Performance evaluation of low GWP large glide temperature zeotropic mixtures applied in air source heat pump for DHW production

Highlights

• Large glide temperature low GWP zeotropic mixtures are used for DHW production.

• The mixtures improve the heat pump performance notably.

• EER reaches 2.56 of heat pump for producing 65 °C hot water with environment temperature of −20 °C.

• Temperature match degree is proposed to reflect the condensing process.

Abstract

This study proposes the low GWP zeotropic mixtures applied in the air source heat pump for the domestic hot water production using the city water of 15 °C. The large glide temperature in zeotropic mixture side was applied to match temperature lift in hot water side in the condenser. Fifteen zeotropic mixtures, composed of the low boiling compositions (R744, R41, and R170) and high boiling point compositions (RE170, R600a, R1234yf, R1234ze(E), and R290), were compared. A new indicator named temperature match degree (TMD) was applied to evaluate the heat exchanging process. Besides, the system performances, including energy efficiency ratio (EER), exergy efficiency (η_ex), volumetric heating capacity (q_v, kJ/m³), are evaluated under different conditions. The results showed that the heat pumps using the zeotropic mixtures show better performance than that of the corresponding pure composition, except for q_v. The maximum EER, η_ex and q_v reaches 4.17, 0.472 and 3702 kJ/m³ with the mole fraction of R744/RE170 of 0.3/0.7.TMD was also improved by the zeotropic mixtures, and the maximum TMD reaches 76 % with the mole fraction of RE170/R600a of 0.6/0.4, which is triple for R600a. The results prove the potential of the heat pumps using the large glide temperature zeotropic mixtures applied in domestic hot water production.

Introduction

The building sector accounts for about 40 % of the society’s total energy consumption, and the domestic hot water (DHW) demand contributes an important section of energy consumption in the residential building, especially in cold regions [1], [2]. Air source heat pumps, as their high energy conversion efficiency and emission reduction ability, are regarded as an effective method to replace the electric water heaters, gas and fuel water heaters [3], [4]. The domestic hot water shows a large temperature lift under the heating process. The outlet water temperature ranges from 55 to 65 °C with the city water inlet for the standard requirement of different countries [5], [6]. For conventional subcritical cycle, the system performance is reduced significantly due to the temperature dis-matching under the heat exchanging process between the refrigerants and heat transfer medium [7], [8]. For the trans-critical cycle, the refrigerant and heat transfer medium show a good temperature match, and the low boiling point refrigerants are applied to achieve the trans-critical cycle due to the low critical temperature [9], [10], [11]. The trans-critical cycle usually shows high efficiency for domestic hot water production due to the high volumetric heating capacity and temperature match during the heat transfer process [12]. The large-scale applications of trans-critical cycle are limited by the two respects: 1) high initial costs and safety risks due to the high operating pressure of low boiling point refrigerants [13], [14]; 2) the large throttling loss during the expansion process [15], [16]. Therefore, to reduce the operating pressure and enhance the heat pump efficiency, various high boiling point refrigerants are added into the low boiling point refrigerants. Besides, the zeotropic mixtures show glide temperature to meet the thermal match during the heat transfer process and improve the heat pump performance. Therefore, numerous studies have been focusing on the application of large glide temperature zeotropic mixtures on heat pumps.

Currently, R744 is a widely used low boiling point component for the large glide temperature zeotropic mixtures, especially for the heat pumps. Jahar and Souvik [17] built the thermodynamic model of the high-temperature heat pump and assessed the possibility of zeotropic mixtures R744/R600a and R744/R600 to replace pure R114 for medium and high temperature heat pump. Compared with pure refrigerants, the mixtures improved the heat pump performance significantly due to the matching of temperature glides, and the heat pump showed the highest COP with the mass fraction of R744 of around 0.3. The operating pressures are reduced to 3.4 MPa. Luo et al. [18] built the experimental test to investigate the potential of applying zeotropic mixture R744/R600a in the air source heat pump based on the single stage compression cycle. The heating COP reached 1.834 for producing 75 °C with the low temperature environment of −30 °C. The single stage cycle using the mixture outperformed the two-stage and cascade system using R134a and R410A in cold regions. Sun et al. [19] introduced the zeotropic mixture R744/R32 applied in the water-to-water heat pump. The mass fraction of mixture affected the heat pump performance significantly and the highest performance occurred with the mass fraction R744/R32 of 0.6/0.4. Ju et al. [20], [21] conducted the experimental and simulated study of the zeotropic mixture R744/R290 applied in the heat pump water heater for domestic hot water production. The optimum R744/R290 (12 %/88 % mass fraction) showed the high potential to replace the R22 for heat pump water heaters. The maximum COP of the heat pump reached 4.731, which is 11.00 % over than that of pure R22. Zühlsdorf et al. [22], [23] conducted the thermodynamic evaluation of eco-friendly zeotropic mixtures applied in booster heat pumps, including 18 pure working fluids and corresponding zeotropic mixtures. Compared with pure R134a, the best performing mixture improved the COP by 47 %. The temperature glide of the mixtures improved the heat pump thermodynamic performance. Dai et al. [24], [25] conducted thermodynamic assessment of the CO₂-based mixtures applied in the heat pump cycle based on the energetic and exergetic model, including the domestic hot water production and drying process. The result showed that the system thermodynamic performance, including COP and exergy efficiency, is improved zeotropic mixture. The high boiling point working fluids reduced the system operating pressure. Zhang et al. [26] proposed the mixture R744/DME applied in a trans-critical cycle for producing domestic hot water, the operating pressure was reduced by the mass fraction of DME increasing. Similar results were found in the following literature [27], [28].

Overall, previous studies are major in focusing on the R744-based zeotropic mixtures applied in heat pumps. R41 and R170 showed a similar critical and boiling point as R744, which also showed the potential to work as the low point component for the large glide temperature zeotropic mixtures. Currently, a few studies have been focused on R41 and R170 working as the low boiling point refrigerants for large zeotropic mixtures, especially for heat pump conditions. Park and Jung [29], [30] proposed the mixture R170/R290 to replace pure R22 in the heat pump, the mixture R170/R290 at 4 %/96 % mass fraction showed a similar COP and capacity with those of R22, while the COP was lower than that of R290. Tang et al. [31] conducted the theoretical study of zeotropic mixture R41/R1234ze(E) applied in the heat pump water heater. The results showed that the heat pump using the mixture R41/R1234ze(E) (9 %/91 % mass fraction) has the maximum COP, which was 14.15 % higher than that of pure R22. The energetic and exergetic model also widely applied to evaluate the refrigerant cycle and component thermodynamic performance, such as the auto-cascade refrigeration cycle and heat exchanger [32], [33], [34], [35], [36], [37].

In this study, large glide temperature low-GWP zeotropic mixtures are proposed as the working fluids in air source heat pump for domestic hot water production, including three low boiling point components (R744, R41, and R170) and five high boiling point components (R290, R600a, RE170, R1234yf and R1234ze(E)), and there are fifteen pairs of zeotropic mixtures are compared and analyzed. All the selected refrigerants show low GWP and zero ODP. The detailed properties of the refrigerants are presented in Section 2. For the heat pump, the thermodynamic performance, energy efficiency ratio, exergy efficiency, operating pressure, volumetric heating capacity and temperature match degree are analyzed and compared for various working pairs. Besides, the effect of temperature of environment and operating on the heat pump performance are conducted. The major novelties of the present study are listed as follows:

• The large glide temperature low-GWP zeotropic mixtures are proposed to apply in the heat pump water heaters for domestic hot water production for the sub-critical cycle.

• A new indicator named temperature match degree (TMD) is proposed to reflect the approaching degree of the actual condensing to the ideal Lorenz process.

• High phase change latent refrigerant RE170 and low flammable R1234yf and R1234ze(E) are proposed as the high boiling point component for the large glide temperature mixtures.

The results could provide the basic theoretical foundation for the further research and improvement of the heat pump water heater using the large glide temperature zeotropic mixtures.

2. Research methodology.

The air source heat pump domestic hot water heater using the low GWP binary zeotropic mixtures is shown in Fig. 1, including the temperature-entropy and schematic diagram. The heat pump includes a compressor, gas cooler/condenser, throttle valve, and air source evaporator. The heat pump also consists of three loops, including the refrigerant loop, domestic hot water loop, and air loop.

Table 1 shows the basic properties of selected refrigerants. Eight refrigerants are selected as the component of the large glide temperature zeotropic mixture, including three low boiling point fluids (R170, R41 and R744) and five high boiling point fluids (R290, R600a, RE170, R1234yf and R1234ze(E)). All refrigerants show low GWP. The refrigerants are sorted by normal boiling point. The low boiling point fluids show low critical temperature and high critical pressure. The high boiling point fluids have high critical temperature and low critical pressure except for RE170. RE170 shows high critical temperature and high critical pressure.

Fig. 2 displays the latent heat of phase change from saturated liquid to saturated vapor under varied pressure of selected pure refrigerants [39]. Among the selected refrigerants, the refrigerants (R744, R170, R41, and RE170) show higher latent heat than the refrigerants (R1234yf, R1234ze(E), R600a, and R290) when the pressure overs 2.0 MPa.

There are several consumptions for the energetic and exergetic model [40], [41].

• Flows in each components keep steady.

• Throttling process keeps isenthalpic.

• Neglecting the pressure drop in the system.

• No heat exchange occurs for the system except for evaporator and condenser.

• Neglecting the kinetic and potential energy changes.

The cycle performance of domestic hot water heater is calculated based on the thermodynamic first and second laws. The energetic and exergetic efficiency indices include the energy efficiency ratio (EER), exergy efficiency (η_ex), volumetric heating capacity (q_v, kJ/m³), and system total exergy destruction per heating capacity unit (Ėx_,d,tot, kW/kW).

The specific exergy of working fluids at j point of a component is calculated by Equation (1) [37]:

$${\mathit{ex}}_j=\left(h_j-h_0\right)-T_0\left(s_j-s_0\right)$$

The reference temperature (T₀) and pressure (P₀) are set as 25 °C and 0.101 MPa [23].

The mass flow rate, power consumption, exergy efficiency and destruction of the compressor are calculated by Equations (2), (3). (4) [25]:

$$W_{\mathit{comp}}={\dot{m}}_1\times \left(h_2-h_1\right)={\dot{m}}_1\times \left(h_{2s}-h_1\right)/{\eta }_s$$

$$W_{\mathit{comp}}+{\dot{m}}_1\times {\mathit{ex}}_1={\dot{m}}_1\times {\mathit{ex}}_2+{\mathit{Ex}}_{d,comp}$$

$${\mathit{Ex}}_{d,comp}={\dot{m}}_1\times T_0\times (s_2-s_1)$$

Where, ṁ₁represents the mass flow rate in the system, kg/s; W_comp represents the power consumption of the compressor, kW; Ex_d,comp represents the exergy destruction of the compressor, kW;.η_s represents the isentropic efficiency of the compressor [40], [41], which is calculated by Equation (5)

$${\eta }_s=1.003-0.121\times \frac{P_2}{P_1}$$

The condenser/gas cooler is tube-in-tube type, and the refrigerant flows through the inner tube and the condensing water flows through the annular region with a counterflow type. The exergy destruction and heating capacity of the condenser are calculated as follows:

The heating capacity can be calculated by Equation (6) [25]

$$Q_h={\dot{m}}_2\times \left(h_2-h_3\right)={\dot{m}}_{\mathit{wa}}\times C_{\mathit{pw}}\times (T_{\mathit{wa},out}-T_{\mathit{wa},in})$$

The exergy destruction can be calculated by Equations (7), (8)[25]

$${\dot{m}}_2\times {\mathit{ex}}_2+{\dot{m}}_{\mathit{wa},in}\times {\mathit{ex}}_{\mathit{wa},in}={\dot{m}}_3\times {\mathit{ex}}_3+{\dot{m}}_{\mathit{wa},out}\times {\mathit{ex}}_{\mathit{wa},out}+{\mathit{Ex}}_{d,cond}$$

$${\mathit{Ex}}_{d,cond}={\dot{m}}_2\times \left({\mathit{ex}}_2-{\mathit{ex}}_3\right)-{\dot{m}}_{\mathit{wa},in}\times \left({\mathit{ex}}_{w,in}-{\mathit{ex}}_{w,out}\right)$$

The exergy destruction of the throttle valve is calculated by Equations (9), (10) [25]

$${\dot{m}}_1\times {\mathit{ex}}_3={\dot{m}}_1\times {\mathit{ex}}_4+{\mathit{Ex}}_{d,va}$$

$${\mathit{Ex}}_{d,va}={\dot{m}}_1\times {(ex}_3-{\mathit{ex}}_4)$$

The exergy destruction of the evaporator is calculated by Equations (11), (12) [25]

$${\dot{m}}_4\times {\mathit{ex}}_4+{\dot{m}}_{\mathit{air}}\times {\mathit{ex}}_{\mathit{air},in}+W_{\mathit{fan}}={\dot{m}}_1\times {\mathit{ex}}_1+{\dot{m}}_{\mathit{air}}\times {\mathit{ex}}_{\mathit{air},out}+{\mathit{Ex}}_{d,evap}$$

$${\mathit{Ex}}_{d,evap}={\dot{m}}_4\times ({\mathit{ex}}_4-{\mathit{ex}}_1)+{\dot{m}}_{\mathit{air}}\times ({\mathit{ex}}_{\mathit{air},in}-{\mathit{ex}}_{\mathit{air},out})+W_{\mathit{fan}}$$

Where, W_fan is the power consumption of the fan in the evaporator and calculated by Equation (13) [42], [43]

$$W_{\mathit{fan}}=\frac{\mathrm{\Delta }P\times L}{{\eta }_t\times {\eta }_c}$$

Where, ΔP represents the head of fan, Pa; L represents the volumetric flow rate, m³/h; η_t represents the total efficiency; and η_c represents the mechanical efficiency [42], [43].

Then, the energy efficiency ratio (EER) is calculated by Equation (14)

$$\mathit{EER}=\frac{Q_h}{W_{\mathit{com}}+W_{\mathit{fan}}}$$

The volumetric heating capacity (q_v, kJ/m³) is calculated by Equation (15)

$$q_v=\frac{Q_h}{m_1\times v_1}$$

The system total exergy destruction is calculated by Equation (16) [25]

$${\mathit{Ex}}_{d,tot}={\mathit{Ex}}_{d,comp}+{\mathit{Ex}}_{d,cond}+{\mathit{Ex}}_{d,va}+{\mathit{Ex}}_{d,evap}$$

The system exergy efficiency is calculated by Equation (17) [25]

$${\eta }_{\mathit{ex}}=1-\frac{{\mathit{Ex}}_{d,tot}}{W_{\mathit{comp}}+W_{\mathit{fan}}}$$

The system exergy destruction per heating capacity is calculated by Equation (18) [44]

$${\dot{E}x}_{d,tot}=\frac{{\mathit{Ex}}_{d,tot}}{Q_h}$$

For the ideal Lorenz cycle, the refrigerant is condensed from point 2′ to point 3′, while for the actual condensing cycle, the refrigerant is condensed along with the point 2–2 s − 2 L −3. The temperature difference is 5 °C and keeps the same during the condensing process for ideal process, while the actual temperature difference is variation during the condensing process and the average difference is set as 5 °C in this study. To reflect the approaching degree of the actual condensing to the ideal Lorenz process, the temperature match degree is introduced in this study. The temperature match degree (TMD) is calculated as Equation (19). When the $S_{2-6^{\text{'}}-2L-3-3^{\text{'}}-2^{\text{'}}-2}$ equals zero, the TMD equals unity, and the condensing process presents the ideal Lorenz process.

$$\mathit{TMD}=\frac{S_{2^{\text{'}}-3^{\text{'}}-5-6-2\text{'}}}{S_{2^{\text{'}}-3^{\text{'}}-5-6-2\text{'}}+S_{2-6^{\text{'}}-2L-3-3^{\text{'}}-2^{\text{'}}-2}}$$

Table 2 displays the operating parameters of the heat pump. In the condenser, the weighted average logarithmic mean temperature difference is set as 5 °C, and the value is calculated by Equation (20). In the evaporator, the average logarithmic mean temperature difference is set as 10 °C for air-cooled type. The environment air temperature ranges from −20 to 30 °C to cover yearly round conditions. The water temperature in the condenser outlet ranges from 55 to 65 °C, which covers the sanity water demand in different countries [5].

With the consideration of the temperature glide during the condensing process and to reduce the effect of pinch point, the condensing process is divided into three sub section as shown in Fig. 3, including the super heating section, two-phase section and subcooling section. The logarithmic mean temperature difference in the condenser ΔT_cond is defined as the weighted average of the three sections [45], and calculated by Equation (20)

$$\frac{1}{\mathrm{\Delta }{T}_{\mathrm{c}\mathrm{o}\mathrm{n}\mathrm{d}}}=\frac{\mathrm{a}}{\mathrm{\Delta }{T}_{\mathrm{m},\mathrm{s}\mathrm{p}}}+\frac{\mathrm{b}}{\mathrm{\Delta }{T}_{\mathrm{m},\mathrm{t}\mathrm{p}}}+\frac{\mathrm{c}}{\mathrm{\Delta }{T}_{\mathrm{m},\mathrm{s}\mathrm{c}}}$$

Where, a, b, and c are the fraction of heat transferred in each section, and calculated as below[45]:

$$a=\frac{Q_{\mathit{sp}}}{Q_h},b=\frac{Q_{\mathit{tp}}}{Q_h},c=\frac{Q_{\mathit{sc}}}{Q_h}$$

$$Q_{\mathit{sp}}={\dot{m}}_{\mathit{ref}}\times \left(h_2-h_{2s}\right)$$

$$Q_{\mathit{tp}}={\dot{m}}_{\mathit{ref}}\times \left(h_{2s}-h_{2l}\right)$$

$$Q_{\mathit{tp}}={\dot{m}}_{\mathit{ref}}\times \left(h_{2l}-h_3\right)$$

$$\mathrm{\Delta }{T}_{m,sp}=\frac{{(T}_{r,2}-T_{w,6})-{(T}_{r,2s}-T_{w,6\text{'}})}{\mathrm{l}\mathrm{n}\frac{{(T}_{r,2}-T_{w,6})}{{(T}_{r,2s}-T_{w,6\text{'}})}}$$

$$\mathrm{\Delta }{T}_{m,tp}=\frac{{(T}_{r,2s}-T_{w,6\text{'}})-{(T}_{r,2l}-T_{w,5\text{'}})}{\mathrm{l}\mathrm{n}\frac{{(T}_{r,2s}-T_{w,6\text{'}})}{{(T}_{r,2l}-T_{w,5\text{'}})}}$$

$$\mathrm{\Delta }{T}_{m,sc}=\frac{{(T}_{r,2l}-T_{w,5\text{'}})-{(T}_{r,3}-T_{w,5})}{\mathrm{l}\mathrm{n}\frac{{(T}_{r,2l}-T_{w,5\text{'}})}{{(T}_{r,3}-T_{w,5})}}$$

For the evaporating process, the logarithmic mean temperature difference in evaporator (ΔT_evap) is calculated by Equation (28) [45]

$$\mathrm{\Delta }{T}_{\mathit{evap}}=\frac{{(T}_8-T_4)-{(T}_7-T_{1\text{'}})}{\mathrm{l}\mathrm{n}\frac{{(T}_8-T_4)}{{(T}_7-T_{1\text{'}})}}$$

Based on the model, the procedure is presented as shown in Fig. 4. The calculation procedure needs some input parameters, displayed as below. The mixture type and mole fraction, the hot water inlet and outlet temperature (T₅ and T₆), the air inlet and outlet temperature (T₇ and T₈), the subcooled and super heating temperature (T_sp and T_sc), and the mass flow rate of refrigerant (ṁ₁) in the system is set to 1 kg/s. In the following section, the performance of the heat pump water heater will be investigated in detail.