Paper No. 128
`
`SIMULATION OF R-1234yf PERFORMANCE IN A TYPICAL
`AUTOMOTIVE SYSTEM
`
`Claudio Zilio,(a) J. Steven Brown, (b)(c) and Alberto Cavallini(a)
`
`(a)Dipartimento di Fisica Tecnica, Università di Padova, Padova, 35131, Italy
`(a)Department of Mechanical Engineering, Catholic University of America, Washington, DC 20064 USA
`(c)Corresponding author:
`Department of Mechanical Engineering, Catholic University of America, 620 Michigan Ave, NE, Washington,
`DC 20064 USA
`brownjs@cua.edu; fax:1-202-319-5173
`
`ABSTRACT
`
`Simulations are conducted using R-1234yf (2,3,3,3-tetrafluoropropene; CF3CF=CH2) in a typical baseline
`R-134a small-size European automotive air-conditioning system, where the baseline R-134a system has a
`nominal cooling capacity of 5.8 kW at a compressor volumetric flow rate of 7.8 m3 h-1. If R-1234yf is
`used as a drop-in replacement in this baseline system, its cooling capacity is 2.0 % lower than the R-134a
`value, and its COP is 1.0 % lower than the R-134a value. If on the other hand, the two systems are
`compared at equal cooling capacities, the COP values of the R-1234yf system are 0 % to 4 % lower than
`the R-134a values over operating conditions from idle to highway speeds. While both systems would
`benefit from the use of a liquid-line/suction-line heat exchanger, R-1234yf would benefit somewhat more
`from its use than would R-134a. Also, R-1234yf could benefit from the optimization of the heat
`exchanger circuitries. The thermodynamic and transport properties of R-1234yf are estimated from
`Brown et al. (2009a). The simulation results and analyses presented in this paper demonstrate the
`attractiveness of R-1234yf as a potential replacement for R-134a in automotive applications.
`
`1. INTRODUCTION
`
`The European Union’s f-gas regulations (Regulation (EC) No 842/2006 and Directive 2006/40/EC)
`specify that beginning on January 1, 2011 new models and on January 1, 2017 new vehicles fitted with air
`conditioning can not be manufactured with fluorinated greenhouse gases having global warming
`potentials (GWP) greater than 150. Recently, R-1234yf, which has a 100-year time horizon GWP of four
`relative to carbon dioxide (Nielsen et al. 2007) has been investigated (e.g., Spatz and Minor 2008) as a
`possible replacement fluid for R-134a in automotive applications. This paper uses thermodynamic and
`transport property estimates of R-1234yf (provided in an accompanying paper by Brown et al., 2009a) to
`simulate its potential performance capabilities in an automotive system. These results are compared with
`the baseline refrigerant R-134a.
`
`2. IDEAL VAPOR COMPRESSION REFRIGERATION CYCLE PERFORMANCE
`
`In order to gauge the performance potential of R-1234yf, this section compares R-134a and R-1234yf in
`an idealized vapor compression refrigeration cycle. The cycle chosen is meant to be representative of
`typical automotive operating conditions, and is specified by a constant evaporation temperature of 10 °C,
`a constant condensation temperature of 50 °C, an evaporator superheat of 5 °C, a condenser subcooling of
`5 °C, and a compressor isentropic efficiency of 70 %.
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`A modified version of CYCLE_D Ver. 4.0 (Brown et al., 2009b) was used to perform the simulations,
`where REFPROP 8.0 (Lemmon et al., 2007) was used to determine the thermodynamic properties of R-
`134a and the thermodynamic properties of R-1234yf were taken from Brown et al. (2009a). In the case of
`R-134a, REFPROP implements the Equation of State of Tillner-Roth and Baehr (1994) and fluid-specific
`transport property formulations fitted to extensive experimental data.
`
`Figure 1 presents the simulation results of R-134a and Figure 2 presents the simulation results of R-
`1234yf. The Coefficient of Performance (COP) of the R-134a cycle is 4.118 and that of the R-1234yf
`cycle is 3.973, that is, the R-134a COP is 3.6 % greater than the R-1234yf cycle. The Volumetric
`Cooling Capacity (VCC) of the R-134a cycle is 2858 kJ m-3 and that of the R-1234yf cycle is 2633 kJ
`m-3, that is the R-134a VCC is 8.5 % greater than the R-1234yf value.
`
`Several observations can be made. First, the compressor discharge temperature for R-134a is some 7 °C
`higher than the R-1234yf value, implying a larger superheat loss for R-134a and thus a lowering of its
`COP vis-à-vis R-1234yf. This is a direct result of the larger slope of the saturated vapor line for R-
`1234yf. Second, the compressor work for the two refrigerants is similar; whereas, the refrigeration
`capacity per mass is larger for R-134a, thus implying better COP for R-134a. Third, though not obvious
`from the figures, the expansion losses are larger for R-1234yf, implying a reduction in its COP and
`indicating that it would benefit more than R-134a from the use of a liquid-line/suction-line heat exchanger
`(LLSL-HX). For example, for the above described cycle, though with zero subcooling, the COP and
`VCC values for R-134a are 3.898 and 2705 kJ m-3, respectively, and for R-1234yf are 3.700 and 2451 kJ
`m-3, respectively. If the cycles are supplemented with LLSL-HXs with efficiency values of 75 %, the
`values for R-134a are 4.025 and 2821 kJ m-3, respectively, and for R-1234yf are 3.953 and 2658 kJ m-3,
`respectively. That is, the COP and VCC values for the R-134a cycle with a LLSL-HX are 3.3 % and 4.3
`%, respectively, greater than the baseline R-134a cycle. On the other hand, the COP and VCC values for
`the R-1234yf cycle with a LLSL-HX are 6.8 % and 8.4 %, respectively, greater than the baseline R-
`1234yf cycle. Fourth, though not shown in the figures, the R-1234yf suction vapor is some 15 % denser
`than the R-134a value; however, the refrigeration capacity per mass of R-134a is some 25 % greater than
`the R-1234yf value, leading to the better VCC for R-134a.
`
`In the next section, a more complete simulation model capable of capturing non-ideal effects, e.g., exergy
`losses in the heat exchangers, is used to simulate the performance potentials of R-134a and R-1234yf in a
`typical automotive system.
`
`3. “REAL” VAPOR COMPRESSION REFRIGERATION CYCLE PERFORMANCE
`
`The simulation model used is described in detail in Casson et al. (2003), and thus will only be discussed
`briefly here. It consists of a compressor model (either defined by a compressor map or by isentropic and
`volumetric efficiencies), a tube-fin evaporator model, and either a tube-fin condenser model or a mini-
`channel condenser model. For calculation purposes, the heat exchangers are divided into a series of
`small three-dimensional cells. The refrigerant-side heat transfer, pressure drop, and void fraction
`correlations are listed in Table 1. The air-side heat transfer coefficients are assumed to be uniform and
`given by constant values.
`
`A simulation proceeds by iterating on the evaporation and condensation pressures, ceasing when the
`model successfully matches the refrigerant mass flow rate of the compressor to those in the evaporator
`and condenser.
`
`A typical small-size R-134a-based European automotive air conditioning system with a nominal cooling
`capacity of 5.8 kW at a compressor volumetric flow rate (V& ) of 7.8 m3 h-1 is considered. The system
`consists of a mini-channel evaporator, a mini-channel condenser, a compressor, and a thermostatic
`expansion device. In order to use the existing simulation model (which consists of a tube-fin evaporator
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`model), a tube-fin evaporator was designed to match the performance curve provided by the manufacturer
`of the mini-channel evaporator. The geometrical characteristics of the evaporator and the condenser are
`provided in Table 2. For the purposes of this paper, the compressor is defined by a constant isentropic
`efficiency of 70 % and a constant volumetric efficiency of 100 %. The thermostatic expansion device
`establishes the refrigerant mass flow to be the value determined by the iteration process described
`previously.
`
`
`Paper No. 128
`
`
`
`
`
` Figure 1b. P-h State Diagram for R-134a.
`
` Figure 2a. P-h State Diagram for R-1234yf.
`
`
`
`
`
`
`
`
`
` Figure 1a. T-s State Diagram for R-134a.
`
` Figure 2a. T-s State Diagram for R-1234yf.
`
`Table 1. Correlations used for the evaporator and the condenser.
`
`
`Refrigerant-side heat transfer coefficient for the evaporator
`Refrigerant-side heat transfer coefficient for the condenser
`Refrigerant-side pressure drop for the evaporator
`Refrigerant-side pressure drop for the condenser
`Refrigerant void fraction
`U-bend pressure drop
`Air-side heat transfer coefficient for the evaporator
`Air-side heat transfer coefficient for the condenser
`
`Kattan et al. (1998)
`Cavallini et al. (2006)
`Friedel (1979)
`Cavallini et al. (2006)
`Rouhani and Axelsson (1970)
`Equivalent length = 30D
`Manufacturer supplied data
`Chang & Wang (1997)
`
`
`
`
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`Table 2. Geometrical description of the mini-channel condenser and the micro-fin tube evaporator.
`
`Paper No. 128
`
`
`
`
`
`Micro-fin tube evaporator
`Longitudinal tube spacing1
`12.7 mm
`Transverse tube spacing2
`21.0 mm
`Tube length
`250 mm
`Inside tube diameter
`6.95 mm
`Number of rows
`4
`Number of tubes per row
`12
`Number of refrigerant circuits
`6
`Fin spacing
`1.8 mm
`Fin thickness
`0.11 mm
`Simulation elements per tube
`10
`Ratio of external to internal area
`12.1
` 1parallel to air flow 2normal to air flow
`
`
`
`
`
`
`
`
`Mini-channel condenser
`Width of flattened tube
`Height of flattened tube
`Length of flattened tube
`Spacing between flattened tubes
`Number of flattened tubes
`Number of refrigerant passes
`Number of flattened tubes/pass
`# of minichannels/flattened tube
`Minichannel diameter
`Fin spacing
`Fin height
`Fin length
`Fin thickness
`Simulation elements/flattened tube
`Ratio of external to internal area
`
`16.51 mm
`1.65 mm
`630 mm
`10.54 mm
`44
`2
`36/8
`13
`0.787 mm
`1.2 mm
`8.9 mm
`16.51
`0.1 mm
`20
`8.67
`
`
`The operating conditions are typical of those for an automotive system and are provided in Table 3. Note:
`the condenser subcooling values were found by performing a series of simulations where the subcooling
`values were incrementally increased from zero. The subcooling values used in the remaining simulations
`(and reported in Table 3) are the ones that maximized the COP values.
`
`
`Table 3. Operating conditions.
`
`
`
`Micro-fin tube evaporator
`
`Mini-channel condenser
`5.0 °C (R-134a)
`Subcooling
`7.0 °C (R-1234yf)
`Inlet air temperature
`35.0 °C
`3150 m3 h-1
`Volumetric flow rate of air
`3.0 m s-1
`Face velocity of air
`Air-side heat transfer coefficient1
`142 W m-2 K-1
`
`
`
`5.0 °C
`Superheat
`35.0 °C
`Inlet air temperature
`60.0 %
`Inlet air relative humidity
`500 kg h-1
`Mass flow rate of air
`4.0 m s-1
`Face air velocity
`Air-side heat transfer coefficient1 115 W m-2 K-1
`1Effective value, inclusive of fin efficiency
`
`
`4. SIMULATION RESULTS FOR A “REAL” CYCLE
`
`4.1 Performance with Constant Compressor Volumetric Capacity
`Table 4 provides simulation results for R-134a and R-1234yf for a constant compressor V& of 7.8 m3 h-1,
`that is, they represent a drop-in replacement of R-134a with R-1234yf. Several interesting observations
`can be made. First, the compressor discharge temperature for R-1234yf is approximately 6.0 °C less than
`the R-134a value. Second, the cooling capacity for R-1234yf is 2.0 % less than the R-134a value. Third,
`refm&
`for R-1234yf is some 20 %
`the COP for R-1234yf is 1.0 % less than the R-134a value. Fourth,
`greater than the R-134a value due to the denser suction vapor of R-1234yf. Fifth, the evaporation
`temperature of the R-1234yf system is 0.51 °C greater than the R-134a value, and the condensation
`temperatures of the R-1234yf system is 0.19 °C greater than the R-134a system. Despite these differences,
`the results confirm the attractiveness of R-1234yf as a potential replacement for R-134a in automotive
`applications. In the next sections, simulation results will be presented first for constant cooling capacity
`and then secondly for equal COP values.
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`Table 4. R-134a and R-1234yf simulation results for constant compressor V& of 7.8 m3 h-1.
`
`
`
`R-134a
`R-1234yf
`
`refm&
`
`(kg h-1)
`144.1
`173.1
`
`Qevap
`(W)
`5832
`5715
`
`Tevap
`(°C)
`9.66
`10.17
`
`Tdisch
`(°C)
`67.47
` 61.37
`
`Qcond
`(W)
`7253
`7141
`
`Tcond
`(°C)
`50.33
`50.52
`
`COP
`
`4.06
`4.02
`
`
`
`
`
`4.2 Performance with Constant Cooling Capacity
`The objective of this subsection is to compare the two systems at constant cooling capacities. The results
`are presented in Table 5, where the cooling capacity range was chosen to approximately cover compressor
`speeds from idle to highway driving. Several interesting observations can be made. First, the compressor
`discharge temperatures for R-1234yf are approximately 3.9 °C to 7.6 °C less than the R-134a values.
`refm&
`values for R-
`Second, the COPs for R-1234yf are 0 % to 4.3 % lower than the R-134a values. Third,
`1234yf are some 20 % to 26 % greater than the R-134a values due to the denser suction vapor of R-
`1234yf. Fourth, the evaporation temperatures of the R-1234yf system are from 0.17 °C smaller to 0.41
`°C greater than the R-134a values, while the condensation temperatures of the R-1234yf system are from
`0.49 °C to 0.95 °C greater than the R-134a values.
`
`Table 5. R-134a and R-1234yf simulation results for constant cooling capacities.
`
`refm&
`
`(kg h-1)
`75.6
`92.1
`100.3
`121.0
`144.1
`178.3
`162.4
`202.5
`188.2
`237.2
`
`Qevap
`(W)
`
`3364
`
`4276
`
`5832
`
`6418
`
`7202
`
`Tevap
`(°C)
`16.85
`17.20
`14.27
`14.68
`9.66
`9.72
`7.73
`7.64
`4.83
`4.66
`
`Tdisch
`(°C)
`57.63
`53.78
`62.31
`56.55
`67.47
`61.80
`71.31
`63.96
`74.63
`66.99
`
`Qcond
`(W)
`3865
`3875
`5102
`5061
`7253
`7332
`8170
`8268
`9469
`9590
`
`Tcond
`(°C)
`44.28
`44.85
`46.48
`46.97
`50.33
`51.01
`51.84
`52.68
`54.07
`55.02
`
`COP
`
`6.71
`6.59
`5.44
`5.44
`4.06
`3.92
`3.65
`3.49
`3.15
`3.03
`
`
`
`
`
`
`
`R-134a
`R-1234yf
`R-134a
`R-1234yf
`R-134a
`R-1234yf
`R-134a
`R-1234yf
`R-134a
`R-1234yf
`
`4.3 Performance for Equal COP Values
`Finally, in this subsection, the two systems are compared at equal energy efficiencies, with the results
`being presented in Table 6. Several interesting observations can be made. First, the compressor
`discharge temperatures for R-1234yf are approximately 6.4 °C to 8.6 °C greater than the R-134a values.
`Second, the cooling capacities for R-1234yf are 1.1 % to 4.4 % lower than the R-134a values. Third,
`refm&
`values for R-1234yf are some 14 % to 21 % greater than the R-134a values due to the denser suction
`vapor of R-1234yf. Fourth, the evaporation temperatures of the R-1234yf system are from 0.38 °C to
`1.02 °C greater than the R-134a values, while the condensation temperatures of the R-1234yf system are
`from 0.03 °C to 0.41 °C greater than the R-134a values.
`
`
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`Table 6. R-134a and R-1234yf simulation results for equal COP values.
`
`Paper No. 128
`
`Qcond
`(W)
`3865
`3654
`5102
`4977
`7253
`7178
`8170
`7993
`9469
`9075
`
`Tcond
`(°C)
`44.28
`44.46
`46.48
`46.83
`50.33
`50.74
`51.84
`52.19
`54.07
`54.10
`
`COP
`
`6.71
`
`5.44
`
`4.06
`
`3.65
`
`3.15
`
`Tdisch
`(°C)
`57.63
`59.66
`62.31
`62.73
`67.47
`68.02
`71.31
`70.60
`74.63
`74.60
`
`
`
`Tevap
`(°C)
`16.85
`17.69
`14.27
`14.85
`9.66
`10.04
`7.73
`8.24
`4.83
`5.85
`
`5. CONCLUSIONS
`
`refm&
`
`(kg h-1)
`75.6
`86.5
`100.3
`118.9
`144.1
`174.6
`162.4
`195.5
`188.2
`223.3
`
`Qevap
`(W)
`3364
`3217
`4276
`4230
`5832
`5714
`6418
`6237
`7202
`6918
`
`
`
`
`
`R-134a
`R-1234yf
`R-134a
`R-1234yf
`R-134a
`R-1234yf
`R-134a
`R-1234yf
`R-134a
`R-1234yf
`
`Simulations were conducted using R-1234yf in a typical baseline R-134a small-size European automotive
`air conditioning system, where the baseline R-134a system has a nominal cooling capacity of 5.8 kW at a
`compressor volumetric flow rate of 7.8 m3 h-1. If R-1234yf is used as a drop-in replacement in this
`baseline system, its cooling capacity is 2.0 % lower than the R-134a value, and its COP is 1.0 % lower
`than the R-134a value. If on the other hand, the two systems are compared at equal cooling capacities, the
`COP values of the R-1234yf system are 0 % to 4 % lower than the R-134a values over operating
`conditions from idle to highway speeds.
`
`While the evaporation and condensation temperatures are somewhat different between the two systems,
`the largest temperature differences can be seen in the compressor discharge temperatures. If R-1234yf is
`used as a drop-in replacement in the baseline system, its compressor discharge temperature is 6.0 °C
`lower than the R-134a value. If on the other hand, the two systems are compared at equal cooling
`capacities, the compressor discharge values of the R-1234yf system are 4.9 °C to 7.6 °C lower than the R-
`134a values over operating conditions from idle to highway speeds. These lower compressor discharge
`temperatures should lead to greater compressor reliability for the R-1234yf system.
`
`While both systems would benefit from the use of a liquid-line/suction-line heat exchanger, R-1234yf
`would benefit somewhat more from its use than would R-134a. Also, R-1234yf could benefit from the
`optimization of the heat exchanger circuitries.
`
`The thermodynamic and transport properties of R-1234yf are estimated from Brown et al. (2009a).
`
`The simulation results and analyses presented in this paper demonstrate the attractiveness of R-1234yf as
`a potential replacement for R-134a in automotive applications.
`
`ACKNOWLEDGEMENTS
`
`The authors thank Mr. Giuseppe Romano for his assistance with the simulations. J. Steven Brown
`expresses his appreciation to the Dipartimento di Fisica Tecnica of the Università di Padova for hosting
`him for a sabbatical stay during the 2008-2009 academic year.
`
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`NOMENCLATURE
`
`COP Coefficient of Performance
`m& mass flow rate (kg h-1)
`Q
` heat transfer rate (W)
` specific entropy (kJ kg-1 K-1)
`s
`VCC Volumetric Cooling Capacity (kJ m-3)
`
`Subscripts
`cond condenser or condensation
`evap evaporator or evaporation
`
`h
`P
`T
`V&
`
`specific enthalpy (kJ kg-1)
`pressure (kPa)
`temperature (°C)
`volumetric flow rate (m3 h-1)
`
`disch compressor discharge
`ref
`refrigerant
`
`
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`
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`3rd IIR Conference on Thermophysical Properties and Transfer Processes of Refrigerants, Boulder, CO, 2009
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