ωcc Tg(°C) 3.32 4.45 5.02 6.44 7.26 8.25 9.83 10.64 12.90 14.79 Ar Te=10°C 15 25 35 45 55 70 80 90 100 110 Tcc Tg(°C) 3.32 4.45 5.02 6.44 7.26 8.25 9.83 10.64 12.90 14.79 Ar Te=10°C Apéndice B: resumen en español 42 42 (a) (b) (c) Fig. 8. (a) Ratio crítico de arrastre, (b) critical COP y (c) temperatura crítica de condensación en funcion de la temperatura de generación usando diferentes refrigerantes. Caso más favorable. 0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 70 80 90 100 110 120 130 ωcc Tg( ° C) R134a R141b R245fa R717 R600a R600 R290 R1234ze(E) R1243zf R1234yf Te=10°C A r=5.02 0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 70 80 90 100 110 120 130 COPcc Tg( ° C ) R134a R141b R245fa R717 R600a R600 R290 R1234ze(E) R1243zf R1234yf Te=10°C A r=5.02 20 25 30 35 40 45 50 55 70 80 90 100 110 120 130 Tcc Tg( ° C) R134a R141b R245fa R717 R600a R600 R290 R1234ze(E) R1243zf R1234yf Tc Te=10°C A r=5.02 43 Modelado y análisis del Sistema de Refrigeración Solar con Eyector para aplicaciones de climatización (a) (b) (c) Fig. 9. (a) Ratio crítico de arrastre, (b) critical COP y (c) temperatura crítica de condensación en funcion de la temperatura de generación usando diferentes refrigerantes. Caso más desfavorable 0 0.2 0.4 0.6 0.8 1 1.2 1.4 1.6 1.8 70 80 90 100 110 120 130 ωcc Tg( ° C) R134a R141b R245fa R717 R600a R600 R290 R1234ze(E) R1243zf R1234yf T e=28°C A r=5.02 0 0.2 0.4 0.6 0.8 1 1.2 1.4 1.6 70 80 90 100 110 120 130 COPcc Tg( ° C) R134a R141b R245fa R717 R600a R600 R290 R1234ze(E) R1243zf R1234yf T e=28°C A r=5.02 30 35 40 45 50 55 60 65 70 80 90 100 110 120 130 Tcc Tg( ° C) R134a R141b R245fa R717 R600a R600 R290 R1234ze(E) R1243zf R1234yf Tc Te=28°C A r=5.02 Apéndice B: resumen en español 44 44 3.3.-Funcionamiento del multi-eyector La figura (10a) muestra el COP del eyector individual y el COP del multieyector. A una temperatura dada del generador y para un solo eyector, el COP aumenta a medida que la temperatura ambiente aumenta, alcanzando el máximo en el punto de diseño. Hasta ahora el expulsor ha funcionado en modo crítico, pero como la temperatura de condensación supera la temperatura de condensación crítica, el expulsor empieza a funcionar en modo subcrítico a partir de este punto, con la consiguiente caída del COP. Se puede observar que la relación de área óptima es menor a medida que aumenta la temperatura ambiente, esto se debe a que la temperatura de condensación también aumenta, y como se mencionó anteriormente, la disminución de la relación de área es una forma de obtener mayores contrapresiones. La otra consecuencia de la reducción del valor del Ar es la caída de la relación de arrastre y por lo tanto el COP también lo hace. Sin embargo, en este caso, la disminución es suave debido a que la temperatura del evaporador también aumenta con la temperatura exterior, lo que significa que el flujo secundario es arrastrado más fácilmente como resultado del aumento de la presión del evaporador. La Fig. (10b) presenta que el COPel sigue la misma tendencia que el COP. (a) (b) Fig. 10. Funcionamiento en función de las temperaturas de un clima cálido para cada eyector optimizado usando R600: (a) COP (b) COPel La Fig. (11) muestra el rendimiento de los tres multiejectores diferentes diseñados para tres niveles de temperatura del generador. Independientemente de la temperatura del generador, cada multieyector debe alcanzar la misma temperatura de condensación crítica, ya que ésta depende de la temperatura ambiente. Por un lado, en un Ar dado, cuanto más alta es la temperatura del generador, más alta es la temperatura crítica de condensación, pero más baja es la del COP. Por otro lado, cuanto más alta es la temperatura del generador, más alto es el Ar que puede alcanzar esa temperatura de condensación crítica necesaria. (a) (b) Fig. 11. Funcionamiento en función de las temperaturas de un clima cálido para cada multi-eyector 0.00 0.05 0.10 0.15 0.20 0.25 0.30 0.35 0.40 0 50 100 150 200 250 300 350 25 26 27 28 29 30 31 32 33 34 35 36 37 38 39 40 41 42 43 COP Numbers of hours (h) Toutdoor (°C) Warm climate (Sevilla/Spain) COP (Ar=11.01) COP (Ar=9.56) COP (Ar=8.3) COP Multiejector Tg=120 ° CTg=120 ° C 0 5 10 15 20 25 30 35 0 50 100 150 200 250 300 350 25 26 27 28 29 30 31 32 33 34 35 36 37 38 39 40 41 42 43 COPel Numbers of hours (h) Toutdoor (°C) Warm climate (Sevilla/Spain) COPel (Ar=11.01) COPel (Ar=9.56) COPel (Ar=8.3) COPel Multiejector Tg=120 ° CTg=120 ° C 0.00 0.05 0.10 0.15 0.20 0.25 0.30 0.35 0.40 0.45 0.50 0 50 100 150 200 250 300 350 25 26 27 28 29 30 31 32 33 34 35 36 37 38 39 40 41 42 43 COP Numbers of hours (h) Toutdoor (°C) Warm climate (Sevilla/Spain) COP Multiejector Tg=120°C COP Multiejector Tg=100°C COP Multiejector Tg=80°C 0 5 10 15 20 25 30 35 40 0 50 100 150 200 250 300 350 25 26 27 28 29 30 31 32 33 34 35 36 37 38 39 40 41 42 43 COPel Numbers of hours (h) Toutdoor (°C) Warm climate (Sevilla/Spain) COPel Multiejector Tg=120°C COPel Multiejector Tg=100°C COPel Multiejector Tg=80°C 45 Modelado y análisis del Sistema de Refrigeración Solar con Eyector para aplicaciones de climatización para diferentes temperaturas de generación, usando R600. 3.4.-Análisis estacional Se supone que la energía térmica suministrada al generador está disponible debido al contexto en el que se propone la aplicación de preenfriamiento. Por esa razón, el SERS se compara con un VCC estándar en términos de COP y COPel. Además, esta comparación se ha llevado a cabo teniendo en cuenta las condiciones reales de funcionamiento durante una temporada de calor en Sevilla utilizando diferentes fluidos de trabajo, a saber, un análisis estacional. Se ha tomado el mismo valor para la eficiencia isoentrópica de la bomba (SERS) y del compresor (VCC ) y es igual a 0,7. De manera similar, las pérdidas electromecánicas de ambos equipos se han asumido como equivalentes ( η_em=0,9). Para cada temperatura ambiente, el COP de la SVC se ha calculado bajo condiciones idénticas a las de la SERS, resultando un COP estacional igual a 6,5. Debe notarse que el COP del SVC es constante para todos los casos, ya que la temperatura del generador no influye en su rendimiento. La Fig. (12a) muestra el COP estacional de R717, R600 y R1234ze(E) para diferentes temperaturas del generador. Se puede ver que R717 logra los mejores rendimientos a niveles más altos de temperatura del generador con un COP igual a 0,25 (Tg=100°C) y 0,37 (Tg=120°C), seguido de cerca por R600 especialmente para el caso en que Tg=100°C (COP=0,24). Sin embargo, el R600 se posiciona como el fluido más adecuado en los casos de bajas temperaturas de generación, alcanzando un COP igual a 0,11. En comparación con el R717 y el R600, el R1234ze(E) tiene un rendimiento ligeramente inferior, pero alcanza un COP aceptable de 0,2 a Tg=100°C. Para temperaturas de generador más altas, el R1234ze(E) no ha sido analizado ya que su temperatura crítica es de unos 109°C. En la Fig. (12b) se comparan el COPel de la SERS y el COP del VCC para las diferentes temperaturas del generador. R600 muestra el mejor rendimiento para la temperatura más alta del generador considerada (Tg=120°C), alcanzando un COPel estacional de 28,3, que comparado con el COP del CCV, resulta en un aumento del 335%. Por el contrario, el amoníaco resulta ser más eficiente desde el punto de vista eléctrico a temperaturas de generador más bajas que el R600 y el R1234ze(E), especialmente en el caso de Tg=80°C, con un aumento del COPel de alrededor del 140% en comparación con el COP del CCV. Se ha encontrado que el R717 alcanza un máximo a Tg=100°C con respecto al COPel y es igual a 20,55, a diferencia del R600 y el R1234ze(E), cuyo COPel aumenta con la temperatura del generador. La mejora eléctrica del R1234ze(E) es menos pronunciada en comparación con los otros refrigerantes estudiados, pero aún así logra una mejora de casi el 40% para Tg=100 con respecto al VCC (a) (b) Fig. 12. (a) COP estacional y (b) COPel estacional en función de la temperatura de generación usando diferentes refrigerantes. . 0 0.05 0.1 0.15 0.2 0.25 0.3 0.35 0.4 80 100 120 Seasonal COP Tg(°C) Warm climate (Sevilla/Spain) R600 R717 R1234ze(E) 0 50 100 150 200 250 300 350 400 80 100 120 Seasonal COPel improvement (%) Tg(°C) Warm climate (Sevilla/Spain) R600 R717 R1234ze(E) Apéndice B: resumen en español 46 46 4.-Conclusiones y trabajo futuro En este estudio, se presenta un novedoso enfoque del SERS que funciona como un sistema de preenfriamiento de aire fresco para apoyar a la UTA. El modelo matemático del eyector ha sido validado con fluidos secos y húmedos. Se ha realizado un screening preliminar con varios fluidos de trabajo para seleccionar los más adecuados para operar en un clima cálido, debido a los requerimientos que esto conlleva para un SERS en cuanto a conseguir altas temperaturas de condensación. En cuanto a la estrategia de optimización, se diseña un multieyector de tres módulos para cada refrigerante seleccionado y para cada temperatura del generador utilizando EES. El rendimiento del SERS se calcula durante toda la temporada de calor en un clima cálido con fluidos de trabajo respetuosos con el medio ambiente, incluyendo R717, R600 y R1234ze(E) como posibles sustitutos del R134a comúnmente utilizado. La principal conclusión de este estudio está presente como sigue: En primer lugar, el COPel estacional del sistema que utiliza el R717 y el R600 con una temperatura del generador de 100°C fue superior al del R1234ze(E) en un 122% y 126%, respectivamente. En segundo lugar, se encontró que el amoníaco tiene el mejor COP estacional igual a 0,37 para Tg=120°C, seguido de cerca por el R600 con un COP estacional=0,33. Por último, el R600 mostró el mayor potencial de ahorro de electricidad al lograr un COPel estacional máximo igual a 28,3 para la mayor temperatura del generador considerada (Tg=120°C). La mejora es más del triple en comparación con un COP estacional de un CCV que funciona en condiciones idénticas. El documento demostró que el SERS trabajando como preenfriamiento de aire fresco usando R600, R717 e incluso R1234ze (E) podría ser una solución prometedora ya que los nuevos edificios están destinados a desperdiciar energía térmica solar durante las estaciones cálidas debido a la normativa de la NZEB. Aún se necesita hacer más análisis. Por un lado, se requieren varias simulaciones dinámicas de estaciones cálidas con diferentes tipos de edificios y, por lo tanto, diferentes cargas de la curva de enfriamiento, incluyendo los recursos solares y los aspectos económicos. Por otro lado, deben probarse investigaciones experimentales que incluyan alternativas factibles del SERS como la presentada en esta investigación.