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Simulative analysis of secondary loop automotive refrigeration systems operated with an HFC and carbon dioxide.

ABSTRACT

Recent attempts to find energy-efficient thermal management systems for electric and plug-in hybrid electric vehicles have led to secondary loop systems as an alternative approach to meet dynamic heating and cooling demands and to reduce refrigerant charge. The choice of refrigerant for the primary refrigeration cycle is an important issue regarding the overall system performance. In this work, an HFC refrigerant (R-134a) and a natural refrigerant (R-744) are evaluated regarding a potential use in secondary loop systems. To meet the demands of R-744 cycles such as higher system pressure, most components have to be redeveloped. Nonetheless the use of the environmentally friendly refrigerant has advantages such as better applicability and performance in heat pump systems under cold ambient conditions. This work presents Modelica-based design and simulations of secondary loop automotive refrigeration systems with the HFC refrigerant in a subcritical and R-744 in a transcritical primary cycle. Under defined use cases and environmental conditions, both primary cycles are compared resulting in a better performance for HFC in a summer scenario and for R-744 under winter conditions.

CITATION: Menken, J., Ricke, M., Weustenfield, T., and Koehler, J., "Simulative Analysis of Secondary Loop Automotive Refrigeration Systems Operated with an HFC and Carbon Dioxide," SAE Int. J. Passeng. Cars - Mech. Syst. 9(1):2016, doi:10.4271/2016-01-9107.

INTRODUCTION

As secondary loop systems become a viable alternative approach for automotive thermal management systems, fundamental design requirements concerning the primary vapor compression cycle need to be specified. An important influencing factor is the choice of refrigerant. This simulative work compares two different primary refrigerant cycles of a secondary loop system: The first system is equipped with R-134a which is widely used throughout the automotive industry. The second system uses R-744 (C[O.sub.2]) as refrigerant. Due to the fact that both systems exhibit fundamental differences in terms of design and operation, the focus of this paper lies in providing comparability if procurable. This is achieved by presuming equal heating or coolant capacities as well as justifiable assumptions for modification based on state-of-the-art technologies for design. For the system evaluation, real-world driving cycles and representative weather data are taken into account.

The general concept of a secondary loop refrigeration system can be found in various fields of application. Wang et al. [1] offer an extensive review on secondary loop refrigeration systems including performance and risk assessment of fammable refrigerants. Applications belonging to the fields of commercial refrigeration, residential air conditioning/heat pumping and mobile air conditioning are presented. Recently, much research has been performed regarding automotive applications in this field. Ghodbane et al. [2] present a secondary loop system with the HFC refrigerant R-152a. Kowsky et al. [3] demonstrate a hermetic encapsulated central thermal management unit for the use in electric and hybrid vehicles using the battery waste heat as energy source during heat pump mode. The occurring energy storage effects in such systems can even be increased by the addition of phase change materials [4] or ice-based latent energy storages [5] in the secondary loops.

Due to the high global warming potential (GWP) of R-134a, the European Union's directive for emissions from air-conditioning systems in motor vehicles constrain the use of such fuorinated greenhouse gases [6]. It is specified that new vehicle models have to be fitted with a refrigerant having a GWP less than 150.

A climate-friendly alternative is the refrigerant R-744 (C[O.sub.2]) with a GWP of 1. Due to its low critical temperature of [T.sub.crit] = 31.1 [degrees]C the R-744 refrigerant cycle discharges heat at a transcritical level. Because of the fact that the cycle is operated at a higher pressure level, R-744 cannot be used as a drop in refrigerant. This is why all refrigerant components have to be engineered from scratch. As shown in Table 1 the volumetric refrigeration capacity of R-744 is 8 times higher compared to R-134a. Therefore a smaller mass flow rate is needed to achieve a certain cooling load which results in a reduced stroke volume of the compressor. Another advantage is based on the low evaporation temperature. Thus, a R-744 equipped heat pump system can be operated on lower ambient temperatures than a R-134a one. As the density declines with pressure, a higher mass flow rate can be obtained in the C[O.sub.2]-system.

METHODS

Boundary Conditions

As the requirements of the vehicle thermal management system change over the year depending on the environmental conditions, the system evaluation is not a trivial task. Additionally, the driving behavior plays an important role. In order to use representative data describing operating conditions of the evaluated system, the boundary conditions have to be specified. As this work uses the waste heat of electric drivetrain components (battery, electric machine, power electronics), real world driving cycles have to be taken into account. Additionally, the environmental conditions influence the vehicle's transient behavior. Therefore, the ambient temperature [T.sub.air,amb], relative humidity [[phi].sub.air,amb] and solar radiation [q.sub.sol] need to be specified. Describing the customer's driving behavior, the Artemis road driving cycle is used. Figure 1 shows the velocity profiles over time [7] as well as the calculated drivetrain losses [8].

Regarding the ambient conditions, Table 2 displays the values for [T.sub.air,amb], [[phi].sub.air,amb] and [q.sub.sol] for the use cases 1, 2 and 3 which represent the three typical heating or cooling loads of the passenger compartment that lead to the dynamic cabin heating and cooling loads described in the following section.

Dynamic Heating and Cooling Loads

For the evaluation of the system configurations, a battery powered electric vehicle is assumed. Because of the lack of an internal combustion engine, the missing heat load during winter and reheat operation mode is provided by a primary refrigerant cycle operated in a heat pump mode using the waste heat of the electric drivetrain components and an electric heater if necessary as heat source. The reheat operation mode is distinguished by dehumidification of the instream air and subsequent re-heating before entering the passenger compartment. Based on experiments and preliminary simulation, the cooling and/or heating load at coolant-to-air heat exchanger is predefined as shown in Figure 2. Positive heat flow rates represent a heat-up of the instream air, whereas a negative sign indicates the air side cooling demand. It has to be mentioned that during re-heat operation, only the cooling demand at the first heat exchanger is controlled. It is assumed that sufficient heating load for the second cabin heat exchanger is provided by a heat surplus at the condenser.

Modeling and Simulation

Refrigeration Cycle

All system setups presented in this work are modeled with the TIL Library in Dymola/Modelica. They consist of an electrical scroll compressor, a condenser, an electrical expansion valve, an evaporator and a refrigerant reservoir on the low pressure side (accumulator). This setup is described as optimal from an operating point of view as subcooling can be set while superheating is close to zero [9]. The simulation of the R-134a model is validated in a previous publication [10]. The R-744 cycle additionally contains an internal heat exchanger. Figure 3 shows the setup of the two primary loop refrigerant cycles modeled in Dymola.

Compressor

The compressor is a crucial component regarding the overall energy consumption of the system. Both refrigerant cycles are driven with an electric loss based scroll compressor described in [11]. The parameters based on measurements of an electric R-134a scroll compressor and then extrapolated for the desired maximum cooling load are given in Table 3. As R-134a and R-744 have different specific volumetric heat capacities, the suction volumes have to be chosen respectively (based on the ratio of suction volumes of mechanical swash-plate-piston compressors) [12].

Plate Heat Exchanger

Plate heat exchangers with their compact dimensions and typically high heat flux densities are used as refrigerant-to-coolant heat exchangers. Figure 4 shows the general flow distribution of both mediums. Due to the unavailability of high pressure C[O.sub.2] plate heat exchangers for automotive applications, the dimensions, pressure losses and heat transfer coefficients are estimated. To maintain comparability the R-134a heat exchangers are also modeled with constant coefficients instead testbed tuned correlations for pressure drop and heat transfer.

Both refrigerant cycles are sized for a cooling capacity of 7.5 kW at ambient conditions of T = 40 [degrees]C and [phi] = 40 % (Table 4).

In a study by Sarkar et al. the ratio between gas cooler area and evaporator area is analyzed. For simultaneous heating and cooling of water with a R-744 cycle the optimum ratio ranged from 1.6 to 1.9. In the simulation model the ratio 1.6 is used for the plate length. Therefore the number of plates is set equal in gas cooler and evaporator. The estimated constant values for pressure drop and heat transfer coefficient for the refrigerant and coolant side are shown in Table 4. Since the power consumption of the coolant pumps is not evaluated in this work, the pressure drop on the plate heat exchanger's coolant side is neglected. In the comparison by Mastrullo et al. [14] it can be seen that the heat transfer coefficient in the R-744 evaporator is approximately twice as high compared to a R-134a evaporator. Regarding the refrigerant side pressure drop within the heat exchanger, fixed values are assumed. The values are in an order of magnitude comparable to experimental results obtained by McEnaney et al. [12]. With this configuration both refrigerant systems are equipped with 40 plates per heat exchanger to provide the demanded cooling capacity.

Expansion Valve

The electric expansion valve restricts the refrigerant flow according to a given pressure drop with the Bernoulli-equation.

[MATHEMATICAL EXPRESSION NOT REPRODUCIBLE IN ASCII] (1)

Accumulator

Both primary cycles are equipped with a refrigerant reservoir on the low pressure side. Due to the design of the accumulator and the oil bleed hole at the bottom of the inner J-tube, the vapor quality of the refrigerant at the accumulator outlet is unsaturated gas [15]. However the outlet vapor quality depends on flow conditions and could be examined in greater detail. As a simplification a constant outlet vapor quality of [x.sub.ref,accu,out]= 0.92 is assumed in all simulations.

Control Strategy

The remaining two degrees of freedom of a developed refrigerant cycle are the compressor power and the opening of the expansion valve [9]. Assuming given flow rates for the secondary fluids at the heat exchangers, the compressor speed is used to fulfll load requirements while the electronic expansion valve (EEV) controls an operational variable. In this work the rotational speed of the compressor ensures the coolant-side demanded heating or cooling load (depending on the use case) at the respective plate heat exchanger. The EEV adjusts either the refrigerant-side subcooling at the R-134a-condenser [16] or the outlet pressure at the R-744-gascooler

For the R-134a system the subcooling is set to [T.sub.sub] = 5 K. As the temperature during condensation stays constant, the refrigerant-side pressure settles. However, as pressure and temperature are independent in a transcritical cycle, the high pressure has to be set. By adjusting the effective flow area of the expansion valve a controller can set the pressure. Correlations for the optimal high pressure as a function of refrigerant-side gascooler outlet and evaporation temperature are given by Liao et al. [17] and Sarkar et al. [18] in equation 2 and 3, respectively:

[P.sub.hp,opt,Liao] = (2.778 - 0.0157 * [T.sub.ref,chi,out]) * [T.sub.re f,chi,out] + (0.381 * [T.sub.ref,chi,out] - 9.34) (2)

[MATHEMATICAL EXPRESSION NOT REPRODUCIBLE IN ASCII] (3)

Complete Vehicle

The simulation of the complete vehicle is done under different ambient conditions and driving cycles. Exemplary, interconnections between all components in a winter and summer case are shown in Figure 5. If the cabin has to be cooled, the cabin heat exchanger is connected with the secondary loop on the evaporator side. After passing the cabin heat exchanger, the coolant flows through the battery in order to maintain the operating temperature limits of this component.. On the hot side, the ambient heat exchanger discharges the heat of the condenser/gas cooler, electrical machine and power electronics. In the winter case the cabin heat exchanger is connected to the condenser/gas cooler side of the primary cycle. In this case the hot water leaving the condenser/gas cooler is used to heat up the instream air to the passenger's compartment. All electric components discharge their heat on the evaporator's side to enhance the performance of the heat pump cycle. By controlling the compressor speed, the given heat flow rate shown in Figure 2 is given as a transient set point.

Electric Drivetrain Components

The electric drivetrain components (battery, electrical machine and power electronics) are modeled as zero dimensional heat capacity. Thus there is no temperature gradient within the component. The capacity is connected with two different heat flow rates. On the one hand, the capacity is thermally connected with the coolant. On the other hand there is the waste heat caused by the electrical flow. The sum of all heat flows is equal to the components heat capacity times the temperature derivation:

[MATHEMATICAL EXPRESSION NOT REPRODUCIBLE IN ASCII] (4)

Coolant-to-Air Heat Exchangers

There are three coolant-to-air heat exchangers needed in the secondary cycles. The ambient heat exchanger in the vehicle's front-end discharges or absorbs heat to/from the environment. In addition the electrical drivetrain components can be cooled passively without the connection to the primary cycle. Two coolant-to-air heat exchangers act as cabin heat exchangers. Thereby the cabin can be cooled, heated or dehumidificated in the reheat mode.

RESULTS

Optimal High Pressure Control

Figure 6 shows the characteristics of the COP as a function of the high pressure and varying coolant-side gascooler inlet temperatures. The operation points with the maximum COP are connected with a dotted line. The efficiency significantly depends on the high pressure, whereby descending the optimal high pressure leads to a radical loss of efficiency at constant coolant inlet temperature. The efficiency loss when exceeding the optimal high pressure is notable smaller.

Comparing the optimal operation points determined by the experimental design with mathematical correlations, Figure 7 shows the different methods. The optimal high pressure calculated with the correlations presented by Liao et al. and Sarkar et al. is in most operation points too small. It can be seen that a simple linear approximation results in a good fit to the actual optimal high pressure:

[P.sub.hp,opt] = 2.08 * Tref,igc,out + 18.7 (5)

Limiting Operation Conditions

As the refrigerant cycle should not be operated below ambient pressure, the low suction pressure of R-134a heat pumps can be a problem during winter operation. The R-134a heat pump cycle can not be operated with lower evaporation pressures than [p.sub.evap] = 1.013 bar respectively with lower evaporation temperatures than [T.sub.evap] = -26.08 [degrees]C. In a simulation, Q = 4000 is maintained with both R-134a and R-744 cycles. Meanwhile the coolant inlet temperature of the evaporator is decreased down to [T.sub.col,chi,in] = -30 [degrees]C. The decreasing refrigerant suction pressure during the temperature drop can be seen in Figure 8. At [T.sub.col,chi,in] = -15 [degrees]C the limiting pressure is reached in the R-134a system and the operation has to be stopped on lower inlet temperatures. As opposed to this the R-744 cycle is operating with an evaporation pressure of [p.sub.R744] = 19.4 bar at this point. An operation under lower temperatures is achievable without any problems.

Evaluation of Use Cases

To compare both systems, the complete vehicle (as shown in Figure 5) is simulated. To compare the R-134a and R-744 system the total energy consumption of the compressor after 1000 seconds of operation is calculated. The results for the Artemis road cycle as boundary condition can be seen in Figure 9.

During Re-heat mode the compressor of both systems is operating with minimum speed, the energy consumption is approximately equal. During cooling mode at [T.sub.amb] = 30 [degrees]C the consumption of the R-744 compressor is 94 Wh higher than the consumption of the R-134a system. In contrast to this, a lower energy consumption can be seen with the R-744 system in heating mode. Attention should be paid to the different system setup. In winter the R-134a system uses a coolant-side high voltage PTC heater with a const. [P.sub.el,PTC] = 1000 W ([eta] = 0.96) for preheating the coolant on the evaporator side. In total the R-134a system's energy consumption after 1000 seconds in the Artemis road cycle is [DELTA]E = 433 Wh higher for tempering the car cabin. Even without the PTC heater the R-134a compressor's energy consumption is higher than the R-744 one. In addition Figure 9 shows the reduction of the relative range due to the operation of the compressor for every operation mode. For heating mode the consumption of the PTC heater is included. During the first 1000 seconds in the Artemis road cycle a distance of 16.36 km is driven. A specific range of 6.6 km/kWh for the electric car is assumed. In the winter case the operation of a R-134a heat pump can reduce the range by 28 %. The range reduction of 10 % with a R-744 heat pump is notable smaller. In the summer case the range is reduced by 8 % operating a R-134a compressor and by 12 % by operating a R-744 system.

CONCLUSION

In this work, an automotive secondary loop system with two different primary cycles (one for R-134a and another for R-744) is modeled and simulated. Therefore, simplifications are made to maintain a comparability of both systems. Typical ambient conditions are assumed as well as realistic driving cycles and component heat loads were presented. To meet the requirements of operating transcritical cycles, different control strategies are compared and a linear correlation for determining an optimal high pressure is presented. It can be seen that a R-744 operated primary cycle can accomplish heat and cooling demands under all conditions. For both R-134a and R-744 systems the minimal operating temperatures are evaluated. As part of a complete vehicle simulation the system's energy consumption is calculated for distinctive use cases. Significant differences in winter and summer case are pointed out.

In comparison to the conventional R-134a cycle the R-744 system seems to be a competitive alternative. The system is able to provide enough heating power with a satisfying COP without using an additional electric heater. In cooling mode at hot ambient temperatures the energy consumption is in fact higher with the R-744 system. But the energy saving potential in heat pump mode is notably larger at the statistically more often occurring low and moderate ambient temperatures.

For further simulations the system is prepared to evaluate the operation under all kinds of situations. This includes other ambient conditions and driving cycles, as well extremely harsh conditions. Also the complete system's energy consumption including all secondary pumps and fans has to be evaluated. Additionally, the driving cycles could be comprised into daily driving patterns [19]. Thereby effects due to heat storage in secondary loops can be observed. Furthermore the use of alternative refrigerants could be evaluated.

Another important step is the validation of the models in the R-744 cycle. The simulations in this paper are based on theoretical assumptions without running similar test bench trials as it was done for the cycle operated with R-134a. For modelling a cycle close to reality, detailed measurement data for the plate heat exchangers as well as for the electrical C[O.sub.2] scroll compressor is necessary.

REFERENCES

[1.] Wang, K., Eisele, M., Hwang, Y., and Radermacher, R., "Review of secondary loop refrigeration systems," International Journal of Refrigeration 33(2):212-234, 2010, doi:10.1016/j.ijrefrig.2009.09.018.

[2.] Ghodbane, M., Craig, T.D., and Baker, J.A., "Demonstration of an Energy-Efficient Secondary Loop HFC-152a Mobile Air Conditioning System," 2007.

[3.] Kowsky, C., Wolfe, E., Leitzel, L., and Oddi, F., "Unitary HPAC System," SAE Int. J. Passeng. Cars - Mech. Syst. 5(2):1016-1025, 2012, doi:10.4271/2012-01-1050.

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[12.] McEnaney, R.P., Boewe, D.E., Yin, J.M., Park, Y.C. et al., "Experimental Comparison of Mobile A/C Systems When Operated With Transcritical C[O.sub.2] Versus Conventional R134a," 7th International Refrigeration and Air Conditioning Conference at Purdue, 1998.

[13.] Sarkar, J., Bhattacharyya, S., and Ram Gopal, M., "Simulation of a transcritical C[O.sub.2] heat pump cycle for simultaneous cooling and heating applications," International Journal of Refrigeration 29, 2006, doi:10.1016/j.ijrefrig.2005.12.006.

[14.] Mastrullo, R., Mauro, A.W., Rosato, A., and Vanoli, G.P., "Comparison of R744 and R134a heat transfer coefficients during flow boiling in a horizontal circular smooth tube," International Conference on Renewable Energies and Power Quality, Valencia, 2009.

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Jan Christoph Menken, Martin Ricke, and Thomas A. Weustenfield

AUDI AG

Juergen Koehler

University of Braunschweig

CONTACT INFORMATION

Jan Christoph Menken

AUDI AG, Ingolstadt, Germany

jan-christoph.menken@audi.de

DEFINITIONS/ABBREVIATIONS

A - Area

accu - Accumulator

amb - Ambient

batt - Battery

chi - Chiller

col - Coolant

COP - Coefficient of performance

[c.sub.p] - Specific heat capacity

E - Energy

EEV - Electric expansion valve

emch - Electric machine

evap - Evaporator

GWP - Global warming potential

HFC - Hydrofuorocarbons

hp - High pressure

igc - Indirect gas cooler

m - Mass

m - Mass flow rate

opt - Optimal

p - Pressure

pwel - Power electronics

Q - Heat flow rate

ref - Refrigerant

sol - Solar

sub - Subcooling

t - Time

T - Temperature

V - Volume

x - Vapor quality

[rho] - Density

[phi] - Relative humidity

Table 1. Comparison of selected parameters for refrigerants R-134a and
R-744
[20, 21, 22, 23].

                              R-134a               R-744

GWP                            1300                1
Autoignition Temperature       743 [degrees]C      none
Cost                           20-30 [euro]/kg     10[euro]/kg
Critical Pressure              40.7 bar            73.8 bar
Critical Temperature          101.1 [degrees]C     31.1 [degrees]C
Boiling Temperature at 1 bar  -26.4 [degrees]C    -78.4 [degrees]C
Volumetric Refrigeration
Capacity at 0
[degrees]C                    2868 kJ/[m.sup.3]   22545 kJ/[m.sup.3]

Table 2. Ambient conditions for the three use cases heat-up, re-heat
and
cool-down.

                              Heat-up  Re-heat  Cool-down

[T.sub.air,amb] [[degrees]C]   5       15        30
[[phi].sub.air,amb] [%]       70       70        42
[q.sub.sol] [W/[m.sup.2]]      0        0       800
[m.sub.col.igc.in] [kg/s]      0.15     0.15      0.15
[m.sub.col.chi.in] [kg/s]      0.15     0.15      0.15

Table 3. Selected compressor parameters for R-134a and R-744.

                                    R-134a  R-744

Suction Volume [cm.sup.3]           34      5.49
Effective Leackage Area [mm.sup.2]   0.028  0.0035
Electric Efficiency [-]              0.764  0.88

Table 4. Plate Heat Exchanger geometry data of R-134a and R-744
condenser/gas
cooler and evaporators.

                                         Condenser   Evaporator
                                                 R-134a
Number of parallel flows per medium [-]    20          20
Total Heat Transfer Area [[m.sup.2]]        0.84954     0.799901
Refrigerant-Side Heat
Transfer Coefficient [W/[m.sup.2]K]      2000        2000
Coolant-Side Heat Transfer
Coefficient [W/[m.sup.2]K]               4000        4000
Refrigerant-Side Pressure Drop [bar]        0.5         0.6
Coolant-Side Pressure Drop [bar]            0           0
                                                 R-744
Number of parallel flows per medium [-]    20          20
Total Heat Transfer Area [[m.sup.2]]        0.753358    0.428084
Refrigerant-Side Heat
Transfer Coefficient [W/[m.sup.2]K]      4000        4000
Coolant-Side Heat
Transfer Coefficient [W/[m.sup.2]K]      4000        4000
Refrigerant-Side Pressure Drop [bar]        2           1
Coolant-Side Pressure Drop [bar]            0           0
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Author:Menken, Jan Christoph; Ricke, Martin; Weustenfield, Thomas A.; Koehler, Juergen
Publication:SAE International Journal of Passenger Cars - Mechanical Systems
Article Type:Report
Date:Apr 1, 2016
Words:4435
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