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Seasonal performance assessment of refrigerants with low GWP as substitutes for R410A in heat pump air conditioning devices
Applied Thermal Engineering 125 (2017) 401–411
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Applied Thermal Engineering journal homepage: www.elsevier.com/locate/apthermeng
Research Paper
Seasonal performance assessment of refrigerants with low GWP as substitutes for R410A in heat pump air conditioning devices Atilla Gencer Deveciog˘lu Dicle University, Department of Mechanical Engineering, 21280 Diyarbakır, Turkey
h i g h l i g h t s Four refrigerants alternatives to R410A were compared. R452B was seen to be a suitable alternative R410A at heating mode. The SEER values of the refrigerants were very close to each other. SEER value for all of the alternative refrigerants was found higher compared to R410A. R446A is more suitable for air conditioners performing cooling only.
a r t i c l e
i n f o
Article history: Received 12 December 2016 Revised 26 May 2017 Accepted 3 July 2017 Available online 5 July 2017 Keywords: SEER SCOP R410A GWP
a b s t r a c t In this study, the effect of the utilization of refrigerants with a low Global Warming Potential (GWP) instead of R410A in air conditioners on seasonal efficiency was evaluated theoretically. In the study, four different alternative refrigerants were used namely, R446A, R447A, R452B and R454B. The GWP values of all of the studied refrigerants were lower than 750. The calculations were made according to the EN 14825 standard. Average seasonal conditions were applied for heating. The results showed that R446A had the lowest heating capacity and electric consumption. R446A required electric energy consumption for extra heating due to its low heating capacity and therefore it had the lowest seasonal coefficient of performance (SCOP) value. The seasonal energy efficiency ratio (SEER) values of the refrigerants were very close to each other and their average was 6.56. R452B had the most suitable with 3.92 SCOP value among the alternative refrigerants. Ó 2017 Elsevier Ltd. All rights reserved.
1. Introduction The European Union (EU) scheduled the utilization of refrigerants with GWP value in air conditioning devices by a regulation [1]. Again, in the EU countries, a new energy label application to be used in air conditioning devices was initiated in the scope of the Regulation of Energy Labeling of Air Conditioner in 2013. In pursuant of this regulation, there must be inscriptions on labels showing the seasonal energy efficiency with regard to both cooling and heating. Performance measurement methods have changed for air conditioners lower than 12 kW and the seasonal efficiency term has been defined by Eco-Design directive that determines minimum efficiency requirements for energy-utilizing products [2]. Three regions were determined for the heating season as colder, average and warmer. In pursuant of 206/2012 EU regulation [3], air conditioning systems must have the minimum energy efficiency
E-mail address: [email protected] http://dx.doi.org/10.1016/j.applthermaleng.2017.07.034 1359-4311/Ó 2017 Elsevier Ltd. All rights reserved.
values shown in Table 1 starting from January 1, 2014. Singleduct and double-duct air conditioners and the devices other than these are classified based on their cooling capacity and the GWP values of the used refrigerants. R410 refrigerant with 2088 GWP value is used frequently in many air conditioning devices [4]. Due to this high GWP value, refrigerants that can be used instead of R410 with low GWP value and high energy parameter value have been searched. In a study conducted to investigate R410A was used instead of R22, the coefficient of cooling performance (COP) values of R22 and R410A were found to be very close to each other at different condenser temperatures [5]. Although not a long time has passed, alternative refrigerants are being investigated due to high GWP value of R410. A study was conducted by using a R32/R290 mixture instead of R410 [6]. The cooling and heating values of R32/R290 mixture were higher than those of R410A and the COP and energy efficiency ratio (EER) values of R410 were higher than those of R32/ R290 mixture. R32 and R410A in domestic air conditioning device
A.G. Deveciog˘lu / Applied Thermal Engineering 125 (2017) 401–411
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Nomenclature COP EER GWP SCOP SEER h h0 _ m P Q T Wel Vdisp m_ suc
Subscripts comp compressor c cooling e evaporator in inlet out outlet L low H high h heating d design bi bivalent ol operating limit a ambient is isentropic vol volumetric on operating state hd heat demand hp heat pump
coefficient of performance energy efficiency ratio global warming potential seasonal coefficient of performance seasonal energy efficiency ratio enthalpy, kJ kg1 enthalpy for isentropic case, kJ kg1 mass flow rate, kg s1 pressure, kPa heat capacity, kW temperature, °C electrical power input to the system, kW displacement specific volume m3 kg1 efficiency number of hours in a year rpm
g
L n_
Table 1 Requirements for minimum energy efficiency. Air conditioners, except double and single duct air conditioners
If If If If a
GWP GWP GWP GWP
of of of of
refrigerant > 150 for air conditioner capacity < 6 kW refrigerant 150 for air conditioner capacity < 6 kW refrigerant > 150 for air conditioner capacity 6–12 kW refrigerant 150 for air conditioner capacity 6–12 kW
Double duct air conditioners
Single duct air conditioners
SEER
SCOPa
EERrated
COPrated
EERrated
COPrated
4.60 4.14 4.30 3.87
3.80 3.42 3.80 3.42
2.60 2.34 2.60 2.34
2.60 2.34 2.60 2.34
2.60 2.34 2.60 2.34
2.04 1.84 2.04 1.84
Heating season: Average.
were compared in both the heating and cooling mode [7]. The cooling and heating capacities and COP values of R32 were reported to be better than those of R410A. R410A and R32, D2Y60 and L41A refrigerants in a 10 kW split air conditioning system were compared for both heating and cooling modes [8]. The GWP values of each of the three alternatives were lower than 750. R32 had the highest cooling capacity. The heating capacity values of R410A and L41A were very close to one another. The refrigerants R410A and R32 and R446A of domestic heat pump were compared under partial loads with the EN 14825 calculation method [4]. There was a 2.1% increase in the EER value when R32 was used and a 1.4% increase when R446A was used. In another study, the air to water heat pump’s seasonal performances, which were designed in pursuant of the EN 14825 were evaluated [9]. In the study, heat pumps with single compressor, multicompressor and variable speed compressor were compared. The theoretical studies comparing the performances of refrigerants with low GWP value in various vapor compression refrigeration systems [10,11] may be beneficial to have an idea prior to the experimental study. In a theoretical study comparing the utilization of refrigerants with low GWP value instead of R410A [12], the COP values of the alternative refrigerants were found to be higher than that of R410A. Response surface method (RSM) was used for mathematical model of the system. RSM is a mathematical modelling and statistical evaluation method used for optimization of input parameters (independent variable) of a system [13].
Related to this subject, Panato et al. [14] have performed analysis through RSM using different refrigerants as drop-in application. They expressed that experimental data were compatible with the results obtained with RSM. All analyzed values of (R2), for all fluids were greater than 0.80. In order to determine optimum operating conditions and the effect of fixed parameters on the system working with R22 and R404A, Çomaklı et al. [15] used experimental design method proposed by Taguchi. It was indicated that COP was mostly affected by the air temperature at condenser inlet. In this study, the energy parameters of refrigerants with low GWP that could be used instead of R410A were calculated theoretically for the average heating and cooling season based on the test point temperatures. In the study, the calculation method provided in the EN14825 standard was used. The refrigerants with low GWP value that could be used instead of R410A were compared based on their SEER and SCOP values. 2. The EN 14825 calculation method A calculation method has been determined for the calculation of the SCOP and SEER values which are mandatory to be placed in the labels of air conditioning devices per the Eco-design Directives. In the calculations made with this method, the reference design conditions shown in Table 2 are considered to maintain these values in air conditioning devices other than the single-duct and doubleduct air conditioning devices [3].
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Table 2 The test and design temperatures for heating and cooling modes and partial load ratios. Point
Cooling
A B C D Tdesign (°C) Tbi (°C) TOL (°C)
Heating (Warmer)
Heating (Average)
Ttest (°C)
Part loads
Ttest (°C)
Part loads
Ttest (°C)
Part loads
Ttest (°C)
Part loads
– 35 30 25 20 35 – –
– 100% 74% 47% 21%
– – 2 7 12 2 7 2
– – 100% 64% 29%
– 7 2 7 12 10 2 7
– 88% 54% 35% 15%
15 7 2 7 12 22 7 15
82% 61% 37% 24% 11%
In the previously used calculation method, nominal efficiency was calculated at a specific ambient temperature (35 °C for cooling and 7 °C for heating), at full load and without considering the energy consumed by auxiliary components. The observance frequency of a 35 °C temperature in real conditions is low. The ambient temperatures during the summer months, when a cooling necessity occurs, range between 16 and 40 °C and change based on geographical location and climatic conditions. Moreover, there is less cooling need at low ambient temperatures. The EN 14825 standard has been developed to calculation the efficiencies of air conditioning devices more realistically [16]. The seasonal data of Strasburg city have been accepted as the basis and the design temperature has been selected as 35 °C for the SEER calculation in the EN 14825 standard [3]. Four test points have been defined in the standard. These are 20, 25, 30, and 35 °C. The device can operate under partial loads at different ambient temperatures since the cooling necessity is different. Therefore, air conditioning devices are tested under different partial loads (100%, 74%, 47% and 21%). The SEER value is calculated based on the EER values calculated at these four test points. Three different climatic zones have been defined for the SCOP calculation in the EN 14825 standard. These are the average climatic zone, warmer climatic zone and colder climatic zone. Strasburg city has been taken as the basis for the average climatic zone, Athens has been taken as the basis for the warmer climatic zone and Helsinki has been taken as the basis for the colder climatic zone. The test and design temperatures and partial load ratios used in the SCOP and SEER calculations are shown in Table 2. In addition to these test and design temperatures, bivalent temperature (Tbi) value has been defined as a temperature when the heat demand of a heat pump starts to be fulfilled completely. The operation limit temperature (TOL) has been defined as the lowest temperature when the heat pump can operate. The observance time of the ambient temperatures in these climatic zones during a year, in other words their frequencies are shown in Fig. 1 [2]. These data are used in the SCOP and SEER calculations. The frequency of temperature observance times, which is one of the starting points of seasonal calculation, has concentrated in the warm zone between +2 and +8 °C as shown in Fig. 1. Similarly, the cooling season need has concentrated between +20 and +25 °C. 2.1. The determination of heat demand used for the calculation of the SCOP The heat demand desired for various test temperatures is calculated by Eq. (1) based on the design temperatures at the climatic zone that is to be calculated. The design temperatures shown in Table 2 [16]. The heat demand curve is formed with this expression.
Pdesignh ¼
Heating (Colder)
T test 16 T designh 16
ð1Þ
The percentages of heat demand at various test temperatures for different climatic zones (warmer, average, colder) are shown in Fig. 2. It must be considered that these percentages (partial load) are not the partial load ratios of a heat pump. The heat demand Pdesignh value and the SCOP are placed on an air conditioning label together. Manufacturers use Pdesignh value to describe the size of a heat pump. Hence, consumers can select the correct device agreeable with their climate condition. 2.2. Calculation of the SCOP value In the study, theoretically calculated values for test temperatures were used and the other intermediate values were determined by interpolation as specified in the EN 14825. Extrapolation can be used within a narrow range. It is necessary to find out how much electric energy is consumed at what temperature in order to determine the compressor electric consumption used in the COP calculation. It is necessary to know what temperature is seen with what frequency when doing this calculation. The observance frequency of temperatures is found for the regions by using the graph shown in Fig. 1. In this study, the average temperature season was used for all refrigerants. Electric consumption, the COP and heating capacity values found by the theoretical study were determined for the four temperature points at the average heating season. Other temperature values were found by interpolation based on these values. Subsequently, partial loads at the test temperatures were determined based on the design temperature (Tdesignh = 10 °C in average heating season) of a 7 kW capacity air conditioning device. Heat demand at the test temperature was determined by multiplying heat demand at the design temperature by partial loads. Total heat demand was found by multiplying heat demand with the observance hour number of the test temperature. Total electric consumption was found by multiplying electric consumption per hour with the observance hour number of the test temperature. The ratio of total heat demand for all test temperatures to total electric consumption gives the seasonal cooling performance coefficient (SCOPon) at the operating state. The SCOP value is calculated in pursuant of the EN14825.
SCOP ¼
Qh Qh SCOPon
þ Waux
ð2Þ
Waux value is the electric energy consumed by the device when it is (off), at standby (sb), when the crankcase heater is on (ch) and the thermostat is off (to) and is calculated as follows [16].
W aux ¼ Lto wto þ Lsb wsb þ Lch wch þ Loff woff
ð3Þ
Here, L value indicates annual operation hour number of a heat pump and w indicates energy consumption.
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Fig. 1. The observance number of ambient temperatures used in calculations for climatic zones.
3. Comparison of refrigerants’ characteristics
120% Average Warmer
100%
Partial load
Colder 80% 60% 40% 20% 0% -18
-8
2
12
22
Ttest (°C) Fig. 2. Determining the heating demand curve at Pdesignh, different climatic conditions.
The missing heat demand amount between the heat demand and the heat pump capacity will be met by a resistance heater. Electric energy consumed by the resistance heater is added to total electric consumption. 2.3. The SEER value calculation In the SEER calculation, electric consumption, the COP and cooling capacity values found by the theoretical study for the four temperature points at the cooling season were determined. Other test temperature values were determined by interpolation based on these values. By multiplying hourly cooling demand with the test temperature observance hour number, total cooling demand was found for that test temperature. Total electric consumption was found by multiplying hourly electric consumption with the test temperature observance hour number. The ratio of total cooling demand for all test temperatures to total electric consumption gives the seasonal energy efficiency ratio at the operating state (SEERon). The SEER value is calculated in pursuant of the EN 14825 as shown in Eq. (4) by dividing annual cooling demand to the sum of power draw for annual cooling and power draw at auxiliary positions [16].
SEER ¼
Qc SEERon
Qc þ Waux
ð4Þ
In this study, refrigerants with low GWP value and better energy parameters were investigated instead of R410A refrigerant used in air conditioning systems especially. R410A refrigerant is used in numerous systems including mono type split air conditioners, VRF systems with hundreds of indoor units, and chillers. Some properties of the refrigerants used in the study are shown in Table 3. Critical temperature values of all refrigerants are close to each other. Therefore, they may be used in a system designed for R410. The GWP values of all alternative refrigerants are lower than 750 as shown in Table 3. The saturation curves of R447A, R452B and R454B refrigerants are very close to each other especially as shown in Fig. 3. R446A (mixture ingredient R32/R600/R1234ze(E); mixture ratio, 29:3:68) HFC/HC/HFO mixture, R447A (mixture ingredient, R32/R125/R1234ze(E); mixture ratio, 68:3.5:28.5), R452B (mixture ingredient, R32/R125/R1234yf; mixture ratio, 67:7:26) and R454B (mixture ingredient, R32/R1234yf; mixture ratio, 68.9:31.1) HFC/ HFO mixture are the refrigerants used instead of R410A (mixture ingredient, R32/R125; mixture ratio, 50:50).
4. Theoretical analysis Energy parameters at the test temperature points were found by a theoretical study. The heating and cooling capacities of the investigated refrigerants were determined for an air conditioning device with a 7 kW cooling capacity. An air conditioning device with a heating feature consists of basic components including a compressor, condenser, expansion valve, evaporator, and fourway valve. During the heating process, the refrigerant’s direction is changed by a four-way valve and this way, the condenser acts like an evaporator (Fig. 4). Some assumptions were made for the system’s analysis; it was assumed that the system operated steady state conditions. Heat and pressure losses in the air conditioning system components and pipes were ignored. Energy consumption of the fan of the evaporator and condenser was ignored. There was no kinetic and potential energy terms, and the auxiliary components had no power consumption. The compressor volumetric efficiency, isentropic efficiency and displacement values were modified based on the test temperature points in the theoretical study to simulate the partial loads given in the EN 14825. Thus, it was supposed that the system operated like a variable speed. These values are shown in Table 4. The study differed with this aspect from the previous theoretical studies.
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Table 3 Some properties of the refrigerants used in the study [17].
Critical temperature (°C) Critical pressure (kPa) Liquid density (kg m3) (25 °C) Boiling point (101.3 kPa) GWP
R410A
R446A
R447A
R452B
R454B
71.3 4901 459 48.5 2088
92.8 4678 442 30.8 490
80.3 5269 421 45.6 580
81.9 5289 421 45.5 695
83.7 5334 418 44.6 465
Fig. 3. Pressure enthalpy change for the refrigerants.
Fig. 4. A schematic illustration of an air conditioner system cycle.
Table 4 Assumptions made for the theoretical analysis: volumetric efficiencies, isentropic efficiencies and displacement ratios according to the test temperatures. Ttest (°C)
gvol gis
Vdisp (m3h1)
10 0.75 0.60 6.20
7 0.80 0.70 6.20
2 0.85 0.85 3.00
7 0.90 0.90 1.70
Superheat and subcooling values were accepted as respectively as 8 and 5 °C during the heating process. Tev value was determined as constant and 10 °C lower than the test temperatures. The condensation temperature was accepted as constant 43 °C for all test temperatures. Superheat and subcooling values were accepted as 10 and 5 °C respectively for the cooling process. The condensation temperature values were accepted to be as constant and 10 °C
12 0.95 0.90 0.70
20 0.9 0.85 6.8
25 0.85 0.8 4.5
30 0.8 0.7 2.8
35 0.75 0.6 2.1
higher than the test temperatures. The evaporation temperature was accepted as constant 0 °C for all test temperatures. 4.1. Calculation of energy parameters Cooling capacity (Qe) of the cooling cycle is calculated as follows based on the above given information [18]:
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_ e;out he;in Þ Q e ¼ mðh
ð5Þ 1
_ mass flow rate (kg s ), he,out and he,in are the enthalpy valHere, m ues (kJ kg1) of the refrigerant at the outlet and inlet of the evaporator. Electric consumed by the compressor:
_ comp;out hcomp;in Þ W el ¼ mðh
ð6Þ
Here, Wel is the electric power (kW) consumed for the system’s operation and hcomp,out and hcomp,in are the enthalpy values (kJ kg1) of the refrigerant at the output and inlet of the compressor. Heat rejected by the condenser or heat given to the environment that is to be heated at the heating mode is found by the following expression:
Q h ¼ W el þ Q e
ð7Þ
Isentropic efficiency of the system is found by the following expression:
gis ¼
0
ðhcomp;out hcomp;in Þ ðhcomp;out hcomp;in Þ
Table 6 Design planning and calculation results. Calculation no.
Ta (°C)
gis
Vdisp
COP
1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17
1 1 1 1 1 1 1 1 0 0 0 0 0 0 0 0 0
1 1 1 1 0 0 0 0 1 1 1 1 0 0 0 0 0
0 0 0 0 1 1 1 1 1 1 1 1 0 0 0 0 0
6.29 3.71 4.44 2.62 5.36 3.17 5.36 3.17 4.76 3.36 4.76 3.36 4.06 4.06 4.06 4.06 4.06
ð8Þ
0
Here, hcomp;out is the enthalpy value corresponding to the constant entropy at the compressor outlet of the refrigerant during the compressing process. hcomp,out and hcomp,in are the enthalpies of the compressor corresponding to the discharge and suction temperature values. Volumetric efficiency value used in the study is determined by Eq. (9).
Table 7 The assumptions for analysis of the model. Ta (°C)
gis
Vdisp (m3h1)
Te (°C)
Superheat (K)
Subcooling (K)
15/45
0.4/0.90
2/8
5/+5
3/9
2/5
8 R454B
_ m
gv ol ¼ _ measured mtheoretical
ð9Þ
R410A
6
R446A
Here, theoretical mass flow rate is calculated as follows:
R447A
ð10Þ
Here, Vdisp is the compressor displacement and the values in Table 4 are accepted. n_ is the rotational speed (rpm) of the compressor and tsuc is the specific volume (m3 kg1) value of the refrigerant at the suction temperature of the compressor. Energy performance of a refrigeration system is defined by the coefficient of performance (COP). The COP is calculated as the ratio of the cooling capacity to the consumption of electrical energy per unit time by the compressor.
COP ¼
Qe W el
ð11Þ
5. Model development Mathematical model of COP based on response surface method: Ta,
gis and Vdisp expressions were used to determine COP which is one of the energy parameters. While COP is reduced with Ta, it is improved as gis increases. Note that gis and Vdisp have been added to the model for considering the effect of variation of partial loads on COP. Evaporation temperature, volumetric efficiency, superheat and subcooling values were assumed to be constants as in theoretical study. For the calculations, Ta, gis and Vdisp have been selected as variable parameters which are given in Table 5.
Table 5 Experimental parameters and their levels. Parameters
Level 1 (1)
Level 2 (0)
Level 3 (+1)
Ta (°C)
20 0.85 2.1
25 0.75 4.5
35 0.60 6.8
gis
3 1
Vdisp (m h
)
COP
_ theoretical m
V disp n_ ¼ tsuc 60
R452B
4
Predicted
2
0 15
20
25
30
35
40
Ta (°C) Fig. 5. The comparison of calculated results with the results obtained from model (predicted).
Theoretical calculations were conducted to investigate the effects of Ta, gis and Vdisp on COP. The calculations in this study were performed using Box-Behnken experimental design planning. The performance outputs through the calculations are presented in Table 6. For the obtained calculation results, statistical analyses were conducted using Design Expert 6.0 software and a suitable regression model was formed. The second order equation composed for COP is
COP ¼ 4:06 ð1:1 T a Þ ð0:72 gis Þ þ ð0:21 T a 2Þ þ ð0:19 T a gis Þ ð12Þ It is seen from the model that value of Vdisp is a variable which does not influence the result. The model gives very close results for the refrigerants studied considering the values given in Table 7. R2 term called as determination coefficient was determined as 0.9996 for COP. Comparison of the computed results and obtained results from mathematical model (predicted) for all refrigerants investigated in
A.G. Deveciog˘lu / Applied Thermal Engineering 125 (2017) 401–411 Table 8 Comparison of experimental data with present model.
407
2.0
COP
R410A R446A R447A R452B R454B
Ta = 35 (°C)
Oruç and Deveciog˘lu [19] R22 R417A R424A
3.77 3.03 3.54
3.13 2.40 2.52
Prediction of model R22 R417A R424A
3.49 3.16 3.81
2.86 2.48 2.49
Wel (kW)
1.6
Ta = 25 (°C)
1.2
Heating mode
0.8 0.4 0.0
6. Results and discussions
0
40
60
80
100
120
80
100
120
(a) 3.0 R410A R446A R447A R452B R454B
2.5 2.0
Cooling mode
1.5 1.0 0.5 0.0 0
In the study, theoretical analysis of air-conditioning device was conducted for different refrigerants considering both heating and cooling applications. Therefore, a difference was obtained compared to previous theoretical investigations generally on refrigeration systems. It can be said that partial load modelling in previous theoretical analyses is also different. In terms of SEER and SCOP values which are necessary for energy labelling of airconditioning devices, the suitable refrigerants with low-GWP were suggested. The variation of electric consumption by partial loads for heating and cooling modes of the refrigerants is shown in Fig. 6. A partial load decrease causes reduce of electric consumption in both modes. Electric consumption values of all refrigerants, except for R446A, are close to one another. R446A has the lowest electric consumption value in both modes. R410A has the highest electric consumption at full load (100%) in both modes. The variation of heating capacities at the ambient temperatures for various refrigerants in heating mode is shown in Fig. 7 a. An increase of ambient temperature decreases heat loss of heating space. Therefore, air conditioning device’s thermal capacity decreases as the ambient temperature decreases with partial load changes at the ambient temperature points. Thermal capacities of R454B and R452B were found as approximately 3–5% lower than that of R410A. Approximately there was a 50% capacity loss for R446A especially. An increase in ambient temperature will increase cooling of the space necessity. The variation of cooling capacities of the air conditioning device with the increase of ambient temperature is shown in Fig. 7b. A decrease in ambient temperature in the systems operating respect to partial load will reduce cooling capacity of the air conditioning device. Cooling capacities change as parallel to heating capacities. Cooling capacities of R454B and R452B refrigerants were determined as approximately 2–5% lower than that of
20
Partial load (%)
Wel (kW)
the study can be seen in Fig. 5. Obviously, predicted results are very close to the calculated results. The validation of proposed model was verified considering experimental study by Oruç and Deveciog˘lu [19]. The ambient temperature values were 25 and 35 °C while gis was 0.50–062 in their investigation. The system was continuously operating at full capacity. The predicted results and experimental results are presented in Table 8. The results are close to each other, however, the difference is some higher for R417A. These differences may have been developed due to the fact that gis and Vdisp values, which are input parameters of the model assumed to operate under partial loads, are different. The obtained model presents compatible results. Therefore a new model has been proposed in the literature for estimating and optimizing COP value of refrigeration systems through suitable input parameters.
20
40
60
Partial load (%)
(b) Fig. 6. The variation of electric consumption by partial loads (a) heating mode and (b) cooling mode.
R410A. Cooling capacity of R446 is approximately 45% lower than that of R410A. The variation of mass flow rate value of the refrigerant that needs to circulate in the system when the air conditioning device is operating by ambient temperatures for various refrigerants is shown in Fig. 8. Mass flow rate values decrease with the decrease of ambient temperatures in cooling mode depending on the volumetric efficiency and displacement values. R446A has the lowest mass flow rate value. Mass flow rate values of R454B and R452B refrigerants are very close to each another. The decrease of mass flow rate value by the ambient temperatures in heating mode is shown in Fig. 8a. It is understood that an air conditioning device cannot achieve the necessary heating even when it operates at full capacity due to a low temperature value (10 °C). When the ambient temperature is 7 °C, mass flow rate value increases with the increase of suction pressure. However, there was a decrease seen in mass flow rate value at partial load as parallel to an increase in the ambient temperature. The condensation temperature was accepted as constant in heating mode. The variation of low pressure (suction pressure) PL value based on partial loads is shown in Fig. 9. An increase in the evaporation temperature causes an increase in low pressure value since the decrease of partial load depends on the ambient temperature. PL value of R410A is approximately 50% higher than that of R446A.
A.G. Deveciog˘lu / Applied Thermal Engineering 125 (2017) 401–411
408
6
900
5
R452B R454B
750
4
R447A R446A
600
3 2
R410A
Heating mode
R410A
R446A R447A
PL (kPa)
Qh (kW)
Heating mode
R452B R454B
450 300
1
150
0 -10
-7
2
7
0
12
0
Ta ( °C)
20
40
60
80
100
120
Partial load (%)
7
Cooling mode
Qe (kW)
Fig. 9. The variation of low pressure with partial load.
R410A
6
R452B R454B
5
R447A
R410A R446A R447A R452B R454B
2700
R446A
4 3
PH (kPa)
2200 2 1 35
30
25
20
Ta ( °C)
Cooling mode
1700
1200
Fig. 7. The variation of heating and cooling capacity by the test temperatures.
700
0.025
0
Heating mode
m (kgs-1)
0.020
40
60
80
100
120
Partial load (%)
R410A R446A R447A R452B R454B
0.015
20
Fig. 10. The variation of high pressure with partial load.
0.010
6.0 0.005
5.5 5.0
0.000 -10
-7
2
7
12
4.5
COP
Ta (°C)
(a)
R4446A R4447A R454B R452B R410A
4.0 3.5
Cooling mode
0.04
R410A R446A R447A R452B R454B
m (kgs-1)
0.03
0.02
3.0 2.5 2.0 -10
-7
2
7
12
Ta ( °C) Fig. 11. The variation of the COP value by the test point temperatures.
0.01
0.00 35
30
25
20
Ta ( °C)
(b) Fig. 8. The variation of mass flow rate value with the test point temperatures (a) heating mode and (b) cooling mode.
The evaporation temperature in cooling mode was accepted as constant. The variation of high pressure value PH by partial loads is shown in Fig. 10. A decrease in the condenser temperature causes a decrease in high pressure value since the decrease of partial load depends on a decrease in the ambient temperature. High pressure values of R454B and R452B refrigerants are very close to each other. Moreover, it is seen in the graph that R410A has the highest PH value and R446A has the lowest PH value.
A.G. Deveciog˘lu / Applied Thermal Engineering 125 (2017) 401–411
The variation of the COP values based on the ambient temperatures for the refrigerants in heating mode is shown in Fig. 11. The COP values of R452B, R454B and R410A at 10 °C are equal to one another. The COP value of R410A was found to be approximately 1% higher than those of R452B and R454B as the ambient temperatures increased. The COP values increased 35% in the average as the ambient temperatures of all these refrigerants increased from 7 to 7. An air conditioning device will take more heat energy from the environment as the ambient temperature increases. It will consume less electric during this process. Therefore, an increase in the ambient temperatures caused the COP values to increase. Since the COP values of R452B and R454B were close to that of R410A, they may be suitable refrigerants for heating. The variation of the EER values by the ambient temperatures in cooling mode is shown in Fig. 12. The EER values of all alternative refrigerants were found to be higher than that of R410A. R446A was determined as the refrigerant with the highest EER value as differing from heating mode. A low electric consumption caused determining a high EER value for R446A. A low electric consumption was due to low evaporator pressure and low discharge pressure. A 2–7% difference was determined between the EER values of the refrigerants. All of the refrigerants can be used instead of R410A owing to these close values. Annual electric consumptions of the refrigerants are shown in Fig. 13. Annual electric consumption for the average heating season where the study was conducted was approximately 3 folds higher than that of the cooling season. More electric is consumed for heating especially at low temperatures in comparison to the cooling mode. Furthermore, when heat loss cannot be met completely, an additional resistance heater steps in and increases electric consumption. R446A consumes approximately 55% more electric energy in comparison to other refrigerants in the heating mode. The reason behind this is that R446A cannot meet heat demand of the volume that is to be heated only by a heat pump in the
7.0 R410A R452B R447A R454B R446A
6.5 6.0 5.5
EER
5.0 4.5 4.0 3.5 3.0 2.5 2.0
35
30
25
20
Ta ( °C) Fig. 12. The variation of the EER value by the test point temperatures.
7000 6000
(kWh)
Annual consumption of electricity
8000
R410A
5000
R452B
4000
R454B 3000
R447A
2000
R446A
1000 0 Heating
409
Cooling
Fig. 13. Annual electric consumption in the heating and cooling modes.
1200
1200 R446A
Qhp
1000
Qhp
R452B
Whp
Whp
1000
Qhd
800
Energy (kWh)
Energy (kWh)
Qhd Tbi
600 400 200
800
Tbi
600 400 200
0
0 -12
-8
-4
0
4
8
12
16
20
-12
-8
-4
0
4
8
12
1200
20
1200 R447A
Qhp
1000
Qhp
R454B 1000
Whp
Whp Qhd
800
Energy (kWh)
Qhd
Energy (kWh)
16
Ta (°C)
Ta (°C)
Tbi
600 400 200
800
Tbi
600 400 200
0
0 -12
-8
-4
0
4
Ta (°C)
8
12
16
20
-12
-8
-4
0
4
8
12
16
20
Ta (°C)
Fig. 14. The variation of heat pump capacity, total heat demand, total electric consumption and Tbi values of the refrigerants by the ambient temperature.
A.G. Deveciog˘lu / Applied Thermal Engineering 125 (2017) 401–411
410
7.0 6.0 5.0 R410A 4.0
R452B R454B
3.0
R447A 2.0
R446A
1.0 0.0 SCOP
SEER
Fig. 15. The variation of the SCOP and SEER values of the refrigerants.
heating mode. The difference between the heat pump’s capacity and the heat demand of the volume that is to be heated is met by a resistance heater. Electric energy consumed by an air conditioning device and electric energy required for back-up heating needs to be added as the SCOP value is calculated in respect of the EN 14825. Thus, annual electric consumption of R446A was determined to be high in comparison to other refrigerants. R446A consumes approximately 44% less electric energy in the cooling mode in comparison to other refrigerants. The variations of heat pump capacity (Qhp), total heat demand (Qhd), total electric consumption (Whp) and Tbi temperature value by the ambient temperatures when alternative refrigerants were used in the heating mode are shown in Fig. 14. Tbi is the bivalent temperature value where heat demand of the volume that is to be heated starts to be met by the heat pump completely. It is desired that Tbi value is at low temperatures. R454B and R452B refrigerants have low Tbi values and their Qhd and Qhp curves are very close to one another. Therefore, it can be said that these two refrigerants can fulfill heat loss required at low temperatures. Tbi value of R446A is very high as shown in Fig. 14. Meeting heat demand of the volume that is to be heated by a heat pump completely is only possible at the points where the ambient temperature is high. Therefore, it is seen that R446A refrigerant is inadequate to meet total heat demand. Thus, extra resistance heater is needed to meet the necessary heat demand. Electric energy consumed by this resistance heater leads to an increase in Whp value that is necessary for heating by R446A. It is seen in the graphs of R446A and R447A refrigerants that the heat pump doesn’t meet total heat demand completely at 10 to +8 °C. R452B and R454B refrigerants met total heat demand by the heat pump at almost all temperatures. A comparison of the SCOP and SEER values for refrigerants, which is the main target of the study, is shown in Fig. 15. The SEER values for all refrigerants were very close to each other. The SEER values of alternative refrigerants were found to be approximately 1–2% higher than that of R410A. The SEER values of all refrigerants were over 6 values. The situation was different for the SCOP values. The SCOP values of R410A, R452B and R454B were very close to one another. The SCOP value of R446A was 58% lower than the SCOP value of R410A. A resistance heater used for heating additionally played a role for the occurrence of this difference as indicated above. 7. Conclusion In this study, the effect of using R446A, R447B, R452B and R454B alternative refrigerants instead of R410A in an air conditioning device on the SCOP and SEER values was investigated theoretically. The GWP values of alternative refrigerants that can be used
instead of R410A are lower than 750. The EU regulations are met by these characteristics. The energy parameters of the refrigerants with the appropriate GWP and ODP values will be determinants for deciding. Many numbers of alternatives make standardization difficult. Maintenance and service may be difficult because different air conditioner manufacturers prefer different refrigerants. Therefore, it will be appropriate to determine the most suitable refrigerant in every aspect (it’s environmental effects, price, energy consumption, etc.) and to design the devices for this refrigerant. The following results were obtained as a result of the research: R452B and R454B refrigerants, whose SCOP results are very close to the SCOP value of R410A, become prominent among the alternatives. The same refrigerants can be accepted as a good alternative for R410A with a higher SEER value than that of R410A. R446A has the lowest energy consumption for cooling and the lowest cooling capacity. R446A may be suitable especially for air conditioning devices needed in hot and tropical climates and only cooling air conditioning devices. However, it will need heat exchangers that are larger than R410A. The SEER values of all alternative refrigerants are higher than that of R410A. R454B refrigerant has the lowest annual energy consumption for heating. The compressor outlet temperatures of alternative refrigerants in cooling mode are lower than that of R410A. Therefore, the system can operate more comfortably and safely. The refrigerant amount circulating in the system decreases in cooling mode as parallel to a decrease in the ambient temperature. The lowest mass flow rate value was determined when R446A was used in the system. R446A has the lowest electric consumption in heating mode in comparison to partial loads. However, it has the highest electric consumption value to meet heat demand since it has a high Tbi value.
References [1] Regulation (EU) No 517/2014 of the European Parliament and the Council of 16 April 2014 on fluorinated greenhouse gases and repealing Regulation (EC) No 842/2006. Official Journal of the European Union. (accessed 14 December 2015). [2] Commission Delegated Regulation (EU) No 626/2011 of 4 May 2011 supplementing Directive 2010/30/EU of the European Parliament and of the Council with regard to Energy Labeling of Air Conditioners. [3] Commission Regulation (EU) No 206/2012 of 6 March 2012, implementing Directive 2009/125/EC of the European Parliament and of the Council with regard to Eco design Requirements for Air Conditioners and Comfort Fans. . [4] S. In, K. Keumnam Cho, B. Lim, C. Lee, Partial load performance test of residential heat pump system with low-GWP refrigerants, Appl. Therm. Eng. 85 (2015) 179–187. [5] W. Chen, A comparative study on the performance and environmental characteristics of R410A and R22 residential air conditioners, Appl. Therm. Eng. 28 (2008) 1–7. [6] Q. Tian, D. Cai, L. Ren, W. Tang, Y. Xie, G. He, F. Liu, An experimental investigation of refrigerant mixture R32/R290 as drop-in replacement for HFC410A in household air conditioners, Int. J. Refrig. 57 (2015) 216–228. [7] X. Xu, Y. Hwang, R. Radermacher, Performance comparison of R410A and R32 in vapor injection cycles, Int. J. Refrig. 36 (2013) 892–903. [8] A. Alabdulkarem, R. Eldeeb, Y. Hwang, V. Aute, R. Radermacher, Testing, simulation and soft-optimization of R410A low-GWP alternatives in heat pump system, Int. J. Refrig. 60 (2015) 106–117. [9] M. Dongellini, C. Naldi, G.L. Morini, Seasonal performance evaluation of electric air-to-water heat pump systems, Appl. Therm. Eng. 90 (2015) 1072–1081. [10] A. Mota-Babiloni, J. Navarro-Esbri, A. Barragan, F. Moles, B. Peris, Theoretical comparison of low GWP alternatives for different refrigeration configurations taking R404A as baseline, Int. J. Refrig. 44 (2014) 81–90.
A.G. Deveciog˘lu / Applied Thermal Engineering 125 (2017) 401–411 [11] F. Moles, J. Navarro-Esbri, B. Peris, A. Mota-Babiloni, A. Barragan-Cervera, Theoretical energy performance evaluation of different single stage vapour compression refrigeration configurations using R1234yf and R1234ze(E) as working fluids, Int. J. Refrig. 44 (2014) 141–150. [12] A. Mota-Babiloni, J. Navarro-Esbri, A. Barragan-Cervera, F. Moles, B. Peris, Analysis based on EU regulation No 517/2014 of new HFC/HFO mixtures as alternatives of high GWP refrigerants in refrigeration and HVAC systems, Int. J. Refrig. 52 (2015) 21–31. [13] D.C. Montgomery, Design and Analysis of Experiments, fourth ed., John Wiley and Sons Inc., New York, 1997. [14] V.H. Panato, M.A. Queiroz, L.M.P. Souza, A.H.P. Antunes, E.P.B. Filho, Analysis of the drop-in operation of a refrigeration system by the response surface methodology, in: 16th International Refrigeration and Air Conditioning Conference, 11–14 July 2016, Purdue, USA. [15] K. Çomakli, F. S ß imsek, O. Çomakli, B. S ß ahin, Determination of optimum working conditions R22 and R404A refrigerant mixtures in heat-pumps using Taguchi method, Appl. Energy 86 (2009) 2451–2458.
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[16] European Committee for Standardization, Standard EN 14825: 2013, Air Conditioners, Liquid Chilling Packages and Heat Pumps, with Electrically Driven Compressors, for Space Heating and Cooling e Testing and Rating at Part Load Conditions and Calculation of Seasonal Performance, 2013. [17] E.W. Lemmon, M.L. Huber, M.O. McLinden, NIST Standard Reference Database 23: Reference Fluid Thermodynamic and Transport Properties-REFPROP, Version 9.1, National Institute of Standards and Technology, Standard Reference Data Program, Gaithersburg, 2013. _ Dinçer, M. Kanog˘lu, Refrigeration Systems and Applications, John Wiley & [18] I. Sons, New York, A.B.D., 2010. [19] V. Oruç, A.G. Deveciog˘lu, Thermodynamic performance of air conditioners working with R417A and R424A as alternatives to R22, Int. J. Refrig. 55 (2015) 120–128.
Contents lists available at ScienceDirect
Applied Thermal Engineering journal homepage: www.elsevier.com/locate/apthermeng
Research Paper
Seasonal performance assessment of refrigerants with low GWP as substitutes for R410A in heat pump air conditioning devices Atilla Gencer Deveciog˘lu Dicle University, Department of Mechanical Engineering, 21280 Diyarbakır, Turkey
h i g h l i g h t s Four refrigerants alternatives to R410A were compared. R452B was seen to be a suitable alternative R410A at heating mode. The SEER values of the refrigerants were very close to each other. SEER value for all of the alternative refrigerants was found higher compared to R410A. R446A is more suitable for air conditioners performing cooling only.
a r t i c l e
i n f o
Article history: Received 12 December 2016 Revised 26 May 2017 Accepted 3 July 2017 Available online 5 July 2017 Keywords: SEER SCOP R410A GWP
a b s t r a c t In this study, the effect of the utilization of refrigerants with a low Global Warming Potential (GWP) instead of R410A in air conditioners on seasonal efficiency was evaluated theoretically. In the study, four different alternative refrigerants were used namely, R446A, R447A, R452B and R454B. The GWP values of all of the studied refrigerants were lower than 750. The calculations were made according to the EN 14825 standard. Average seasonal conditions were applied for heating. The results showed that R446A had the lowest heating capacity and electric consumption. R446A required electric energy consumption for extra heating due to its low heating capacity and therefore it had the lowest seasonal coefficient of performance (SCOP) value. The seasonal energy efficiency ratio (SEER) values of the refrigerants were very close to each other and their average was 6.56. R452B had the most suitable with 3.92 SCOP value among the alternative refrigerants. Ó 2017 Elsevier Ltd. All rights reserved.
1. Introduction The European Union (EU) scheduled the utilization of refrigerants with GWP value in air conditioning devices by a regulation [1]. Again, in the EU countries, a new energy label application to be used in air conditioning devices was initiated in the scope of the Regulation of Energy Labeling of Air Conditioner in 2013. In pursuant of this regulation, there must be inscriptions on labels showing the seasonal energy efficiency with regard to both cooling and heating. Performance measurement methods have changed for air conditioners lower than 12 kW and the seasonal efficiency term has been defined by Eco-Design directive that determines minimum efficiency requirements for energy-utilizing products [2]. Three regions were determined for the heating season as colder, average and warmer. In pursuant of 206/2012 EU regulation [3], air conditioning systems must have the minimum energy efficiency
E-mail address: [email protected] http://dx.doi.org/10.1016/j.applthermaleng.2017.07.034 1359-4311/Ó 2017 Elsevier Ltd. All rights reserved.
values shown in Table 1 starting from January 1, 2014. Singleduct and double-duct air conditioners and the devices other than these are classified based on their cooling capacity and the GWP values of the used refrigerants. R410 refrigerant with 2088 GWP value is used frequently in many air conditioning devices [4]. Due to this high GWP value, refrigerants that can be used instead of R410 with low GWP value and high energy parameter value have been searched. In a study conducted to investigate R410A was used instead of R22, the coefficient of cooling performance (COP) values of R22 and R410A were found to be very close to each other at different condenser temperatures [5]. Although not a long time has passed, alternative refrigerants are being investigated due to high GWP value of R410. A study was conducted by using a R32/R290 mixture instead of R410 [6]. The cooling and heating values of R32/R290 mixture were higher than those of R410A and the COP and energy efficiency ratio (EER) values of R410 were higher than those of R32/ R290 mixture. R32 and R410A in domestic air conditioning device
A.G. Deveciog˘lu / Applied Thermal Engineering 125 (2017) 401–411
402
Nomenclature COP EER GWP SCOP SEER h h0 _ m P Q T Wel Vdisp m_ suc
Subscripts comp compressor c cooling e evaporator in inlet out outlet L low H high h heating d design bi bivalent ol operating limit a ambient is isentropic vol volumetric on operating state hd heat demand hp heat pump
coefficient of performance energy efficiency ratio global warming potential seasonal coefficient of performance seasonal energy efficiency ratio enthalpy, kJ kg1 enthalpy for isentropic case, kJ kg1 mass flow rate, kg s1 pressure, kPa heat capacity, kW temperature, °C electrical power input to the system, kW displacement specific volume m3 kg1 efficiency number of hours in a year rpm
g
L n_
Table 1 Requirements for minimum energy efficiency. Air conditioners, except double and single duct air conditioners
If If If If a
GWP GWP GWP GWP
of of of of
refrigerant > 150 for air conditioner capacity < 6 kW refrigerant 150 for air conditioner capacity < 6 kW refrigerant > 150 for air conditioner capacity 6–12 kW refrigerant 150 for air conditioner capacity 6–12 kW
Double duct air conditioners
Single duct air conditioners
SEER
SCOPa
EERrated
COPrated
EERrated
COPrated
4.60 4.14 4.30 3.87
3.80 3.42 3.80 3.42
2.60 2.34 2.60 2.34
2.60 2.34 2.60 2.34
2.60 2.34 2.60 2.34
2.04 1.84 2.04 1.84
Heating season: Average.
were compared in both the heating and cooling mode [7]. The cooling and heating capacities and COP values of R32 were reported to be better than those of R410A. R410A and R32, D2Y60 and L41A refrigerants in a 10 kW split air conditioning system were compared for both heating and cooling modes [8]. The GWP values of each of the three alternatives were lower than 750. R32 had the highest cooling capacity. The heating capacity values of R410A and L41A were very close to one another. The refrigerants R410A and R32 and R446A of domestic heat pump were compared under partial loads with the EN 14825 calculation method [4]. There was a 2.1% increase in the EER value when R32 was used and a 1.4% increase when R446A was used. In another study, the air to water heat pump’s seasonal performances, which were designed in pursuant of the EN 14825 were evaluated [9]. In the study, heat pumps with single compressor, multicompressor and variable speed compressor were compared. The theoretical studies comparing the performances of refrigerants with low GWP value in various vapor compression refrigeration systems [10,11] may be beneficial to have an idea prior to the experimental study. In a theoretical study comparing the utilization of refrigerants with low GWP value instead of R410A [12], the COP values of the alternative refrigerants were found to be higher than that of R410A. Response surface method (RSM) was used for mathematical model of the system. RSM is a mathematical modelling and statistical evaluation method used for optimization of input parameters (independent variable) of a system [13].
Related to this subject, Panato et al. [14] have performed analysis through RSM using different refrigerants as drop-in application. They expressed that experimental data were compatible with the results obtained with RSM. All analyzed values of (R2), for all fluids were greater than 0.80. In order to determine optimum operating conditions and the effect of fixed parameters on the system working with R22 and R404A, Çomaklı et al. [15] used experimental design method proposed by Taguchi. It was indicated that COP was mostly affected by the air temperature at condenser inlet. In this study, the energy parameters of refrigerants with low GWP that could be used instead of R410A were calculated theoretically for the average heating and cooling season based on the test point temperatures. In the study, the calculation method provided in the EN14825 standard was used. The refrigerants with low GWP value that could be used instead of R410A were compared based on their SEER and SCOP values. 2. The EN 14825 calculation method A calculation method has been determined for the calculation of the SCOP and SEER values which are mandatory to be placed in the labels of air conditioning devices per the Eco-design Directives. In the calculations made with this method, the reference design conditions shown in Table 2 are considered to maintain these values in air conditioning devices other than the single-duct and doubleduct air conditioning devices [3].
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Table 2 The test and design temperatures for heating and cooling modes and partial load ratios. Point
Cooling
A B C D Tdesign (°C) Tbi (°C) TOL (°C)
Heating (Warmer)
Heating (Average)
Ttest (°C)
Part loads
Ttest (°C)
Part loads
Ttest (°C)
Part loads
Ttest (°C)
Part loads
– 35 30 25 20 35 – –
– 100% 74% 47% 21%
– – 2 7 12 2 7 2
– – 100% 64% 29%
– 7 2 7 12 10 2 7
– 88% 54% 35% 15%
15 7 2 7 12 22 7 15
82% 61% 37% 24% 11%
In the previously used calculation method, nominal efficiency was calculated at a specific ambient temperature (35 °C for cooling and 7 °C for heating), at full load and without considering the energy consumed by auxiliary components. The observance frequency of a 35 °C temperature in real conditions is low. The ambient temperatures during the summer months, when a cooling necessity occurs, range between 16 and 40 °C and change based on geographical location and climatic conditions. Moreover, there is less cooling need at low ambient temperatures. The EN 14825 standard has been developed to calculation the efficiencies of air conditioning devices more realistically [16]. The seasonal data of Strasburg city have been accepted as the basis and the design temperature has been selected as 35 °C for the SEER calculation in the EN 14825 standard [3]. Four test points have been defined in the standard. These are 20, 25, 30, and 35 °C. The device can operate under partial loads at different ambient temperatures since the cooling necessity is different. Therefore, air conditioning devices are tested under different partial loads (100%, 74%, 47% and 21%). The SEER value is calculated based on the EER values calculated at these four test points. Three different climatic zones have been defined for the SCOP calculation in the EN 14825 standard. These are the average climatic zone, warmer climatic zone and colder climatic zone. Strasburg city has been taken as the basis for the average climatic zone, Athens has been taken as the basis for the warmer climatic zone and Helsinki has been taken as the basis for the colder climatic zone. The test and design temperatures and partial load ratios used in the SCOP and SEER calculations are shown in Table 2. In addition to these test and design temperatures, bivalent temperature (Tbi) value has been defined as a temperature when the heat demand of a heat pump starts to be fulfilled completely. The operation limit temperature (TOL) has been defined as the lowest temperature when the heat pump can operate. The observance time of the ambient temperatures in these climatic zones during a year, in other words their frequencies are shown in Fig. 1 [2]. These data are used in the SCOP and SEER calculations. The frequency of temperature observance times, which is one of the starting points of seasonal calculation, has concentrated in the warm zone between +2 and +8 °C as shown in Fig. 1. Similarly, the cooling season need has concentrated between +20 and +25 °C. 2.1. The determination of heat demand used for the calculation of the SCOP The heat demand desired for various test temperatures is calculated by Eq. (1) based on the design temperatures at the climatic zone that is to be calculated. The design temperatures shown in Table 2 [16]. The heat demand curve is formed with this expression.
Pdesignh ¼
Heating (Colder)
T test 16 T designh 16
ð1Þ
The percentages of heat demand at various test temperatures for different climatic zones (warmer, average, colder) are shown in Fig. 2. It must be considered that these percentages (partial load) are not the partial load ratios of a heat pump. The heat demand Pdesignh value and the SCOP are placed on an air conditioning label together. Manufacturers use Pdesignh value to describe the size of a heat pump. Hence, consumers can select the correct device agreeable with their climate condition. 2.2. Calculation of the SCOP value In the study, theoretically calculated values for test temperatures were used and the other intermediate values were determined by interpolation as specified in the EN 14825. Extrapolation can be used within a narrow range. It is necessary to find out how much electric energy is consumed at what temperature in order to determine the compressor electric consumption used in the COP calculation. It is necessary to know what temperature is seen with what frequency when doing this calculation. The observance frequency of temperatures is found for the regions by using the graph shown in Fig. 1. In this study, the average temperature season was used for all refrigerants. Electric consumption, the COP and heating capacity values found by the theoretical study were determined for the four temperature points at the average heating season. Other temperature values were found by interpolation based on these values. Subsequently, partial loads at the test temperatures were determined based on the design temperature (Tdesignh = 10 °C in average heating season) of a 7 kW capacity air conditioning device. Heat demand at the test temperature was determined by multiplying heat demand at the design temperature by partial loads. Total heat demand was found by multiplying heat demand with the observance hour number of the test temperature. Total electric consumption was found by multiplying electric consumption per hour with the observance hour number of the test temperature. The ratio of total heat demand for all test temperatures to total electric consumption gives the seasonal cooling performance coefficient (SCOPon) at the operating state. The SCOP value is calculated in pursuant of the EN14825.
SCOP ¼
Qh Qh SCOPon
þ Waux
ð2Þ
Waux value is the electric energy consumed by the device when it is (off), at standby (sb), when the crankcase heater is on (ch) and the thermostat is off (to) and is calculated as follows [16].
W aux ¼ Lto wto þ Lsb wsb þ Lch wch þ Loff woff
ð3Þ
Here, L value indicates annual operation hour number of a heat pump and w indicates energy consumption.
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404
Fig. 1. The observance number of ambient temperatures used in calculations for climatic zones.
3. Comparison of refrigerants’ characteristics
120% Average Warmer
100%
Partial load
Colder 80% 60% 40% 20% 0% -18
-8
2
12
22
Ttest (°C) Fig. 2. Determining the heating demand curve at Pdesignh, different climatic conditions.
The missing heat demand amount between the heat demand and the heat pump capacity will be met by a resistance heater. Electric energy consumed by the resistance heater is added to total electric consumption. 2.3. The SEER value calculation In the SEER calculation, electric consumption, the COP and cooling capacity values found by the theoretical study for the four temperature points at the cooling season were determined. Other test temperature values were determined by interpolation based on these values. By multiplying hourly cooling demand with the test temperature observance hour number, total cooling demand was found for that test temperature. Total electric consumption was found by multiplying hourly electric consumption with the test temperature observance hour number. The ratio of total cooling demand for all test temperatures to total electric consumption gives the seasonal energy efficiency ratio at the operating state (SEERon). The SEER value is calculated in pursuant of the EN 14825 as shown in Eq. (4) by dividing annual cooling demand to the sum of power draw for annual cooling and power draw at auxiliary positions [16].
SEER ¼
Qc SEERon
Qc þ Waux
ð4Þ
In this study, refrigerants with low GWP value and better energy parameters were investigated instead of R410A refrigerant used in air conditioning systems especially. R410A refrigerant is used in numerous systems including mono type split air conditioners, VRF systems with hundreds of indoor units, and chillers. Some properties of the refrigerants used in the study are shown in Table 3. Critical temperature values of all refrigerants are close to each other. Therefore, they may be used in a system designed for R410. The GWP values of all alternative refrigerants are lower than 750 as shown in Table 3. The saturation curves of R447A, R452B and R454B refrigerants are very close to each other especially as shown in Fig. 3. R446A (mixture ingredient R32/R600/R1234ze(E); mixture ratio, 29:3:68) HFC/HC/HFO mixture, R447A (mixture ingredient, R32/R125/R1234ze(E); mixture ratio, 68:3.5:28.5), R452B (mixture ingredient, R32/R125/R1234yf; mixture ratio, 67:7:26) and R454B (mixture ingredient, R32/R1234yf; mixture ratio, 68.9:31.1) HFC/ HFO mixture are the refrigerants used instead of R410A (mixture ingredient, R32/R125; mixture ratio, 50:50).
4. Theoretical analysis Energy parameters at the test temperature points were found by a theoretical study. The heating and cooling capacities of the investigated refrigerants were determined for an air conditioning device with a 7 kW cooling capacity. An air conditioning device with a heating feature consists of basic components including a compressor, condenser, expansion valve, evaporator, and fourway valve. During the heating process, the refrigerant’s direction is changed by a four-way valve and this way, the condenser acts like an evaporator (Fig. 4). Some assumptions were made for the system’s analysis; it was assumed that the system operated steady state conditions. Heat and pressure losses in the air conditioning system components and pipes were ignored. Energy consumption of the fan of the evaporator and condenser was ignored. There was no kinetic and potential energy terms, and the auxiliary components had no power consumption. The compressor volumetric efficiency, isentropic efficiency and displacement values were modified based on the test temperature points in the theoretical study to simulate the partial loads given in the EN 14825. Thus, it was supposed that the system operated like a variable speed. These values are shown in Table 4. The study differed with this aspect from the previous theoretical studies.
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Table 3 Some properties of the refrigerants used in the study [17].
Critical temperature (°C) Critical pressure (kPa) Liquid density (kg m3) (25 °C) Boiling point (101.3 kPa) GWP
R410A
R446A
R447A
R452B
R454B
71.3 4901 459 48.5 2088
92.8 4678 442 30.8 490
80.3 5269 421 45.6 580
81.9 5289 421 45.5 695
83.7 5334 418 44.6 465
Fig. 3. Pressure enthalpy change for the refrigerants.
Fig. 4. A schematic illustration of an air conditioner system cycle.
Table 4 Assumptions made for the theoretical analysis: volumetric efficiencies, isentropic efficiencies and displacement ratios according to the test temperatures. Ttest (°C)
gvol gis
Vdisp (m3h1)
10 0.75 0.60 6.20
7 0.80 0.70 6.20
2 0.85 0.85 3.00
7 0.90 0.90 1.70
Superheat and subcooling values were accepted as respectively as 8 and 5 °C during the heating process. Tev value was determined as constant and 10 °C lower than the test temperatures. The condensation temperature was accepted as constant 43 °C for all test temperatures. Superheat and subcooling values were accepted as 10 and 5 °C respectively for the cooling process. The condensation temperature values were accepted to be as constant and 10 °C
12 0.95 0.90 0.70
20 0.9 0.85 6.8
25 0.85 0.8 4.5
30 0.8 0.7 2.8
35 0.75 0.6 2.1
higher than the test temperatures. The evaporation temperature was accepted as constant 0 °C for all test temperatures. 4.1. Calculation of energy parameters Cooling capacity (Qe) of the cooling cycle is calculated as follows based on the above given information [18]:
A.G. Deveciog˘lu / Applied Thermal Engineering 125 (2017) 401–411
406
_ e;out he;in Þ Q e ¼ mðh
ð5Þ 1
_ mass flow rate (kg s ), he,out and he,in are the enthalpy valHere, m ues (kJ kg1) of the refrigerant at the outlet and inlet of the evaporator. Electric consumed by the compressor:
_ comp;out hcomp;in Þ W el ¼ mðh
ð6Þ
Here, Wel is the electric power (kW) consumed for the system’s operation and hcomp,out and hcomp,in are the enthalpy values (kJ kg1) of the refrigerant at the output and inlet of the compressor. Heat rejected by the condenser or heat given to the environment that is to be heated at the heating mode is found by the following expression:
Q h ¼ W el þ Q e
ð7Þ
Isentropic efficiency of the system is found by the following expression:
gis ¼
0
ðhcomp;out hcomp;in Þ ðhcomp;out hcomp;in Þ
Table 6 Design planning and calculation results. Calculation no.
Ta (°C)
gis
Vdisp
COP
1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17
1 1 1 1 1 1 1 1 0 0 0 0 0 0 0 0 0
1 1 1 1 0 0 0 0 1 1 1 1 0 0 0 0 0
0 0 0 0 1 1 1 1 1 1 1 1 0 0 0 0 0
6.29 3.71 4.44 2.62 5.36 3.17 5.36 3.17 4.76 3.36 4.76 3.36 4.06 4.06 4.06 4.06 4.06
ð8Þ
0
Here, hcomp;out is the enthalpy value corresponding to the constant entropy at the compressor outlet of the refrigerant during the compressing process. hcomp,out and hcomp,in are the enthalpies of the compressor corresponding to the discharge and suction temperature values. Volumetric efficiency value used in the study is determined by Eq. (9).
Table 7 The assumptions for analysis of the model. Ta (°C)
gis
Vdisp (m3h1)
Te (°C)
Superheat (K)
Subcooling (K)
15/45
0.4/0.90
2/8
5/+5
3/9
2/5
8 R454B
_ m
gv ol ¼ _ measured mtheoretical
ð9Þ
R410A
6
R446A
Here, theoretical mass flow rate is calculated as follows:
R447A
ð10Þ
Here, Vdisp is the compressor displacement and the values in Table 4 are accepted. n_ is the rotational speed (rpm) of the compressor and tsuc is the specific volume (m3 kg1) value of the refrigerant at the suction temperature of the compressor. Energy performance of a refrigeration system is defined by the coefficient of performance (COP). The COP is calculated as the ratio of the cooling capacity to the consumption of electrical energy per unit time by the compressor.
COP ¼
Qe W el
ð11Þ
5. Model development Mathematical model of COP based on response surface method: Ta,
gis and Vdisp expressions were used to determine COP which is one of the energy parameters. While COP is reduced with Ta, it is improved as gis increases. Note that gis and Vdisp have been added to the model for considering the effect of variation of partial loads on COP. Evaporation temperature, volumetric efficiency, superheat and subcooling values were assumed to be constants as in theoretical study. For the calculations, Ta, gis and Vdisp have been selected as variable parameters which are given in Table 5.
Table 5 Experimental parameters and their levels. Parameters
Level 1 (1)
Level 2 (0)
Level 3 (+1)
Ta (°C)
20 0.85 2.1
25 0.75 4.5
35 0.60 6.8
gis
3 1
Vdisp (m h
)
COP
_ theoretical m
V disp n_ ¼ tsuc 60
R452B
4
Predicted
2
0 15
20
25
30
35
40
Ta (°C) Fig. 5. The comparison of calculated results with the results obtained from model (predicted).
Theoretical calculations were conducted to investigate the effects of Ta, gis and Vdisp on COP. The calculations in this study were performed using Box-Behnken experimental design planning. The performance outputs through the calculations are presented in Table 6. For the obtained calculation results, statistical analyses were conducted using Design Expert 6.0 software and a suitable regression model was formed. The second order equation composed for COP is
COP ¼ 4:06 ð1:1 T a Þ ð0:72 gis Þ þ ð0:21 T a 2Þ þ ð0:19 T a gis Þ ð12Þ It is seen from the model that value of Vdisp is a variable which does not influence the result. The model gives very close results for the refrigerants studied considering the values given in Table 7. R2 term called as determination coefficient was determined as 0.9996 for COP. Comparison of the computed results and obtained results from mathematical model (predicted) for all refrigerants investigated in
A.G. Deveciog˘lu / Applied Thermal Engineering 125 (2017) 401–411 Table 8 Comparison of experimental data with present model.
407
2.0
COP
R410A R446A R447A R452B R454B
Ta = 35 (°C)
Oruç and Deveciog˘lu [19] R22 R417A R424A
3.77 3.03 3.54
3.13 2.40 2.52
Prediction of model R22 R417A R424A
3.49 3.16 3.81
2.86 2.48 2.49
Wel (kW)
1.6
Ta = 25 (°C)
1.2
Heating mode
0.8 0.4 0.0
6. Results and discussions
0
40
60
80
100
120
80
100
120
(a) 3.0 R410A R446A R447A R452B R454B
2.5 2.0
Cooling mode
1.5 1.0 0.5 0.0 0
In the study, theoretical analysis of air-conditioning device was conducted for different refrigerants considering both heating and cooling applications. Therefore, a difference was obtained compared to previous theoretical investigations generally on refrigeration systems. It can be said that partial load modelling in previous theoretical analyses is also different. In terms of SEER and SCOP values which are necessary for energy labelling of airconditioning devices, the suitable refrigerants with low-GWP were suggested. The variation of electric consumption by partial loads for heating and cooling modes of the refrigerants is shown in Fig. 6. A partial load decrease causes reduce of electric consumption in both modes. Electric consumption values of all refrigerants, except for R446A, are close to one another. R446A has the lowest electric consumption value in both modes. R410A has the highest electric consumption at full load (100%) in both modes. The variation of heating capacities at the ambient temperatures for various refrigerants in heating mode is shown in Fig. 7 a. An increase of ambient temperature decreases heat loss of heating space. Therefore, air conditioning device’s thermal capacity decreases as the ambient temperature decreases with partial load changes at the ambient temperature points. Thermal capacities of R454B and R452B were found as approximately 3–5% lower than that of R410A. Approximately there was a 50% capacity loss for R446A especially. An increase in ambient temperature will increase cooling of the space necessity. The variation of cooling capacities of the air conditioning device with the increase of ambient temperature is shown in Fig. 7b. A decrease in ambient temperature in the systems operating respect to partial load will reduce cooling capacity of the air conditioning device. Cooling capacities change as parallel to heating capacities. Cooling capacities of R454B and R452B refrigerants were determined as approximately 2–5% lower than that of
20
Partial load (%)
Wel (kW)
the study can be seen in Fig. 5. Obviously, predicted results are very close to the calculated results. The validation of proposed model was verified considering experimental study by Oruç and Deveciog˘lu [19]. The ambient temperature values were 25 and 35 °C while gis was 0.50–062 in their investigation. The system was continuously operating at full capacity. The predicted results and experimental results are presented in Table 8. The results are close to each other, however, the difference is some higher for R417A. These differences may have been developed due to the fact that gis and Vdisp values, which are input parameters of the model assumed to operate under partial loads, are different. The obtained model presents compatible results. Therefore a new model has been proposed in the literature for estimating and optimizing COP value of refrigeration systems through suitable input parameters.
20
40
60
Partial load (%)
(b) Fig. 6. The variation of electric consumption by partial loads (a) heating mode and (b) cooling mode.
R410A. Cooling capacity of R446 is approximately 45% lower than that of R410A. The variation of mass flow rate value of the refrigerant that needs to circulate in the system when the air conditioning device is operating by ambient temperatures for various refrigerants is shown in Fig. 8. Mass flow rate values decrease with the decrease of ambient temperatures in cooling mode depending on the volumetric efficiency and displacement values. R446A has the lowest mass flow rate value. Mass flow rate values of R454B and R452B refrigerants are very close to each another. The decrease of mass flow rate value by the ambient temperatures in heating mode is shown in Fig. 8a. It is understood that an air conditioning device cannot achieve the necessary heating even when it operates at full capacity due to a low temperature value (10 °C). When the ambient temperature is 7 °C, mass flow rate value increases with the increase of suction pressure. However, there was a decrease seen in mass flow rate value at partial load as parallel to an increase in the ambient temperature. The condensation temperature was accepted as constant in heating mode. The variation of low pressure (suction pressure) PL value based on partial loads is shown in Fig. 9. An increase in the evaporation temperature causes an increase in low pressure value since the decrease of partial load depends on the ambient temperature. PL value of R410A is approximately 50% higher than that of R446A.
A.G. Deveciog˘lu / Applied Thermal Engineering 125 (2017) 401–411
408
6
900
5
R452B R454B
750
4
R447A R446A
600
3 2
R410A
Heating mode
R410A
R446A R447A
PL (kPa)
Qh (kW)
Heating mode
R452B R454B
450 300
1
150
0 -10
-7
2
7
0
12
0
Ta ( °C)
20
40
60
80
100
120
Partial load (%)
7
Cooling mode
Qe (kW)
Fig. 9. The variation of low pressure with partial load.
R410A
6
R452B R454B
5
R447A
R410A R446A R447A R452B R454B
2700
R446A
4 3
PH (kPa)
2200 2 1 35
30
25
20
Ta ( °C)
Cooling mode
1700
1200
Fig. 7. The variation of heating and cooling capacity by the test temperatures.
700
0.025
0
Heating mode
m (kgs-1)
0.020
40
60
80
100
120
Partial load (%)
R410A R446A R447A R452B R454B
0.015
20
Fig. 10. The variation of high pressure with partial load.
0.010
6.0 0.005
5.5 5.0
0.000 -10
-7
2
7
12
4.5
COP
Ta (°C)
(a)
R4446A R4447A R454B R452B R410A
4.0 3.5
Cooling mode
0.04
R410A R446A R447A R452B R454B
m (kgs-1)
0.03
0.02
3.0 2.5 2.0 -10
-7
2
7
12
Ta ( °C) Fig. 11. The variation of the COP value by the test point temperatures.
0.01
0.00 35
30
25
20
Ta ( °C)
(b) Fig. 8. The variation of mass flow rate value with the test point temperatures (a) heating mode and (b) cooling mode.
The evaporation temperature in cooling mode was accepted as constant. The variation of high pressure value PH by partial loads is shown in Fig. 10. A decrease in the condenser temperature causes a decrease in high pressure value since the decrease of partial load depends on a decrease in the ambient temperature. High pressure values of R454B and R452B refrigerants are very close to each other. Moreover, it is seen in the graph that R410A has the highest PH value and R446A has the lowest PH value.
A.G. Deveciog˘lu / Applied Thermal Engineering 125 (2017) 401–411
The variation of the COP values based on the ambient temperatures for the refrigerants in heating mode is shown in Fig. 11. The COP values of R452B, R454B and R410A at 10 °C are equal to one another. The COP value of R410A was found to be approximately 1% higher than those of R452B and R454B as the ambient temperatures increased. The COP values increased 35% in the average as the ambient temperatures of all these refrigerants increased from 7 to 7. An air conditioning device will take more heat energy from the environment as the ambient temperature increases. It will consume less electric during this process. Therefore, an increase in the ambient temperatures caused the COP values to increase. Since the COP values of R452B and R454B were close to that of R410A, they may be suitable refrigerants for heating. The variation of the EER values by the ambient temperatures in cooling mode is shown in Fig. 12. The EER values of all alternative refrigerants were found to be higher than that of R410A. R446A was determined as the refrigerant with the highest EER value as differing from heating mode. A low electric consumption caused determining a high EER value for R446A. A low electric consumption was due to low evaporator pressure and low discharge pressure. A 2–7% difference was determined between the EER values of the refrigerants. All of the refrigerants can be used instead of R410A owing to these close values. Annual electric consumptions of the refrigerants are shown in Fig. 13. Annual electric consumption for the average heating season where the study was conducted was approximately 3 folds higher than that of the cooling season. More electric is consumed for heating especially at low temperatures in comparison to the cooling mode. Furthermore, when heat loss cannot be met completely, an additional resistance heater steps in and increases electric consumption. R446A consumes approximately 55% more electric energy in comparison to other refrigerants in the heating mode. The reason behind this is that R446A cannot meet heat demand of the volume that is to be heated only by a heat pump in the
7.0 R410A R452B R447A R454B R446A
6.5 6.0 5.5
EER
5.0 4.5 4.0 3.5 3.0 2.5 2.0
35
30
25
20
Ta ( °C) Fig. 12. The variation of the EER value by the test point temperatures.
7000 6000
(kWh)
Annual consumption of electricity
8000
R410A
5000
R452B
4000
R454B 3000
R447A
2000
R446A
1000 0 Heating
409
Cooling
Fig. 13. Annual electric consumption in the heating and cooling modes.
1200
1200 R446A
Qhp
1000
Qhp
R452B
Whp
Whp
1000
Qhd
800
Energy (kWh)
Energy (kWh)
Qhd Tbi
600 400 200
800
Tbi
600 400 200
0
0 -12
-8
-4
0
4
8
12
16
20
-12
-8
-4
0
4
8
12
1200
20
1200 R447A
Qhp
1000
Qhp
R454B 1000
Whp
Whp Qhd
800
Energy (kWh)
Qhd
Energy (kWh)
16
Ta (°C)
Ta (°C)
Tbi
600 400 200
800
Tbi
600 400 200
0
0 -12
-8
-4
0
4
Ta (°C)
8
12
16
20
-12
-8
-4
0
4
8
12
16
20
Ta (°C)
Fig. 14. The variation of heat pump capacity, total heat demand, total electric consumption and Tbi values of the refrigerants by the ambient temperature.
A.G. Deveciog˘lu / Applied Thermal Engineering 125 (2017) 401–411
410
7.0 6.0 5.0 R410A 4.0
R452B R454B
3.0
R447A 2.0
R446A
1.0 0.0 SCOP
SEER
Fig. 15. The variation of the SCOP and SEER values of the refrigerants.
heating mode. The difference between the heat pump’s capacity and the heat demand of the volume that is to be heated is met by a resistance heater. Electric energy consumed by an air conditioning device and electric energy required for back-up heating needs to be added as the SCOP value is calculated in respect of the EN 14825. Thus, annual electric consumption of R446A was determined to be high in comparison to other refrigerants. R446A consumes approximately 44% less electric energy in the cooling mode in comparison to other refrigerants. The variations of heat pump capacity (Qhp), total heat demand (Qhd), total electric consumption (Whp) and Tbi temperature value by the ambient temperatures when alternative refrigerants were used in the heating mode are shown in Fig. 14. Tbi is the bivalent temperature value where heat demand of the volume that is to be heated starts to be met by the heat pump completely. It is desired that Tbi value is at low temperatures. R454B and R452B refrigerants have low Tbi values and their Qhd and Qhp curves are very close to one another. Therefore, it can be said that these two refrigerants can fulfill heat loss required at low temperatures. Tbi value of R446A is very high as shown in Fig. 14. Meeting heat demand of the volume that is to be heated by a heat pump completely is only possible at the points where the ambient temperature is high. Therefore, it is seen that R446A refrigerant is inadequate to meet total heat demand. Thus, extra resistance heater is needed to meet the necessary heat demand. Electric energy consumed by this resistance heater leads to an increase in Whp value that is necessary for heating by R446A. It is seen in the graphs of R446A and R447A refrigerants that the heat pump doesn’t meet total heat demand completely at 10 to +8 °C. R452B and R454B refrigerants met total heat demand by the heat pump at almost all temperatures. A comparison of the SCOP and SEER values for refrigerants, which is the main target of the study, is shown in Fig. 15. The SEER values for all refrigerants were very close to each other. The SEER values of alternative refrigerants were found to be approximately 1–2% higher than that of R410A. The SEER values of all refrigerants were over 6 values. The situation was different for the SCOP values. The SCOP values of R410A, R452B and R454B were very close to one another. The SCOP value of R446A was 58% lower than the SCOP value of R410A. A resistance heater used for heating additionally played a role for the occurrence of this difference as indicated above. 7. Conclusion In this study, the effect of using R446A, R447B, R452B and R454B alternative refrigerants instead of R410A in an air conditioning device on the SCOP and SEER values was investigated theoretically. The GWP values of alternative refrigerants that can be used
instead of R410A are lower than 750. The EU regulations are met by these characteristics. The energy parameters of the refrigerants with the appropriate GWP and ODP values will be determinants for deciding. Many numbers of alternatives make standardization difficult. Maintenance and service may be difficult because different air conditioner manufacturers prefer different refrigerants. Therefore, it will be appropriate to determine the most suitable refrigerant in every aspect (it’s environmental effects, price, energy consumption, etc.) and to design the devices for this refrigerant. The following results were obtained as a result of the research: R452B and R454B refrigerants, whose SCOP results are very close to the SCOP value of R410A, become prominent among the alternatives. The same refrigerants can be accepted as a good alternative for R410A with a higher SEER value than that of R410A. R446A has the lowest energy consumption for cooling and the lowest cooling capacity. R446A may be suitable especially for air conditioning devices needed in hot and tropical climates and only cooling air conditioning devices. However, it will need heat exchangers that are larger than R410A. The SEER values of all alternative refrigerants are higher than that of R410A. R454B refrigerant has the lowest annual energy consumption for heating. The compressor outlet temperatures of alternative refrigerants in cooling mode are lower than that of R410A. Therefore, the system can operate more comfortably and safely. The refrigerant amount circulating in the system decreases in cooling mode as parallel to a decrease in the ambient temperature. The lowest mass flow rate value was determined when R446A was used in the system. R446A has the lowest electric consumption in heating mode in comparison to partial loads. However, it has the highest electric consumption value to meet heat demand since it has a high Tbi value.
References [1] Regulation (EU) No 517/2014 of the European Parliament and the Council of 16 April 2014 on fluorinated greenhouse gases and repealing Regulation (EC) No 842/2006. Official Journal of the European Union.
A.G. Deveciog˘lu / Applied Thermal Engineering 125 (2017) 401–411 [11] F. Moles, J. Navarro-Esbri, B. Peris, A. Mota-Babiloni, A. Barragan-Cervera, Theoretical energy performance evaluation of different single stage vapour compression refrigeration configurations using R1234yf and R1234ze(E) as working fluids, Int. J. Refrig. 44 (2014) 141–150. [12] A. Mota-Babiloni, J. Navarro-Esbri, A. Barragan-Cervera, F. Moles, B. Peris, Analysis based on EU regulation No 517/2014 of new HFC/HFO mixtures as alternatives of high GWP refrigerants in refrigeration and HVAC systems, Int. J. Refrig. 52 (2015) 21–31. [13] D.C. Montgomery, Design and Analysis of Experiments, fourth ed., John Wiley and Sons Inc., New York, 1997. [14] V.H. Panato, M.A. Queiroz, L.M.P. Souza, A.H.P. Antunes, E.P.B. Filho, Analysis of the drop-in operation of a refrigeration system by the response surface methodology, in: 16th International Refrigeration and Air Conditioning Conference, 11–14 July 2016, Purdue, USA. [15] K. Çomakli, F. S ß imsek, O. Çomakli, B. S ß ahin, Determination of optimum working conditions R22 and R404A refrigerant mixtures in heat-pumps using Taguchi method, Appl. Energy 86 (2009) 2451–2458.
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[16] European Committee for Standardization, Standard EN 14825: 2013, Air Conditioners, Liquid Chilling Packages and Heat Pumps, with Electrically Driven Compressors, for Space Heating and Cooling e Testing and Rating at Part Load Conditions and Calculation of Seasonal Performance, 2013. [17] E.W. Lemmon, M.L. Huber, M.O. McLinden, NIST Standard Reference Database 23: Reference Fluid Thermodynamic and Transport Properties-REFPROP, Version 9.1, National Institute of Standards and Technology, Standard Reference Data Program, Gaithersburg, 2013. _ Dinçer, M. Kanog˘lu, Refrigeration Systems and Applications, John Wiley & [18] I. Sons, New York, A.B.D., 2010. [19] V. Oruç, A.G. Deveciog˘lu, Thermodynamic performance of air conditioners working with R417A and R424A as alternatives to R22, Int. J. Refrig. 55 (2015) 120–128.