Performance comparison of heat pumps using low global warming potential refrigerants with optimized heat exchanger designs

Journal Pre-proofs Performance comparison of heat pumps using low global warming potential refrigerants with optimized heat exchanger designs Byeongsu Kim, Sang Hun Lee, DongChan Lee, Yongchan Kim PII: DOI: Reference:

S1359-4311(19)36511-1 https://doi.org/10.1016/j.applthermaleng.2020.114990 ATE 114990

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Applied Thermal Engineering

Received Date: Revised Date: Accepted Date:

19 September 2019 10 January 2020 22 January 2020

Please cite this article as: B. Kim, S. Hun Lee, D. Lee, Y. Kim, Performance comparison of heat pumps using low global warming potential refrigerants with optimized heat exchanger designs, Applied Thermal Engineering (2020), doi: https://doi.org/10.1016/j.applthermaleng.2020.114990

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Performance comparison of heat pumps using low global warming potential refrigerants with optimized heat exchanger designs Byeongsu Kim, Sang Hun Lee, DongChan Lee, Yongchan Kim* Department of Mechanical Engineering, Korea University 145 Anam-ro, Seongbuk-gu, Seoul 02841, Republic of Korea

Abstract The objective of this study is to analyze the performance improvement of heat pumps with low global warming potential (GWP) refrigerants over conventional R-410A heat pumps by optimizing the heat exchanger designs. A simulation model is developed for heat pumps with low GWP refrigerants and validated by comparing the predicted and measured data for R-410A and R-32. Based on the developed model, the diameter of the tube and the number of paths in the indoor and outdoor heat exchangers are optimized for each alternative refrigerant using the genetic algorithm. The performances of the optimized heat pumps with the low GWP refrigerants are compared with that of the R-410A heat pump to provide a selection guide for alternative refrigerants. The average performance improvements of the heat pumps with optimized heat exchanger designs were estimated to be 6.0%, 3.4%, and 2.4% for R-32, DR-5, and L-41a, respectively, compared with that of the baseline system. Thus, the R-32 heat pump delivered the highest performance improvement with the optimized heat exchanger designs.

Keywords: Alternative refrigerant, Low GWP, Heat pump, Heat exchanger design, Performance improvement

* Corresponding author. E-mail address: [email protected]

Nomenclature A

area (m2)

Bo

Bond number

C

heat capacity (kJ K–1)

D

diameter (mm)

Dc

fin collar outside diameter (mm)

Dh

hydraulic diameter (mm)

E

enhancement factor

F

reduced parameter

f

friction factor

Fp

fin pitch (mm)

Fr

Froude number

Fs

fin spacing (mm)

G

mass flux (kg s–1 m–2)

h

heat transfer coefficient (kW m–2 K–1)

Hz

compressor frequency (s–1)

i

enthalpy (kJ kg–1)

j

Colburn factor

k

thermal conductivity (kW m–1 K–1)

L

length (m)

Lh

louver height (mm)

Lp

louver pitch (mm)

m

mass flow rate (kg s–1)

N

number of paths

n

coefficient of equation

Nu

Nusselt number

P

pressure (kPa)

PF

penalty factor

Pl

pitch of heat exchanger tube column (mm)

Pr

Prandtl number

Pt

pitch of heat exchanger row (mm)

Q

heat transfer rate (kW)

R

number of longitudinal tube rows

r

radius (m)

Re

Reynolds number

Rx

parameter for effect of increasing heat transfer area

S

suppression factor

Sh

slit height (mm)

Sn

number of slits in an enhanced zone

Ss

breadth of a slit in the direction of airflow (mm)

T

temperature (℃)

t

fin thickness (mm)

U

overall heat transfer coefficient (kW m–2 K–1)

u

velocity (m s–1)

V

volume (m3)

W

power consumption (kW)

X

measured variables

Greek letters



effectiveness



two-phase frictional multiplier



efficiency



density (kg m–3)



uncertainty

Subscripts air

air

c

condensing

comp

compressor

cond

condenser

crit

critical

d

discharge

disp

displacement

e

evaporating

evap

evaporator

i

inner side

ID

indoor

IDU

indoor unit

in

in

isen

isentropic

l

liquid

max

maximum

min

minimum

normal

normalized

o

outer side

OD

outdoor

ODU

outdoor unit

out

out

pool

pool boiling

ref

refrigerant

sc

subcooling

sh

superheat

tp

two-phase

vol

volumetric

Acronyms COP

coefficient of performance

CR

compression ratio

EER

energy efficiency ratio

EEV

electronic expansion valve

GWP

global warming potential

HC

hydrocarbon

HFC

hydrofluorocarbon

HFO

hydrofluoroolefin

NTU

number of transfer units

1. Introduction Based on the 1997 Kyoto Protocol, the applications of hydrofluorocarbon (HFC) refrigerants have been restricted due to their high global warming potentials (GWPs). In response to this regulation, major global refrigerant companies have developed various low GWP refrigerants. Specifically, extensive studies have been conducted to determine alternative refrigerants with low GWPs to replace R-410A used in residential air-conditioners and heat pumps. However, comparative performance evaluations of heat pumps using low GWP refrigerants should be performed to determine an optimal alternative to R-410A. Many studies have been performed to compare the thermodynamic properties of alternative refrigerants and to predict their performances using a simple cycle analysis. Bobbo et al. [1] analyzed the thermodynamic properties of hydrofluoroolefin (HFO) refrigerants and their mixtures. Devecioglu and Oruc [2] reported that R-1234yf, L-40, R-444B, and DR-5 can be employed as alternatives to R-134a, R-404A, R-22, and R-410A, respectively, based on the analyses of their thermodynamic properties. Shaik and Babu [3,4] compared the thermodynamic performance of HFC-hydrocarbon (HC) mixtures in a refrigeration cycle. Choudhari and Sapali [5] investigated the possibility of replacing R-22 with R-290 using a simulation analysis. In addition, López-Belchí and Illán-Gómez [6] compared the condensation heat transfer characteristics of R-32 and R-410A inside mini-channels. Lillo et al. [7] analyzed the flow boiling characteristics of R-32 in a stainless-steel tube and compared the results with those of R-410A. Longo et al. [8] compared the flow boiling and condensation heat transfer characteristics of alternative refrigerants of R-410A and R-404A in a brazed plate heat exchanger. The drop-in performances of alternative refrigerants for residential air conditioners and heat pumps have been analyzed through simulations and experiments. Domanski et al. [9] proposed optimal alternative refrigerants for small- and medium-sized air conditioners in

heating and refrigeration applications based on their respective thermodynamic analyses. They concluded that refrigerant mixtures could be suitable alternatives to R-410A and R-404A considering coefficient of performance (COP), volumetric capacity, flammability, and GWP. Babiloni et al. [10] reviewed studies related to the performances and thermodynamic characteristics of alternative refrigerants for R-410A residential air conditioners. Moreover, Sagia and Rakopoulos [11] compared the performances of geothermal heat pumps by applying R-22, R-407B, and a binary mixture of R-32 and R-134a. Nawaz et al. [12] compared the dropin performances of R-290 and R-600a in an R-134a air-to-water heat pump through simulations. The R-290 heat pump exhibited a similar performance to the R-134a heat pump. Shen et al. [13] compared the drop-in performances of alternatives to R-22 and R-410A in a rooftop heat pump. R-452B was suggested to be the best alternative refrigerant to R-410A. Afshari et al. [14] analyzed the efficiencies of heat pumps by applying R-407C and R-404A. In addition, Tian et al. [15], Xu et al. [16], and Cheng et al. [17] experimentally evaluated the drop-in performances of R-290 and R-32 in R-410A residential air conditioners. The drop-in performance of L-41b was experimentally investigated in an R-410A heat pump [18], and the seasonal performances of R-446A, R-447A, and R-454B heat pumps were theoretically compared with that of an R-410 heat pump [19]. Yao et al. [20] measured the drop-in performance of R-32 in direct evaporative all fresh air handling units using R-410A. The performance of the R-32 system without optimization was slightly lower than that of the R410A system. Xu et al. [21] and Baek et al. [22] also conducted experiments to evaluate the performance improvement of vapor injection heat pumps using R-32 and CO2, respectively. Chen et al. [23] compared the performances of air-to-water vapor injection heat pumps using R-410A, R-447B, and R-452B. The COPs of the vapor injection heat pumps using R-447B and R-452B were 1.4–2.0% and 0.4–3.8%, respectively, higher than that of the vapor injection heat pump using R-410A.

These previous studies associated with low GWP refrigerants have focused on comparing simple thermodynamic cycles and drop-in performances of heat pumps. The components of heat pumps using low GWP refrigerants should be optimally redesigned to improve their actual performances at various operating conditions. In addition, a design guide to the optimum heat flux condition for each alternative refrigerant should be provided for effective system operation. However, studies on performance improvements of heat pumps with alternative refrigerants through the optimization of component designs are rather limited in the open literature. The objective of this study was to analyze the performance improvement of heat pumps with low GWP refrigerants over the R-410A heat pump by optimizing the heat exchanger designs. An air-to-air heat pump simulation model was developed and validated by comparing the predictions with the measured data under various operating conditions. With the developed simulation model, the performance of the heat pump was estimated for various alternative refrigerants under standard cooling and heating conditions. The design parameters of the indoor and outdoor heat exchangers, such as the tube diameter and number of paths, were optimized for each alternative refrigerant. In addition, the performances of the optimized heat pumps were compared to provide a selection guide for alternative refrigerants.

2. Experiments Fig. 1 shows a schematic of the experimental setup for measuring the performance of the heat pump under various operating conditions. The heat pump consists of a hermetic scroll compressor with a displacement volume (Vdisp) of 31.6 cm3 rev–1, an electronic expansion valve (EEV), and a couple of fin-tube heat exchangers. Based on preliminary tests, the optimal refrigerant charge amount to yield the maximum energy efficiency ratio (EER) and COP under standard cooling and heating conditions was determined to be 3800 g for R-410A and 3000 g for R-32. A louver fin-tube heat exchanger with a heat transfer area of 44.9 m2 was used for

the condenser (in the cooling mode) and evaporator (in the heating mode) in the outdoor unit, and a slit fin-tube heat exchanger with a heat transfer area of 28.9 m2 was used for the evaporator (in the cooling mode) and condenser (in the heating mode) in the indoor unit. The number of outdoor heat exchanger paths in the cooling mode was eight, and that in the heating mode was 16. The refrigerant mass flow rate ( m ) was controlled by the EEV. The orifice diameters of the EEV were 3.0 mm and 2.4 mm for R-410A and R-32, respectively. Detailed specifications of the heat pump components are listed in Table 1. The performance of the heat pump was measured using an air-enthalpy-type psychrometric calorimeter. The air flow rate was calculated using the nozzle method based on the measured pressure difference across the nozzle. The pressure and temperature of the refrigerant were measured using pressure transducers with an accuracy of ±0.25% and T-type thermocouples with an accuracy of ±0.2 ℃, respectively, at both ends of each component. The total power consumption of the compressor (Wcomp) and fans (WIDU Fan, WODU Fan) was measured using a power meter with an accuracy of ±0.2%. The temperature, pressure, refrigerant mass flow rate, and power consumption were measured for 10 min, and the time-averaged values were used for the performance evaluation. As defined in Eq. (1), the cooling or heating capacity (Q) was calculated based on the air flow rate and enthalpy difference across the indoor unit. As defined in Eq. (2), the EER (or COP) was calculated using the ratio of the cooling (or heating) capacity to power consumption.

Q  m air (iair , out  iair ,in ) EER  or COP  

(1)

Q Wcomp  WIDU Fan  WODU Fan

(2)

As given in Eq. (3), the total uncertainty of the reduced parameter ( F ) was determined using the uncertainty propagation method with directly measured values [24]. The total

uncertainty of the individual parameters ( X ) was estimated as the sum of the systematic and random errors. 2

2

2

 F   F   F  F   X1    X2      Xn   X1   X2   Xn 

(3)

For the total uncertainty of the capacity (F=Q), the measured variables of X1, X2, and X3 were m air , iair,out, and iair,in., respectively. For the total uncertainty of the EER (F=EER or COP), the measured variables of X1 and X2 were Q and W, respectively. The enthalpies were calculated based on the measured pressure and temperatures of air. As a result, the average uncertainties of the capacity and EER (or COP) were estimated as ±3.19% and ±3.18%, respectively. The maximum uncertainties of the capacity and EER (or COP) among the tested cases were ±4.88% and ±5.03%, respectively.

3. Simulation model 3.1. Heat pump simulation model A simulation model for heat pumps using alternative refrigerants was developed based on the Oak Ridge National Laboratory method [25]. Fig. 2 represents the flow of the heat pump simulation model. The convergence of the model was checked by its subcooling, superheat, and capacity. The specifications of the components (compressor, heat exchanger, and expansion device) and operating conditions were provided as input data. In addition, the condensing (Tc) and evaporating (Te) temperatures were initially assumed, and the target superheat (Tsh) and subcooling (Tsc) were set. Simulations were conducted in the compressor, condenser, and evaporator models (in that sequence). The refrigerant mass flow rate, compressor outlet condition, and power consumption were calculated using the compressor model. Subsequently, the condenser performance was calculated, and the convergence for the

subcooling was checked. An iterative calculation was performed by adjusting the condensing temperature until the convergence for subcooling was satisfied. Accordingly, the evaporator performance was estimated, and the convergence for the superheat was checked. An iterative calculation was performed in this case and by adjusting the evaporating temperature to obtain the converged superheat. The convergence of the capacity was then checked, and an iterative calculation was performed by regulating the compressor frequency (Hz). After all the convergence criteria were satisfied, the simulation results were stored. For the system optimization, the aforementioned process was repeated by varying the subcooling until the maximum EER (or COP) was obtained. Theoretically, the system efficiency increases when the superheat is decreased. Therefore, in the system optimization, the superheat was fixed at a controllable minimum value for better convergence. The heat exchanger optimization for each alternative refrigerant was performed using the genetic algorithm [26]. The objective function of the optimization was set to maximize the EER (or COP) at each operating condition. The population size and crossover fraction were 200 and 0.8, respectively. The convergence of the objective function was checked by a tolerance of 10–6, and the maximum number of iterations was limited to 100. The ranges of the optimization variables are shown in Table 2. The heat exchanger performance was analyzed according to design parameters and operating conditions. Moreover, the performance of the heat pump with the optimized heat exchanger designs was estimated using the developed simulation model. In this study, R-32, R-446A, DR-5, and L-41a were adopted as alternative refrigerants of R-410A in heat pumps. The properties of alternative refrigerants were estimated using REFPROP 9.1 [27]. The basic physical properties of these alternative refrigerants are listed in Table 3.

3.2. Compressor model

The performance of a hermetic scroll compressor was analyzed using an efficiency model. As defined in Eq. (4), the mass flow rate of the refrigerant was determined by the compressor frequency, displacement, and volumetric efficiency (  v o l ). As defined in Eq. (5), the volumetric efficiency of the compressor was approximated using a multiple regression as the function of compressor frequency, condensing temperature, and evaporating temperature based on the performance data of the compressor provided by compressor manufacturers.

m  vol HzVdisp compin

(4)

vol  0.76846  0.00277Hz  0.00169Tc  0.00168Te

(5)

As defined in Eq. (6), the power consumption of the compressor was calculated in terms of the refrigerant mass flow rate, enthalpy difference across the compressor, and isentropic efficiency (  isen ). The isentropic efficiency was also derived using a method similar to that used for calculating the volumetric efficiency, as shown in Eq. (7).

Wcomp 

m (icomp out , isen  icomp in )

 isen

isen 1.65516  0.00502Hz  0.00780Tc  0.00779Te

(6)

(7)

3.3. Heat exchanger model Air-cooled heat exchangers with louver fins and slit fins were employed in the condenser and evaporator, respectively. The tube-by-tube method was used to analyze the heat exchanger performance. This method used the exit state of the preceding segment as the inlet state of the subsequent segment. The heat exchanger was separated into several uniform segments. The



-number of transfer units (NTU) method was used to calculate the heat transfer

rate, which was derived using the product of maximum possible heat transfer rate (Qmax) and

effectiveness (  ). The NTU was determined from the overall heat transfer coefficient (UA). The related equations are presented in Eqs. (8)–(11):

ln  ro ri  1 1 1    UA href Ai hair Ao 2 kL

(8)

UA Cmin

(9)

 1e(NTU)

(10)

Q   Qmax

(11)

NTU 

For the air, the heat transfer coefficient and friction factor of the outdoor unit with louver fins and those of the indoor unit with slit fins were calculated using the Wang et al. correlation [28,29]. For the refrigerant, the single-phase heat transfer coefficient was calculated using the Dittus–Boelter correlation [30]. The condensation and evaporation heat transfer coefficients were calculated using the correlations suggested by Cavallini et al. [31] and Gungor and Winterton [32], respectively. The single- and two-phase pressure drops were calculated using the Blasius [33] and Newell and Shah [34] correlations, respectively. The referred empirical formulas for the heat transfer coefficient and pressure drop in the heat exchanger analysis are listed in Table 4. The heat exchanger model was developed using a uniform path design based on a given number of heat exchanger columns. The power consumptions of the indoor and outdoor fans were calculated using the measured data according to the volumetric air flow rate.

3.4. Model validation Fig. 3 shows the validation results of the heat pump simulation model by comparing the predictions with experimental data for the heat pumps that employ R-410A and R-32. Thirtynine experimental cases were conducted by varying the outdoor temperatures in the range of

29–35 ℃ in the cooling mode and –15–7 ℃ in the heating mode. The indoor temperature was fixed at 27 ℃ and 20 ℃ in the cooling and heating modes, respectively. Experiments were performed by varying the target capacity in the range of 4.3–14.5 kW in the cooling mode and 5.8–16.7 kW in the heating mode by controlling the compressor frequency and air flow rate. Fig. 3 shows that the calculated capacity and power consumption were consistent with the measured data within deviations of ±1% and ±10%, respectively.

4. Results and discussion 4.1. Drop-in performance A drop-in simulation was performed for a split-type heat pump designed for R-410A with nominal cooling and heating capacities of 14.5 kW and 16.7 kW, respectively. In addition, the nominal cooling and heating capacities of the other alternative refrigerants were the same as those of R-410A. Fig. 4 shows the normalized EERs (or COPs) of heat pumps that employ alternative refrigerants. To achieve a uniform path design, eight outdoor heat exchanger paths were maintained in the cooling mode, and 16 were maintained in the heating mode. The outdoor temperatures in the cooling and heating modes were maintained at 35 ℃ and 7 ℃, respectively, whereas the indoor temperatures were maintained at 27 ℃ and 20 ℃, respectively. In the cooling mode, the normalized EERs of the heat pumps using R-32, R-446A, DR-5, and L-41a over that of the R-410A heat pump were 104.9%, 93.4%, 101.2%, and 98.5%, respectively. However, in the heating mode, the normalized COPs of the heat pumps using R-32, R-446A, DR-5, and L-41a were 102.7%, 97.6%, 96.9%, and 100.8%, respectively. As listed in Table 3, the latent heats of the alternative refrigerants were observed to be higher than that of R-410A. Accordingly, as listed in Table 5, the refrigerant mass flow rates of the heat pumps using the alternative refrigerants were substantially lower than that of the R410A heat pump in obtaining a similar target capacity. In this case, the compressor frequency

was reduced to obtain a lower refrigerant mass flow rate, which led to an increase in the EERs (or COPs). However, the liquid densities of the alternative refrigerants were lower than that of R-410A. In this case, the compressor frequency was increased to maintain the required refrigerant mass flow rate, which led to a decline in the EERs (or COPs). The counter-effects of the higher latent heat and lower liquid density were critical in determining the performance of alternative refrigerants. For the R-32 heat pump, the increase in the latent heat was observed to be more substantial than the decrease in the liquid density, which led to an increase in the EER (or COP). In addition, as the compression ratio (CR) increased, the EER and COP decreased due to the reduction in compression efficiency. As shown in Table 5, the R-446A heat pump exhibited a higher compression ratio than those using other refrigerants in the heating and cooling modes, which resulted in a substantially lower EER (or COP) than that of the R-410A heat pump. This was because the saturation pressure of R-446A was 70–330 kPa lower than those of the other refrigerants at a given temperature. Furthermore, the DR-5 heat pump demonstrated a substantially lower COP than the R-410A heat pump due to its higher compression ratio in the heating mode.

4.2. Optimization of heat exchanger design In this study, preliminary simulations were conducted to determine the possible optimum ranges of the tube diameter (D) and number of paths (N) in the heat exchanger with the low GWP refrigerants. Based on the preliminary simulations, the possible optimum ranges of the tube diameter and number of paths were determined to be 5–9 mm, and 4–16 for the outdoor unit and 3–12 for the indoor unit, respectively. Accordingly, the optimization of the heat exchanger design was conducted by comparing the performances of the heat pumps with the variations in the tube diameter and number of paths. The tube diameters of the indoor and outdoor heat exchangers were set to 5, 7, and 9 mm, which are commonly used in the actual

heat pumps. The number of heat exchanger paths was set to 4, 8, and 16 for the outdoor unit (64 columns) and 3, 6, and 12 for the indoor unit (12 columns), which enables a uniform path design based on the number of heat exchanger columns. The other design parameters of the heat exchanger and the target capacities were fixed at values specified in the drop-in simulation. Fig. 5 shows the normalized EERs of the heat pumps that used alternative refrigerants as compared to that of the R-410A heat pump in the cooling mode according to the tube diameter and number of paths in the evaporator. As shown in Fig. 5(a), the normalized EER of the R-32 heat pump was the highest in the cooling mode, whereas that of the R-446A heat pump was the lowest. The normalized EERs of the heat pumps using the alternative refrigerants decreased with an increase in the evaporator tube diameter. The EERs of these pumps decreased by 5.5– 10.0% with the increase in the evaporator tube diameter from 5 to 9 mm due to the decreased heat transfer coefficient with decreased refrigerant velocity. Fig. 5(b) shows the normalized EERs of the heat pumps in the cooling mode according to the number of evaporator paths. As the number of evaporator paths was increased, the EERs increased and then peaked when the number of evaporator paths was nine because of a substantial decrease in the pressure drop. However, when the number of evaporator paths was increased to more than nine, the EERs began to decrease because of a substantial decrease in the heat transfer coefficient. Therefore, the optimal number of evaporator paths for the alternative refrigerants was determined to be nine in the cooling mode. The heat transfer coefficient of R-32 decreased by 38.9%, 27.8%, and 22.3% with the increase in the number of evaporator paths from 3 to 6, 9, and 12, whereas the pressure loss dropped by 42.4%, 27.7%, and 20.7%, respectively. As the number of evaporator paths was increased from 3 to 6, the decrease in the heat transfer coefficient was lower than that in the pressure loss. However, as the number of evaporator paths was increased from 3 to 12, this trend changed to its opposite.

Fig. 6 shows the normalized EERs of the heat pumps in the cooling mode according to the tube diameter and number of paths in the condenser. As shown in Fig. 6(a), for all the alternative refrigerants, the maximum variation in the EERs was less than 1.5% with an increase in the condenser tube diameter. Therefore, the effect of the condenser tube diameter on the EER was obviously negligible. This is because the following factors affecting the EER compensated each other: the reduction in the heat transfer coefficient, increase in the heat transfer area, and reduction in the pressure loss with increasing tube diameter. Fig. 6(b) shows the normalized EERs of the heat pumps according to the number of condenser paths. The optimal number of condenser paths in the cooling mode was observed to be 10 for the R-32 heat pump and 12 for the other refrigerants. With the optimal number of condenser paths, the EER of the R-32 heat pump was 11.5% higher than that of the R-446A heat pump. Fig. 7 shows the normalized COPs of the heat pumps in the heating mode according to the tube diameter and number of paths in the evaporator. In the heating mode, the COP of the R32 heat pump was observed to be the highest, whereas that of the DR-5 heat pump was the lowest. This is because DR-5 had a higher compression ratio and lower latent heat with a relatively higher vapor density as compared with the other alternative refrigerants at a given evaporating temperature. As shown in Fig. 7(a), the normalized COPs of these heat pumps dropped by 2.7–5.0% with the increase in the evaporator tube diameter from 5 to 7 mm due to the dominant effect of the reduced heat transfer coefficient. However, as the evaporator tube diameter was increased from 7 to 9 mm, the variation in the COP remained nearly constant within a deviation of 1.0% due to the counter-balancing effects of the reduced heat transfer coefficient and pressure drop. Therefore, the optimal evaporator tube diameter for achieving a maximum COP was observed to be 5 mm in the heating mode. Fig. 7(b) shows that the normalized COPs of the heat pumps peaked at a certain number of evaporator paths, which was considered the optimal value. In the heating mode, the optimal number of evaporator paths in

the heat pumps using R-32 and DR-5 was 12, whereas that in the heat pumps using R-446A and L-41a was 10. For all the alternative refrigerants, the maximum performance improvement of heat pumps obtained by the optimization of the number of evaporator paths was within the range of 3.5–6.0%. Fig. 8 shows the normalized COPs of heat pumps in the heating mode according to the tube diameter and number of paths in the condenser. In the heating mode, as shown in Fig. 8(a), the R-32 heat pump demonstrated a COP that was at least 1.9–3.1% higher than those of the heat pumps using other refrigerants. In addition, the normalized COPs of the heat pumps dropped by 2.6–7.6% with the increase in the condenser tube diameter from 5 to 9 mm due to the dominant effect of the decreased heat transfer coefficient. Therefore, the optimal condenser tube diameter in the heating mode was determined to be 5 mm. Fig. 8(b) shows the normalized COPs of the heat pumps according to the number of condenser paths. As the number of condenser paths was increased from 3 to 12, the COPs of these pumps dropped by 3.5–7.3% due to the decreased heat transfer coefficient with a lower mass flow rate. Specifically, the DR5 heat pump showed the largest decline in COP based on the number of condenser paths, which was due to the substantial decrease in the heat transfer coefficient. The decrease in the refrigerant velocity of the DR-5 heat pump with an increased number of condenser paths was greater than those of the other alternative refrigerants due to the lower liquid density. Accordingly, the heat transfer coefficient decreased rapidly with the increase in the number of condenser paths, which led to a considerable decrease in the COP.

4.3. Comparison of performances with the optimized heat exchanger design The refrigerant mass flux is critical in determining the heat transfer performance and pressure drop in the heat exchanger. Therefore, determining the optimal range of the refrigerant mass flux with the maximum EER (or COP) for each alternative refrigerant is necessary. Fig.

9 shows the normalized EERs (or COPs) of the heat pumps that used alternative refrigerants as compared to that of the R-410A heat pump according to the mass fluxes in the evaporator and condenser. The EERs (or COPs) of the heat pumps with the optimized design parameters were estimated according to the mass fluxes of the condenser and evaporator. Based on this contour plot, the optimum mass flux for each alternative refrigerant could be determined to achieve the maximum EER (or COP). In the cooling mode, as shown in Fig. 9(a), the R-32 heat pump showed the highest EER at evaporator and condenser mass fluxes of 450–700 and 300–450 kg s–1 m–2, respectively. The optimal evaporator and condenser mass fluxes of the DR-5 heat pump were observed to be in the ranges of 500–800 and 300–550 kg s–1 m–2, respectively. In the heating mode, as shown in Fig. 9(b), the heat pumps using R-32 and L-41a demonstrated the highest COP at evaporator mass fluxes of 250–350 and 350–550 kg s–1 m–2 and condenser mass fluxes of 400–600 and 450–650 kg s–1 m–2, respectively. Detailed results for the heat pumps using other alternative refrigerants are shown in Fig. 9. The optimal evaporator and condenser mass fluxes tend to increase with an increase in the liquid density of these alternative refrigerants. As the refrigerant density is increased, the mass flux should be increased to maintain a constant refrigerant velocity. Therefore, the performance of heat pumps employing these alternative refrigerants could be improved by controlling the mass fluxes in the evaporator and condenser. In addition, although the heat exchangers were optimized separately for the cooling and heating modes, the heat pump could be optimized by setting the optimal mass fluxes for each heat exchanger within the overlapping range for the cooling and heating modes. Fig. 10 shows the normalized EERs (COPs) of heat pumps with the optimized heat exchanger designs over that of the R-410A heat pump. In the cooling mode, the normalized EERs of the heat pumps using R-32, R-446A, DR-5, and L-41a were estimated at 105.8%, 93.9%, 101.6%, and 98.8%, respectively. In the heating mode, the normalized COPs of the heat

pumps using R-32, R-446A, DR-5, and L-41a were estimated at 106.2%, 103.4%, 105.1%, and 106.0%, respectively. Therefore, the R-32 heat pump with the optimized heat exchanger design exhibited the highest EER and COP in the cooling and heating modes, respectively. In addition, the DR-5 heat pump with the optimized heat exchanger design exhibited a higher EER and COP than the R-410A heat pump. As previously discussed, the R-446A heat pump demonstrated a higher compression ratio than the heat pumps that employ other alternative refrigerants due to its lower saturation pressure at a given temperature. Therefore, this heat pump showed the lowest EER and COP due to its higher compressor power consumption. With the application of the optimized heat exchanger designs in these heat pumps, the average performance improvements (EER and COP) over the drop-in performance were estimated at 6.0%, 3.4%, and 2.4% for the R-32, DR-5, and L-41a heat pumps. Accordingly, the R-32 heat pump demonstrated the highest performance improvement with the optimized heat exchanger designs. By contrast, the R-446A heat pump with the optimized heat exchanger design showed a 1.3% reduction in average performance. Table 6 lists the detailed performance data corresponding to the simulations performed.

4. Conclusion A simulation model for heat pumps with low global warming potential (GWP) refrigerants was developed and validated by comparing the predictions and measured data for the R-410A and R-32 heat pumps. Based on the drop-in simulations conducted, the normalized energy efficiency ratios (EERs) of the heat pumps using R-32 and L-41a over that of the R-410A heat pump were observed to be 104.9% and 98.5%, whereas the normalized coefficient of performances (COPs) of the heat pumps using R-32 and L-41a were 102.7% and 100.8%, respectively. The tube diameter and number of heat exchanger paths were optimized for each alternative refrigerant to maximize the EER (or COP) using the genetic algorithm. As the tube

diameters in the evaporator and condenser were decreased with the specified number of heat exchanger paths, the EERs and COPs of the heat pumps using the alternative refrigerants increased because of the higher heat transfer coefficient. Accordingly, the optimal evaporator and condenser tube diameters were determined to be 5 mm. As the number of heat exchanger paths were varied with a given tube diameter, the EERs and COPs peaked at certain values. For the R-32 heat pump in the cooling mode, the optimal numbers of evaporator and condenser paths were determined to be 9 and 12, respectively. Similarly, for the R-32 heat pump in the heating mode, the optimal number of evaporator paths was determined to be 12. The performance of the optimized heat pump using the low GWP refrigerants was then compared with that of the R-410A heat pump. The average performance improvements of the heat pumps with optimized heat exchanger designs over that of the R-410A heat pump were estimated at 6.0%, 3.4%, and 2.4% for the R-32, DR-5, and L-41a heat pumps. Thus, the R-32 heat pump with the optimized heat exchanger design exhibited a higher performance improvement than that of the conventional R-410A heat pump.

Acknowledgments This work was supported by a grant from the Energy Technology Development Program (No. 20172020108580) of the Korea Institute of Energy Technology Evaluation and Planning (KETEP) funded by the Ministry of Trade, Industry & Energy, Republic of Korea.

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Figure captions Fig. 1.

Schematic of the experimental setup.

Fig. 2.

Flow of the heat pump simulation model.

Fig. 3.

Validation results of the heat pump simulation model.

Fig. 4.

Normalized drop-in performance of each alternative refrigerant over that of R-410A in (a) cooling and (b) heating modes.

Fig. 5.

Normalized EER according to (a) tube diameter and (b) number of paths in the evaporator.

Fig. 6.

Normalized EER according to (a) tube diameter and (b) number of paths in the condenser.

Fig. 7.

Normalized COP according to (a) tube diameter and (b) number of paths in the evaporator.

Fig. 8.

Normalized COP according to (a) tube diameter and (b) number of paths in the condenser.

Fig. 9.

Normalized EER (or COP) according to mass fluxes in the evaporator and condenser in (a) cooling and (b) heating modes.

Fig. 10. Normalized EER (or COP) of heat pumps with optimized heat exchanger designs for each refrigerant in (a) cooling and (b) heating modes.

Outdoor Chamber 800 Hz

EEV

Outdoor heat exchanger M

P

T

P

T

System controller

Compressor

Data acquisition system

Data logger

4-way valve

P T M

P T

Indoor heat exchanger Cooling Heating Powermeter

Indoor Chamber

Fig. 1. Schematic of the experimental setup.

Input data

Compressor Adjust condensing temperature Adjust evaporating temperature

Condenser

Subcooling

Evaporator

Adjust compressor frequency Change target subcooling

Superheat

Capacity

Efficiency

Stop

Fig. 2. Flow of the heat pump simulation model.

20

8

(a)

1%

10

Predicted W (kW)

Predicted Q (kW)

R-410A R-32

6

15

-1%

TID (C)

5

0

(b)

R-410A R-32

: 20, 27 TOD (C) : -157, 2935

0

5

10

15

Measured Q (kW)

20

10%

4

TID (C)

2

0

-10%

: 20, 27 TOD (C) : -157, 2935

0

2

4

6

Measured W (kW)

Fig. 3. Validation results of the heat pump simulation model.

8

EERnormal (%)

110 105 100 95 90 110

COPnormal (%)

(a) Cooling (T : 27 C, T : 35 C) ID OD

105

R-410A

R-32

R-446A

DR-5

L-41a

DR-5

L-41a

(b) Heating (T : 20 C, T : 7 C) ID OD

100 95 90

R-410A

R-32

R-446A

Fig. 4. Normalized drop-in performance of each alternative refrigerant over that of R-410A in (a) cooling and (b) heating modes.

120

120 (a) Cooling 115 Dcond : 7 mm Ncond : 8

Nevap : 12

105 100 95

Devap : 5 mm

105 100 95 90

90 85

R-32 R-446A DR-5 L-41a

Ncond : 8

110

EERnormal (%)

EERnormal (%)

110

(b) Cooling Dcond : 7 mm 115

R-32 R-446A DR-5 L-41a

5

7 Devap (mm)

9

85

3

6

9

12

Nevap

Fig. 5. Normalized EER according to (a) tube diameter and (b) number of paths in the evaporator.

120

120 115

(a) Cooling Ncond : 8 Devap : 5 mm

Nevap : 12

105 100 95

(b) Cooling Dcond : 7 mm

R-32 R-446A DR-5 L-41a

Devap : 5 mm

110

Nevap : 12

105 100 95 90

90 85

115

EERnormal (%)

EERnormal (%)

110

R-32 R-446A DR-5 L-41a

5

7 Dcond (mm)

9

85

4

8

12

16

Ncond

Fig. 6. Normalized EER according to (a) tube diameter and (b) number of paths in the condenser.

120 115

120 (a) Heating Dcond : 5 mm Ncond : 12

Nevap : 16

105 100 95 90

115

(b) Heating Dcond : 5 mm

R-32 R-446A DR-5 L-41a

Ncond : 12

Devap : 7 mm

110

COPnormal (%)

COPnormal (%)

110

R-32 R-446A DR-5 L-41a

105 100 95

5

7 Devap (mm)

9

90

4

8

12

16

Nevap

Fig. 7. Normalized COP according to (a) tube diameter and (b) number of paths in the evaporator.

120 115

120 (a) Heating Ncond : 12 Devap : 7 mm

Nevap : 16

105 100 95 90

115

(b) Heating Dcond : 5 mm

R-32 R-446A DR-5 L-41a

Devap : 7 mm

Nevap : 16

110

COPnormal (%)

COPnormal (%)

110

R-32 R-446A DR-5 L-41a

105 100 95

5

7 Dcond (mm)

9

90

3

6

9

12

Ncond

Fig. 8. Normalized COP according to (a) tube diameter and (b) number of paths in the condenser.

Fig. 9. Normalized EER (or COP) according to the mass fluxes in the evaporator and condenser in (a) cooling and (b) heating modes.

EERnormal (%)

115 110 105 100 95 90 115

COPnormal (%)

(a) Cooling (TID : 27 C, TOD : 35 C)

110

R-32

R-446A

DR-5

L-41a

(b) Heating (TID : 20 C, TOD : 7 C)

105 100 95

R-32

R-446A

DR-5

L-41a

Fig. 10. Normalized EER (or COP) of heat pumps with optimized heat exchanger designs for each refrigerant in (a) cooling and (b) heating modes.

Table captions Table 1 Specifications of the heat pump components. Table 2 Range of variables for the genetic algorithm. Table 3 Basic properties of R-410A and possible alternative refrigerants. Table 4 Referred empirical formulas for heat transfer and pressure drop from existing studies. Table 5 Drop-in simulation results for the heat pump designed for R-410A. Table 6 Simulation results for the heat pump with optimized heat exchanger designs.

Table 1 Specifications of the heat pump components. Compressor

Expansion device

Indoor unit

Type

Inverter scroll

Displacement volume (cc)

31.6

Frequency (Hz)

10–135

Type

EEV

Orifice diameter (mm)

2.4, 3.0

Pulse range (step)

0–500

Fan type

Centrifugal

Row Column Tube

Fin

Outdoor unit

Number

3

Pitch (mm)

12

Number

12

Pitch (mm)

21

Diameter (mm)

5

Thickness (mm)

0.22

Length (mm)

2088

Material

Copper

Type

Slit

Material

Aluminum

Thickness (mm)

0.1

Pitch (mm)

1.21

Number of paths

12

Fan type

Axial

Row Column Tube

Fin

Number

2

Pitch (mm)

18.2

Number

64

Pitch (mm)

21

Diameter (mm)

7

Thickness (mm)

0.22

Length (mm)

950

Material

Copper

Type

Louver

Material

Aluminum

Thickness (mm)

0.1

Pitch (mm)

1.81

Number of paths (Cooling/Heating)

8/16

Table 2 Range of variables for the genetic algorithm. Variable

Lower bound

Upper bound

Tube diameter, D (mm)

5

9

Number of the paths, N

3

12

Tube diameter, D (mm)

5

9

Number of the paths, N

4

16

Compressor frequency, Hz (s–1)

15

150

Subcooling, Tsc (℃)

3

25

Indoor heat exchanger

Outdoor heat exchanger

Table 3 Basic properties of R-410A and possible alternative refrigerants.

Refrigerant

Composition (wt%)

Tcrit (℃)

Pcrit (kPa)

Density at 25 ℃ (kg m–3)

Latent heat (kJ kg–1)

Liquid

Vapor

Te at 10 ℃

Tc at 45 ℃

Safety GWP class

R-410A

R-32/R-125 = 50/50

4926

72.1

1058

66.0

208.5

148.0

A1

2088

R-32

-

5783

78.1

961

47.3

298.9

224.0

A2L

675

R-446A

R-32/R-1234ze(E)/R-600 = 68/29/3

5727

86.0

1010

43.6

259.6

202.0

A2L

460

DR-5

R-1234yf/R-32 = 27.5/72.5

5219

76.8

981

52.5

253.0

188.8

A2L

490

L-41a

R-1234yf/R-1234ze(E)/R-32 = 15/12/73 5514

80.9

994

49.2

257.7

195.5

A2L

490

Table 4 Referred empirical formulas for heat transfer and pressure drop from existing studies. Fluid

Classification Heat transfer correlation n2

Air side

Slit fin [29]

Pressure drop correlation n3

n2

 F  S  j  1.0691Re  s   s  Rn4  Dc   Sh 

n3

n2

Louver fin [28]

n2

n3

n4

 F  P S  f  1.201Re  s   t   s  (R)n5 (Sn )n6  Dc   Pl   Sh  n1

n1

n4

F  L  P j  1.1373Re n1  p   h   l  ( R)0.3545    Pl   Lp   Pt 

n3

n4

L   h   Lp   R n 5 (ln(Re)  4.0) 1.093

 Fp   D  f  0.06393Re    h   Dc   Dc  n1

Refrigerant side

L u 2 D 2

Single-phase [33]

Nu  0.023Re0.8 Prn

P  f

Two-phase evaporation [32,34]

htp  Ehl  Shpool

P  PF

2l2 fl G 2 L l D

Two-phase condensation [31,34]

Nu  0.05Re Pr Rx (Bo Fr)

P  PF

2l2 fl G 2 L l D

0.8

1/3

n

t

Table 5 Drop-in simulation results for the heat pump designed for R-410A.

CR

Td (℃)

Mass flow rate (kg h–1)

2778.5

2.89

76.3

328.7

762.5

2828.0

3.71

86.8

280.5

4222

979.0

2807.9

2.87

92.3

213.6

16722

4912

775.4

2880.7

3.72

110.5

183.7

Cooling

14367

4698

808.1

2541.0

3.14

89.3

257.9

Heating

16637

5142

660.5

2626.1

3.98

102.6

216.5

Cooling

14504

4375

924.8

2686.4

2.90

83.3

265.9

Heating

16583

5162

749.5

2950.9

3.94

102.4

222.1

Cooling

14424

4471

886.0

2644.7

2.99

86.4

259.3

Heating

16567

4957

737.6

2762.5

3.75

99.7

221.1

Refrigerant Mode

Capacity (W)

Power Pe (W) (kPa)

Pc (kPa)

R-410A

Cooling

14460

4416

960.4

Heating

16645

5022

Cooling

14509

Heating

R-32 R-446A DR-5 L-41a

Table 6 Simulation results for the heat pump with optimized heat exchanger designs.

CR

Td (℃)

Mass flow rate (kg h–1)

2834.9

2.91

94.3

217.0

775.5

2839.2

3.66

108.9

183.3

4691

805.4

2538.0

3.15

89.3

256.7

16773

4893

657.6

2449.1

3.72

96.5

220.1

Cooling

14408

4332

930.3

2686.4

2.89

83.5

265.3

Heating

16592

4761

752.3

2674.4

3.55

92.9

227.4

Cooling

14424

4459

892.9

2659.4

2.98

86.9

259.7

Heating

16565

4714

735.1

2618.0

3.56

94.9

221.0

Refrigerant Mode

Capacity (W)

Power Pe (W) (kPa)

Pc (kPa)

R-32

Cooling

14407

4158

973.3

Heating

16605

4717

Cooling

14419

Heating

R-446A DR-5 L-41a

Highlights  Heat pump simulation using low global warming potential refrigerants is performed. 

Optimum heat exchanger designs are determined for each alternative refrigerant.



Optimum mass flux range for achieving the maximum performance is proposed.



R-32 heat pump with the optimized design shows a performance improvement of 6%.