Numerically Optimized Ejector Geometry for Ejector Refrigeration Systems With Low-Global Warming Potential Working Fluids

The waste heat from automobile exhausts, industrial processes, and air-conditioning facilities is tending toward global warming and ozone layer depletion. For example, a typical vapor compression refrigeration system (VCRS) used to provide air-conditioning in automobiles increases fuel consumption by up to 12–17%. In addition, the burning of fossil fuels in automobiles emits exhaust gases that cause long-term damage to the environment. In addition, the energy wasted through exhaust gases, coolant, and engine compartments of any automobile tolls up to 65% of total consumption [1,2]. Therefore, the utilization of this waste heat to operate a cooling system is an encouraging way of energy management with reduced fuel consumption. There are a few promising technologies that provide a cooling effect by recovering the process heat, such as absorption, adsorption, desiccant, and ejector refrigeration systems (ERS). Among these, the ERS is drawing more attention in recent days due to its simple and compact structure with cooling capacity from a few Watts to a kilowatts range, economic viability, and ease of maintenance, in spite of a lower coefficient of performance (COP) in the range of 0.1–0.4. However, the adsorption, absorption, and desiccant systems are not suitable for small-range applications like automotive refrigeration [3,4]. Therefore, further investigation of an ERS-based cooling system is needed to understand the complex flow dynamics across the ejector toward superior cooling capacity and COP.

To improve the COP of an ERS, several investigations are performed by varying the operating parameters, design specifications, and refrigerant types [5]. These studies indicate that condenser, evaporator, and generator temperatures are the key attributes significantly influencing primary and secondary flowrates within the system. In parallel, area ratio (AR), primary nozzle exit position (NXP), length and diameter of constant-area mixing section, convergence angle of constant-pressure mixing section, and divergence angle of the diffuser are other important geometric parameters that affect the overall ERS performance [6,7]. Efficient entrainment of secondary flow into the ejector necessitates a minimum primary flow pressure and optimal nozzle Mach number and ejector area ratio [8]. Although the ERS performance is predominantly affected by AR, the NXP also significantly impacts ejector behavior [9]. Commonly, the analytical model of an ejector is configured based on two approaches: constant-pressure mixing (CPM) and constant-area mixing (CAM), depending on the location of the NXP at the suction chamber or the constant-area section, respectively. However, the CPM superimposes the CAM modeling approach due to its utmost performance, ease of solution, and accuracy [10]. According to the ESDU design guide, the optimum NXP is 0.5–1.0 times the mixing chamber diameter, and the change in NXP can give a 100% rise in COP and cooling capacity [11]. Practically, it is difficult to specify a single optimum value for NXP due to complex flow structures and fast dynamics of working parameters [12,13]. For instance, the optimal NXP for refrigerant R141b is observed to be 1.5 times and 1.7–3 times the mixing chamber diameter in two separate investigations [14,15].

Most of the aforementioned studies are based on the 1D-CPM analogy. According to this model, at a specific distance from the NXP, the secondary flow becomes choked and begins to merge with the primary flow at constant pressure [16,17]. However, experiments confirm that the mixing phenomena inside the ejector are not simplistic [18,19]. Irreversible losses caused by an oblique shock wave from the primary nozzle are not considered by 1D models, demonstrating that the 1D-CPM theory with a single shock wave in the mixing section cannot be applied satisfactorily to all physical conditions [20]. Therefore, knowledge of the inner dynamics of this complex flow phenomenon is required to understand the behaviour working fluids, geometrical parameters, and working conditions in the mixing process. The computational fluid dynamics (CFD) technique is one such tool for visually evaluating the ejector flow field. A CFD simulation study employing six turbulence models and air as the working fluid for an ejector demonstrates excellent agreement with experimental results [21]. As air is not suitable for cooling applications, later studies focused on hydrochlorofluorocarbon (HCFC) and hydrofluorocarbon (HFC) refrigerants. CFD investigations of ERS using steam [22] and refrigerants R141b [23], R142b [20], and R1234yf [9] show that the secondary flow saturation temperature, AR, and NXP enhance ER, critical backpressure, and overall performance. It reports that the suction chamber geometry and secondary flow entry velocity do not considerably affect the flow pattern of the ejector. Therefore, the 3D and axisymmetric CFD models give almost similar results for CAM and CPM steam ejectors [24]. The CFD investigations on different ejector geometries at various operating circumstances using steam [6] and refrigerant R134a [25] reveal that the primary nozzle geometry and shock circle location can substantially affect the ejector's performance. Besides, these studies disclose that the ER is more sensitive to the throat diameter, surface roughness, diverging part length, and NXP of the primary nozzle.

HFC-134a is the utmost popular refrigerant used in refrigeration and AC applications. However, the leakage of this refrigerant can directly damage the environment owing to its higher global warming potential (GWP) value. The Montreal Protocol's Kigali amendment restricts the usage of R134a refrigerants and encourages the refrigeration industries to adopt zero-ozone depletion potential (ODP) and low-GWP refrigerants [26]. In recent years, a few new refrigerants have been introduced, such as R152a, R290 [27,28], R717, R600a [29,30], R1234yf, R1234ze(E) [31,32], R1233zdE [33], R515a, R456a, and R516a [34]. Recent studies also explored nano refrigerants, which are homogeneous mixtures of R134a, R152a, and R290 with TiO₂ and Al₂O₃ nanoparticles (0–5 wt%) [35]. While nanoparticle concentration enhances performance, it may also adversely affect ejector longevity due to intricate flow and shock wave phenomena inside the mixing chamber. Among all, the refrigerants R152a, R600a, and R290 are highly flammable, while the performance of R515a, R456a, and R1234ze(E) is very poor compared to R134a. Therefore, R1234yf and R1233zdE have garnered a lot of interest as potential replacements for the most widely used R134a refrigerant [36]. However, the use of R134a alternatives in ejectors and the impact of ejector geometry on these alternatives and their flow dynamics have not been extensively studied. Hence, the present investigation explores the qualitative performance of recently developed sustainable refrigerants and conventional refrigerants.

The preceding literature on ERS predominantly focuses on the impact of operational and geometrical factors on their performance, utilizing conventional refrigerants like R134a. However, a comprehensive understanding of ejector performance under the influence of a wide range of eco-friendly refrigerants and their intricate flow dynamics within the ejector is yet to be extensively examined. Furthermore, optimizing ejector geometry and operating conditions specifically for these environmentally friendly refrigerants remains largely unexamined. This lack of understanding regarding the optimal utilization of environmentally friendly refrigerants within ERS hinders the development of sustainable and efficient cooling solutions that align with global environmental goals. In view of this, the current work aims to examine the influence of operational and geometrical parameters on an ejector with five eco-friendly R134a alternate refrigerants, namely R1234yf, R1243zf, R152a, R513a, and R440a, having almost similar thermo-physical properties of R134a by using ansys fluent 20.0 software. The CFD model of the ejector is developed and validated with experimental data [37]. Later, investigations are conducted to compare the ER for all the chosen refrigerants under various operating conditions, such as condenser, evaporator, and generator temperatures, by adopting the specific ejector geometry from the experimental study [38]. The impact of geometrical factors, specifically AR and NXP, on the ER is thoroughly examined by evaluating seven different ARs and five distinct NXPs for each of the selected refrigerants. Further, the study is extended to find the optimal NXP in an extensive range of operating circumstances. Ultimately, an in-depth examination of the complex flow dynamics inside the ejector is carried out.

The compressor of a traditional VCRS is replaced with a superheated ejector for any ERS (Fig. 1). The typical ejector has four major parts: the primary nozzle, the suction chamber, the constant-pressure and constant-area mixing chambers, and the diffuser. First, the primary flow from the generator with higher pressure and temperature enters and expands in the primary nozzle, creates suction, and induces secondary flow from the evaporator into the suction chamber. Next, both streams combine in the mixing chamber and undergo a normal shock wave. Finally, this mixed-stream reaches condenser pressure by recovering all the kinetic energy in the diffuser. Subsequently, the mixed stream dissipates its heat at the condenser, splitting into the primary and secondary flows. The secondary flow expands in the expansion valve, and offers cooling by absorbing heat, $Qe$⁠, in the evaporator, while the primary flow is pumped back to the generator and absorbs heat, $Qg$⁠.

As the change in enthalpy across the generator and evaporator are almost equal, the COP of the ERS is solely dependent on the ER [39].

The other geometrical parameters required for modeling the ejector are depicted in Fig. 1. Among all, the primary nozzle exit diameter $(Dne)$ (defines the primary flow's driving force), constant-pressure mixing chamber entrance diameter $(Dpm)$ (at this position, secondary flow start entrains into the primary flow), and diffuser diameter $(Dd)$ are also few critical parameters that have equal importance in modeling of the ejector [40,41]. Table 1 shows the ejector dimensions used in the current dynamics model [38].

Refrigerant leaks can create various environmental issues like ozone depletion and global warming. The Kigali Amendment to the Montreal Protocol emphasizes the need to find environmentally friendly alternatives for high GWP refrigerants such as R134a. Therefore, finding a suitable replacement refrigerant for ERS is critical [16,31]. In view of this, five zero-ODP and low-GWP refrigerants are examined as alternatives for R134a, and their thermo-physical properties are reported in Table 2.

The dense meshes are prearranged in the primary nozzle exit and mixing section regions to handle the high incline properties developed due to the shock waves. The axisymmetric 2D model is used for ejector design since the performance and wall pressure distribution in the axisymmetric and 3D models are very similar (Fig. 2) [24]. Due to less running time requirements and better prediction of the separation and recirculation of the jet flows, the realizable k–ɛ with scalable wall function turbulence model is chosen [24,42,43]. The SIMPLE algorithm is employed for pressure velocity coupling to solve the discretized equations. The second-order upwind is used to discretize the convective and diffusive terms for high accuracy [44]. The detailed simulation algorithm is described in Fig. 3. In the upcoming section, the ejector behavior is evaluated under various boundary conditions, with seven refrigerants and different geometrical parameters, such as AR and NXP.

Table 3 demonstrates the boundary conditions adopted for the present analytical study. The saturation pressure corresponds to the evaporator, generator, and condenser temperatures $(Tevap,Tgen,andTcond)$ are considered as the pressures of the upstream (secondary and primary flows) and the downstream (mixed flow) flows. Variable AR is obtained by changing the $(Dt)$ value in the designed ejector (Table 4). Convergence criteria of 10⁻⁶ are used for accuracy and good convergence. The mass flowrate variation with the iterations plot is used to find the minimal number of iterations necessary to reach convergence. The ER values are calculated by observing the changes in the $m˙s$ and $m˙p$⁠. The physical parameters such as molecular weight, viscosity, thermal conductivity, and specific heat of the refrigerants at the boundary conditions are taken from the refprop 10.0 database.

Equation (9) is used to calculate the relative error percentage in ERs for various grids mentioned in Table 5. Beyond grid number 22,335, the relative error in ER is less than 1%, and also, the axial static pressure distribution along the ejector length is comparatively close (Fig. 4). Hence, a quadrilateral mesh of size 22,335 is used for further investigations.

The current CFD model is validated using experimental findings of an R134a-based ejector [37]. Table 6 signifies the experimental working conditions adopted for verifying the numerical observations of the present model. ER variation with NXP is compared for two selective working conditions (Fig. 5). For both conditions, ER increases with an increase in NXP, showing good agreement with experimental findings. However, the disparity between the experimental and numerical outcomes is noticed to be within ±10% and ±5% for conditions 1 and 2, respectively. This minor deviation is mostly because of (a) the wall roughness of the experimental ejector and (b) the consideration of the working fluid as an ideal gas [45].

In the ERS, the ejector is an essential component that regulates refrigerant flow across the evaporator and condenser. Therefore, a detailed performance study is conducted on the ejector for a wide variety of environment-friendly alternates of R134a refrigerant such as R1234yf, R1243zf, R152a, R513a, R440a, at a wide-ranging $Tcond,Tevap$⁠, and $Tgen$⁠. Furthermore, the study is extended to optimize the critical geometric parameters such as AR and NXP.

The ejector geometry with an AR of four is adopted for investigating the influence of operating circumstances such as the $Tcond,Tevap$⁠, and $Tgen$ on the ER from Wang et al. [38].

In line with topical climatic conditions, the $Tcond$ is altered from 32 °C to 47 °C for investigating the critical point temperature corresponding to critical back pressure for all the selected refrigerants by keeping the $Tgen$ and $Tevap$ constant at 90 °C and 5 °C, respectively. For the refrigerants R1234yf, R1234zf, R152a, R513a, R134a, and R440a, the critical temperature is observed in the order of 36 °C, 36 °C, 38 °C, 38 °C, 42 °C, and 42 °C, respectively. The ER is maximum and constant for all the refrigerants below their respective critical temperatures. In this region, both primary and secondary flows are chocked, and they do not depend on $Tcond$⁠. With the continuous increment of $Tcond$⁠, only the primary flow is choked, whereas the secondary flow still depends on $Tcond$⁠. Therefore, the ER is certainly decreased for all the refrigerants after their critical temperatures (Fig. 6). It is noticed that R1234yf and R1243zf give maximum ER for $Tcond$ of 37 °C and followed by it drops. However for other refrigerants like R134a and R440a, a maximum ER is observed beyond $Tcond$ of 40 °C. In between $Tcond$ of 37 °C and 40 °C, the maximum ER is attained with refrigerants R152a and R513a.

Focusing on the automobile and industrial AC applications, the $Tevap$ is varied from 0 °C to 14 °C to examine the ER at a fixed $Tcond$ and $Tgen$ of 32 °C and 90 °C, respectively. For a fixed $Tgen$⁠, the primary flowrate, $m˙p$⁠, remains constant. As the $Tevap$ rises, the differential pressure between the secondary flow inlet and suction chamber increases, allowing more secondary flow, $m˙s$⁠, into the ejector suction chamber. Therefore, the ER rises with the $Tevap$ for all the refrigerants (Fig. 7). R1234yf and R1243zf refrigerants achieve the highest ER of 0.45 at 14 °C $Tevap$⁠, while R134a has the lowest ER of 0.156 at 0 °C. Increasing $Tevap$ from 0 °C to 14 °C raises the ER by approximately 50% for R1234yf, R1243zf, R152a, and R513a refrigerants while nearly doubling it for R134a and R440a.

It is recommended to design the condenser in such a way that the $Tcond$ should be above the ambient temperature in order to transfer heat through free convection. Also, the $Tevap$ should not be too high or too low because it influences the power consumption. Therefore, a case study with the $Tevap$ and $Tcond$ of 5 °C and 32 °C, respectively, is conducted to identify the optimal $Tgen$ for each refrigerant. At these operating conditions, the optimal $Tgen$ for each refrigerant is identified to be in the range of 81–87 °C (Fig. 8). Optimal $Tgen$ and corresponding maximum ER are identified for various refrigerants: R1234yf and R1243zf at 84 °C with 0.378 and 0.387, respectively, R440a at 84 °C with 0.344, R134a and R152a at 81 °C with 0.339 and 0.355, and R513a at 87 °C with 0.346.

Figure 9 represents the contours of pressure and velocity for all refrigerants at $Tcond$⁠, $Tevap$⁠, and $Tgen$ of 32 °C, 5 °C, and 90 °C, respectively. For all, an array of expansion waves emerges from the primary nozzle exit and develops a converging duct to entrain the induced flow. Normal shock waves are formed at different regions for different refrigerants. In the case of R1234yf and R1243zf, the shock wave is almost disappeared, and the supersonic expansion wave is extended into the constant-area section exit with high velocity. The over-expansion of these expansion waves creates a flow field that pre-compresses the primary flow before mixing. The entrained induced flow combines with the primary flow and reduces its velocity. The combined flow experiences a weak normal shockwave in the diffuser and reaches the backpressure. This phenomenon can be the reason for the lesser critical back pressures of R1234yf and R1243zf than the other refrigerants, as shown in Fig. 6. For refrigerants R152a and R513a, an oblique shockwave is formed at the diffuser's inlet and propagates into the diffuser. In R134a and R440a, a convex-shaped shockwave is formed at the diffuser's inlet, and the velocity instantly falls to the subsonic levels.

The ejector's throat diameter,$Dt$⁠, is altered from 3.2 mm to 2.6 mm (with 1 mm step reduction) to examine the influence of AR. The impact of the NXP on the ER is also investigated. In the following investigations, the $Tcond$⁠, $Tevap$⁠, and $Tgen$ are fixed at 32 °C, 5 °C, and 90 °C, respectively, to run the ejector in the critical mode for all refrigerants.

The effective flow area for the primary flow reduces with the reduction in throat diameter. As a result, more suction is created, and more induced flow is entrained. Thus the increase in AR increases the ER by decreasing the $m˙p$ and increasing the $m˙s$ (Fig. 10). For all ARs, R1234yf and R1243zf achieve higher and almost equal ER values. At an AR of 5.325, the maximum ER of 0.51 is attained with R1243zf and R1234yf. Conversely, R440a refrigerant yields the lowest ER of 0.13 at an AR of 3.516. As the AR increases from 3.52 to 5.325, around 82% increment in the ER is achieved with R1234yf, R1243zf, and R152A refrigerants, while it is raised 2.3 times with R134a, R513a, and R440a.

At a given AR of four, as the NXP changes from 1 mm to 20 mm, the ER of all the refrigerants rises initially and is followed by reduction (Fig. 11). As the NXP moves away from the mixing section, the secondary flow's chocking position shifts toward upstream. Consequently, the hypothetical converging duct area of the secondary flow increases and allows more amount of induced flow. Therefore, the ER rises for all the refrigerants initially. If the NXP moves further, the primary flow's momentum decreases, resulting in a lower ER. It is not possible to define a single optimum NXP for all the refrigerants because of their difference in thermo-physical properties. For R1234yf, R1243zf, R513a, and R134a, the highest ER is achieved at 10 mm NXP. R440a achieves its peak at 15 mm NXP, while R152a's ER continuously increases within the tested NXP range. The optimum NXP for all the refrigerants is found to be in the range of 10–15 mm. At all the NXPs, the ER values are maximum with R1234yf refrigerant, and the highest ER of 0.355 is achieved at 10 mm NXP. Therefore, R1234yf is considered to optimize the NXP for a wide range of operating conditions.

The effect of NXP on the ER for a wide variation of $Tcond$⁠, $Tevap$⁠, and $Tgen$ is shown in Fig. 12. For all NXPs, the ER upsurges with a rise in $Tevap$ and $Tgen$⁠, in contrast, it decreases with an increase in $Tcond$⁠. With the variation of NXP from 1 mm to 20 mm, the ER initially increases by 1.21 times and attains 0.35766 at NXP of 10 mm, and subsequently decreases by 0.91 times to the $Tcond$ of 35 °C. However, for $Tcond$ of 41 °C, a sharp increase of ER up to 7.06 times at NXP of 10 mm and decreases by 3.48 times to the peak value is observed. From this, it is obvious that the NXP influence on ejector performance is more significant at higher $Tcond$⁠. A drop of 57.32% in the ER value can be noticed with a rise in $Tcond$ from 35 °C to 41 °C at the optimum NXP of 10 mm.

For a relatively higher $Tevap$ of 6 °C and beyond, the ER decreases first and then increases to the peak value. However, at a lower $Tevap$ of $≤0∘C$⁠, the ER increases initially, followed by decreases. The maximum ERs of 0.412, 0.372, and 0.330 for the $Tevap$ of 9 °C, 6 °C, and 3 °C are achieved at 10 mm NXP. For $Tevap$ of 0 °C and −3 °C, the peak ERs of 0.281 and 0.304 are obtained at 15 mm NXP. For every variation of $Tevap$ around 3 °C, a minimum of 13% rise in ER is obtained at all NXPs.

The maximum ER at optimum NXP of 10 mm increases by 36.74% when the $Tgen$ changes from 75 °C to 85 °C. At lower $Tgen$⁠, the ER decreases significantly beyond the optimum value of NXP, which indicates that the impact of NXP is higher at lower $Tgen$⁠. From this analysis, it can be comprehended that increasing NXP in a specific range can enhance the ER irrespective of the operating conditions. Moreover, the highest ER values are attained at 10 mm NXP. Hence, the optimum NXP for the present investigation is considered to be 10 mm.

The pressure and velocity contours and axial flow distribution throughout the ejector length with refrigerant R1234yf for various NXPs at $Tgen$⁠, $Tevap$⁠, and $Tcond$ of 90 °C, 5 °C, and 32 °C, respectively, are described in Figs. 13 and 14. Diamond-shaped expansion waves of the primary jet are formed from the primary nozzle exit at all NXPs. In the case of 1 mm and 5 mm NXPs, the supersonic converging primary jet is extended to the CAM section exit with high velocity and replaces the normal shock wave. Consequently, the induced flow chokes while mixing with the primary flow and reduces its velocity. The combined stream decelerates to the subsonic speed in the diffuser section and reaches the backpressure. Additionally, for NXP of 5 mm, a minor shock wave is formed immediately after the CAM section exits. An increase in NXP reduces the over-expansion of the primary jet and develops a normal shockwave in the diffuser section. In the case of 10 mm, 15 mm, and 20 mm NXPs, the secondary flow gets choked at the mixing section entrance by the supersonic duct of the primary flow. For these, a convex-shaped normal shock wave is generated from the CAM section exit at a distance of 5 mm to 2 mm (Fig. 14). The changes in the mixing processes and the pressure differences of the two streams are responsible for the variation in ER with NXP. In addition, an increase in NXP than the optimum value results in an extensive loss of kinetic energy, which decreases the mass flowrate. It may be noted that the maximum suction-induced pressure differential poses the maximum ER at NXP of 10 mm.

Figures 15 and 16 illustrate the pressure, Mach number, and velocity contours of R1234yf refrigerant at $Tcond,Tgen$⁠, and $Tevap$ of 32 °C, 90 °C, and 5 °C, respectively. The high-pressure primary flow at 30.7 bar expands at the primary nozzle throat and attains the sonic velocity (Fig. 15). Due to the formation of oblique expansion shock waves in the primary jet core, the flow velocity at the primary nozzle exit is very high. This supersonic-primary jet acts as a wall and forms a convergent duct that creates suction of 1.75 bar in the mixing chamber to entrain the induced flow into the ejector (Fig. 15). The low-pressure (3.71 bar), low-velocity secondary flow is entrained through this effective area, gains momentum along with the primary flow and reaches a maximum velocity of around 637 m/s at a 5 mm distance from the entrance of constant-area section (Fig. 16). The over-expansion of this combined stream reduces its velocity and recompresses at the end portion of the pseudo shock region. At a 1.3 mm distance from the CAM section exit, a typical shock wave has formed and been transported into the diffuser, resulting in a sudden rise in pressure and a significant decrease in velocity. In the diffuser, all available kinetic energy is transformed into pressure energy, and the mixed flow reaches the condenser pressure of 8.23 bar with a velocity of 30.7 m/s.

The current study numerically evaluates the thermodynamic performance of an ejector for air-conditioning by comparing R134a with five zero-ODP and low-GWP alternatives, namely, R1234yf, R1243zf, R152a, R513a, and R440a under various operating conditions. The efficacy of the proposed model is validated against the experimental outcomes of a superheated ejector [37]. Besides this, the impact of design variables such as AR and NXP on the ER is also investigated using seven different ARs and five different NXPs for all refrigerants. The complex flow behavior inside the ejector is examined for all refrigerant types and NXPs, and optimal NXP is determined for various operating circumstances. Significant observations are as follows:

• Refrigerants R1234yf and R1243zf yield the highest ER values in the critical mode of operation. However, the critical back pressure of these refrigerants is lower than that of others.

• Increasing $Tevap$ improves ER. All selected refrigerants outperform R134a in ER, with R1234yf and R1243zf achieving the highest values.

• At the $Tcond$ and $Tevap$ of 32 °C and 5 °C, the optimal $Tgen$ is identified in the range of 81–87 °C, corresponding to each refrigerant.

• Increasing the AR from 3.52 to 5.325 increases the ER by about 82% with the refrigerants R1234yf, R1243zf, and R152a. However with refrigerants R134a, R513a, and R440a, the ER increases almost 2.3 times the initial value.

• A single optimum NXP cannot be determined for all the refrigerants because of their difference in thermo-physical properties. For most refrigerants tested, the highest ER is achieved at 10 mm NXP. R440a achieves its peak at 15 mm NXP, while R152a's ER continuously increases within the tested NXP range.

• At lower $Tgen$ or higher $Tcond$⁠, the impact of NXP is higher. Irrespective of the operating conditions for refrigerant R1234yf, the optimum NXP is obtained as 10 mm, which is 1.67 times the CAM chamber diameter.

From all the parametric investigations, among the tested refrigerants, R1234yf and R1243zf show excellent performance with the least ODP and GWP, and are proposed as suitable alternatives to R134a in ejector refrigeration systems for automotive or industrial applications.

The Department of Science and Technology (DST), New Delhi, under the Scheme for Young Scientists and Technologists (SYST) with DST Sanction Order No. SP/YO/565/2018, provides funding for the present work.

There are no conflicts of interest.

The datasets generated and supporting the findings of this article are obtainable from the corresponding author upon reasonable request.

e =

relative error (%)

h =

enthalpy (J/kg)

u =

velocity (m/s)

A =

area of cross section (m²)

D =

diameter (m)

E =

overall energy (J)

L =

length (m)

P =

pressure (kPa)

T =

temperature (°C)

W =

work (kW)

ṁ =

mass flowrate (kg/s)

 Q_e =

cooling capacity (kW)

Q_g =

generator heat (kW)

d =

diffuser section

p =

primary flow

s =

secondary flow

t =

primary nozzle throat

am =

constant-area mixing section

ne =

primary nozzle exit

pm =

constant-pressure mixing section

gen =

generator

cond =

condenser

evap =

evaporator

  $αeff$ =

effective thermal conductivity (W/mK)

$δij$ =

Kronecker delta function

Δ =

change in the value

$μeff$ =

effective dynamic viscosity (Ns/m²)

ρ =

density (kg/m³)

τ =

stress tensor (Pa)

ω =

entrainment ratio

1D =

one-dimensional

3D =

three-dimensional

AR =

area ratio

CAM =

constant-area mixing

CFD =

computational fluid dynamics

COP =

coefficient of performance

CPM =

constant-pressure mixing

ER =

entrainment ratio

ERS =

ejector refrigeration system

GWP =

global warming potential

NXP =

nozzle exit position

ODP =

ozone depletion potential

VCRS =

vapor compression refrigeration system