Purdue University
`Purdue e-Pubs
`International Refrigeration and Air Conditioning
`Conference
`
`School of Mechanical Engineering
`
`2000
`
`Suction Line Heat Exchanger for R134A
`Automotive Air-Conditioning System
`M. Preissner
`University of Maryland
`
`B. Cutler
`University of Maryland
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`R. Radermacher
`University of Maryland
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`C. A. Zhang
`Visteon Automotive Systems
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`Follow this and additional works at: http://docs.lib.purdue.edu/iracc
`
`Pre ssner M.; Cutler B.; Radermacher R.; and Zhang C. A. "Suct on L ne Heat Exchanger for R134A Automot ve A r Cond t on ng
`System" (2000). International Refrigeration and Air Conditioning Conference. Paper 494.
`http://docs.l b.purdue.edu/ racc/494
`
`Th s docu e as bee ade ava ab e
`add o a
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`a o .
`Co p e e p oceed gs ay be acqu ed
`He
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`oug Pu due e-Pubs, a se v ce o e Pu due U ve s y L b a es. P ease co ac epubs@pu due.edu o
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` p
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` a d o CD-ROM d ec y o e Ray W. He
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`ck Labo a o es a ttps://e g ee
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`g.pu due.edu/
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`Arkema Exhibit 1129
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`SUCTION LINE HEAT EXCHANGER
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`FOR R134A AUTOMOTIVE AIR-CONDITIONING SYSTEM
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`Marcus Preissner
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`Brett Cutler
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`Reinhard Radermaeher
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`Center of Environmental Energy Engineering
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`Department of Mechanical Engineering
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`University of Maryland
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`College Park, MD 20742
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`TEL: 301-405-5003, FAX: 301-405-5247
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`e-mail: mpreissn@eng.umd.edu
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`Chao A. Zhang
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`Visteon Corporation
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`45000 Helm St.
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`Plymouth, NH 48170
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`TEL: 734—45l—8857, FAX: 734-416-6940
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`e-mailz czhang3 @visteon.com
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`ABSTRACT
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`The performance of an R134a automotive prototype air-conditioning system was tested in
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`the laboratory with and without an internal heat exchanger. At a higher condenser air
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`temperature of 40°C and a restrictive idling air flow rate of 1.0 m/s, the COP and the capacity
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`increased on the order of 5 to 10 % with a suction line heat exchanger with 60 % effectiveness. If
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`the system is operated not at the optimum expansion device setting or it is undercharged, the
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`improvement
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`exchanger, the low side pressure drop is critical and must be minimized so as not to compensate
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`the performance improvement.
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`INTRODUCTION
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`The performance of R134a automotive air-conditioning systems has been continuously
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`improved over the last years. However, the potential benefits associated with the implementation
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`of a suction line heat exchanger (SLHX) - also termed internal heat exchanger — have not been
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`paid much attention according to the published literature. A SLHX transfers energy from the
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`refrigerant leaving the condenser to the suction gas, resulting in a lower inlet enthalpy at the
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`evaporator, providing a higher cooling capacity. The competing effects with respect
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`Eighth International Refrigeration Conference at
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`Purdue University, West Lafayette, IN, USA - July 25-28, 2000
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`performance are a larger enthalpy change across the compressor and a lower mass flow rate, both
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`effects caused by the lower suction density. An Rl34a prototype system was operated with and
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`without a SLHX, the performance comparison is presented in this paper.
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`EXPERIMENTAL SETUP
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`The SLHX was tested in the Rl34a automotive air-conditioning test facility at the Center
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`of Environmental Energy Engineering (CEEE). The evaporator is placed in a closed loop, and an
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`air handling unit controls the specified temperature and humidity of the air stream entering the
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`evaporator. The air inlet and outlet temperatures are measured with a grid of 9 thermocouples
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`upstream and downstream of the evaporator, the humidity entering and leaving is determined by
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`two dew point meters. In addition, the moisture condensing on the coil is collected in order to
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`verify the dew point measurement. The air flow rate is varied with a variable speed fan, and is
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`calculated from the pressure drop across a nozzle in the loop. The air flow rate measurement was
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`calibrated with an electric heater prior to the beginning of the experiments. From the temperature
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`and humidity differences, and the air flow rate, the cooling capacity is calculated. The heat loss
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`from the experimental test section to the ambient is calibrated and incorporated in the capacity
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`calculation. A differential pressure transducer measures the air side pressure drop across the
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`The condenser is mounted in an open duct within a Variable climate environmental
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`chamber. The air flow rate,
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`measured, and calculated in the same manner as for the evaporator.
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`The schematic setup of the Rl34a loop is shown in Figure 1. The compressor is placed in
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`the same climate chamber as the condenser, and consequently, is exposed to the condenser air
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`inlet temperature (outdoor temperature). This condition is close to the arrangement in a vehicle,
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`where the compressor is exposed to slightly higher temperatures in the engine compartment. The
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`compressor is driven by an electric motor operated with a variable frequency drive. A non-
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`contact torque meter is mounted between the motor and the compressor, and together with the
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`measured compressor speed, the input power to the compressor is determined. The refrigerant
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`leaving the compressor passes the condenser and a mass flow meter, and can then be routed
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`through the suction line heat exchanger, or bypass it. The suction line heat exchanger transfers
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`heat to the suction gas, providing more cooling capacity and improving the cycle COP. After
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`passing the expansion device (manually or electronically controlled), the refrigerant evaporates
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`in the evaporator and enters the accumulator. Liquid refrigerant and oil are stored in this device —.
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`a small amount of the oil-refrigerant mixture passes through a bleed hole at the bottom inside of
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`the accumulator to the main refrigerant line to ensure proper compressor lubrication. The
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`refrigerant loop is equipped with instream thermocouples and pressure transducers at the inlet
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`and outlet of the main components.
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`Note that the system is neither a fixed orifice system nor a TXV (thermostatic expansion
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`device) arrangement. The superheat was varied using charge and expansion valve settings, and
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`the degree of subcooling resulted from the particular system test conditions and components. The
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`operation with a receiver, which controls the high side pressure and the degree of subcooling is
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`planned in fiiture tests with a new prototype.
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`Eighth International Refrigeration Conference at
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`Purdue University, West Lafayette, IN, USA — July 25-28, 2000
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`—W
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`—m Outdoor
`——@> Air Duct
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`—w
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`Gas Cooler
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`-——w
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`Expansion
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`Suction
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`Accumulator
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`Compressor
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`Figure 1: Schematic setup ofgthe R134a cycle
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`COMPONENTS AND TEST CONDITIONS
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`A state-of-the—art piston compressor and custom made heat exchangers were employed in
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`the R134a system. They are of comparable dimensions and performance when compared to
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`current systems, the actual dimensions are not of primary interest in this study. The evaporator
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`inlet temperature was set at 27°C with a relative humidity of 50 %, and the air flow rate was
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`fixed at 580 ‘m3/h. The air velocity across the condenser was restricted to 1.0 rrds at a compressor
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`speed of 1000 RPM, and increased to 2.5 m/s at 1800 RPM. The air inlet temperature was varied
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`between 25 and 40°C (40 % humidity). These test conditions are summarized in Table 1.
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`Line Heat
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`Exchanger Evaporator
`Wt
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`Indoor
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`Humidity
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`Outdoor
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`Compressor
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`Frontal Air
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`idling: 1000
`idling: 1.0
`40
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`driving: 1800 driving: 2.5
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`Table 1: Test conditions
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`Eighth International Refrigeration Conference at
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`Purdue University, West Lafayette, IN, USA — July 25-28, 2000
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`Relative
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`RESULTS AND DISCUSSION
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`The remaining parameters to be adjusted during testing were the refrigerant charge and
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`the opening of the expansion device. The expansion device was controlled so that
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`evaporation temperature was held constant at various values depending on the compressor speed.
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`At 1000 RPM, the evaporation temperatures tested were 5°C, 8°C, and 12°C, where 8°C showed
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`the best performance. At 1800 RPM, the capacity to be transferred at the evaporator is larger,
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`therefore the evaporation temperatures tested were 2°C, 5°C, and 8°C, where 2°C showed the best
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`performance. It is not possible to run lower temperatures as freezing of moisture on the
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`evaporator coil must be avoided.
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`The performance of the system both with and without a suction line heat exchanger for a
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`compressor speed of 1000 RPM and a condenser air inlet temperature of 40°C is shown in
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`Figures 2 and 3. The data points to the left of the curves represent the performance of the system
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`operated at undercharged conditions. The refrigerant leaves the condenser as a 2-phase fluid for
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`higher evaporation temperatures (8°C), and with small amounts of subcooling at
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`evaporation temperatures. The superheat at these operating conditions is on the order of 15 K,
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`which is close to the air inlet temperature. The tests using a SLHX (with the same refrigerant
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`charge) show a larger capacity and COP. Note that the system volume remained the same for
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`tests with and without SLHX, as no system modifications were carried out, and only valve
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`settings were changed.
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`the
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`performance did not change significantly with increasing charge. For the higher evaporation
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`temperature of 8°C, the charge was increased until the superheat was about 5 K, resulting in a
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`subcooling at the condenser of about 20 K. At the lower evaporation temperature of 5°C, the
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`superheat was about 10 K and the subcooling 30 K, Note that the degree of superheat and
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`subcooling does not change significantly whether a SLHX is applied or not. For 8°C evaporation
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`temperature, the capacity improved by 8 % and the COP by 10 %. At lower charge or lower
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`evaporation temperatures (non-optimum operating points) - the benefit of the suction line heat
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`exchanger can be even larger.
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`the
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`At a compressor speed of 1800 RPM,
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`performance improvement with the suction line heat exchanger is only on the order of 5 %. At
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`lower condenser air inlet temperatures of 25°C at 1000 RPM, the performance with and without
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`SLHX was equivalent, however, at undercharged conditions and not optimum evaporation
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`temperatures, the improvement was 0 to 5 %. At 25°C condenser air inlet temperature and
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`1800 RPM, no performance increase was noticeable.
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`When designing a SLHX, the refrigerant pressure drop is a major parameter to consider.
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`The high pressure side is not very crucial as the pressure drop in the subcooled region is low due
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`to the high density of the liquid, and it is not difficult to design the heat exchanger with a low
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`pressure drop. Moreover, the small high side pressure drop does not change the refrigerant
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`temperature significantly. However it should be avoided that the refrigerant gets into a 2—phase
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`state, as the pressure drop of 2—phase fluid in the liquid line gets significantly larger. On the
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`suction side however, the pressure drop reduces the suction density resulting in a lower mass
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`flow rate, a higher pressure ratio, both leading to a lower system performance.
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`The heat exchanger used here was designed carefully so that the low side pressure drop is
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`negligible. It was below the measurement accuracy of 10 kPa in all operating conditions tested.
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`Eighth International Refrigeration Conference at
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`Due to the focus on the low pressure drop, the heat exchangers effectiveness ranged only from
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`The system was tested with a second SLHX. At the mass flow rates encountered in the
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`the low side pressure drop was between 40 and 70 kPa. Although having an
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`effectiveness of 60 %, the pressure drop is too high and annihilates the performance benefit
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`Bold symbols and solid lines:
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`vowpen symbols and dotted lines:
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`Without
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`4.5
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`4.0
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`Eis
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`TJ“d°°" = 27°C
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`T___0Utd00l' 7- 40°C
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`——
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`El
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`2500
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`l——
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`Condenser Outlet Pressure [kPa]
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`—G * 0
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`~n— 5°C
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`‘ ‘ ‘ ‘ El
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`I]
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`T_indoor = 27°C
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`-I-_outdo0r= 4006
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`"“"‘1‘ 21
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`Condenser Outlet Pressure [kPa]
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`Bold symbols and solid lines:
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`“gt” Smxb I
`d d tt d I,
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`pen sym 0 san
`o e
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`Figures 2 and 3: Cooling capacity and COP with and without SLHX
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`Eighth International Refrigeration Conference at
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`To further quantify the performance decrease associated with an additional pressure drop
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`at the low pressure side, calculations were carried out with a program based on fitted cycle
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`parameters from the experimental data set. For the same test conditions as the data presented in
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`Figures 2 and 3, the penalty of the pressure drop on the low pressure side of the suction line heat
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`exchanger is shown in Figure 4. A 50 kPa pressure drop reduces capacity and COP by about
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`4.5
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`Figure 4: Performance penalty of pressure drop at the suction side
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`CONCLUSIONS
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`Depending on the operating conditions and design, a suction line heat exchanger can
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`improve the performance of an automotive system. At higher air temperatures of 40°C at the
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`condenser and with a low air velocity of 1.0 m/s, which typically occurs in idling conditions, a
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`suction line heat exchanger can improve the capacity as well as the COP by about 5 to 10 %. At .
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`higher air velocities and lower temperatures, the benefit nearly vanishes. If the air-conditioning
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`system is operated at either not at the optimum evaporator temperature or with insufficient
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`charge (both of which can occur in a vehicle), the benefit of the suction line heat exchanger can
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`be even larger.
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`After having demonstrated that the performance improvements of a suction line heat
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`exchanger are not negligible, a further design with a higher efficiency is planned. However, the
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`pressure drop at the low pressure side is the most important parameter to consider in a suction
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`line heat exchanger design. With the new heat exchanger, a broader range of operating
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`conditions is planned to be tested.
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`ACKNOWLEDGMENTS
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`We greatly appreciate the support from Visteon and the University of Maryland.
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`Eighth International Refrigeration Conference at
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`Purdue University, West Lafayette, IN, USA - July 25-28, 2000
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`Capacity[kW],COP
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`2.5
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`-0- Capacity
`—EI— COP
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