Energy 27 (2002) 391–404
www.elsevier.com/locate/energy
The effects of improper refrigerant charge on the
performance of a heat pump with an electronic expansion
valve and capillary tube
J.M. Choi, Y.C. Kim
*
Department of Mechanical Engineering, Korea University, Anam-dong, Sungbuk-ku, Seoul 136-701, South Korea
Received 28 February 2001
Abstract
For inverter heat pumps and multi-type heat pumps, conventional expansion devices such as capillary
tubes, short tube orifices, and thermostatic expansion valves (TXVs) are being gradually replaced with
electronic expansion valves (EEVs) because of the increasing focus on comfort and energy conservation.
In this study, the effects of off-design refrigerant charge on the performance of a water-to-water heat pump
are investigated by varying refrigerant charge amount from ⫺20% to +20% of full charge in a steady state,
cooling mode operation with expansion devices of capillary tube and EEV. The characteristics of the heat
pump with an EEV are compared with those with a capillary tube. The capillary tube system is more
sensitive to off-design charge as compared with the EEV system. Cooling capacity and COP of the EEV
system show little dependence on refrigerant charge, while those are strongly dependent on outdoor conditions. In general, for a wide range of operating conditions the EEV system shows much higher performance as compared with the capillary tube system. The performance of the EEV system can be optimized
by adjusting the EEV opening to maintain a constant superheat at all test conditions.  2002 Elsevier
Science Ltd. All rights reserved.
1. Introduction
A heat pump is a major electrical energy consuming residential appliance. To make an energy
efficient heat pump system, the compressor should have high efficiency and be optimized well
with other parts. In addition, the amount of refrigerant charge in the heat pump is another primary
parameter influencing energy consumption. Undercharge or overcharge of refrigerant into the heat
* Corresponding author. Tel.: +82-2-3290-3366; fax: +82-2-921-5439.
E-mail address: yongckim@korea.ac.kr (Y.C. Kim).
0360-5442/02/$ - see front matter  2002 Elsevier Science Ltd. All rights reserved.
PII: S 0 3 6 0 - 5 4 4 2 ( 0 1 ) 0 0 0 9 3 - 7
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J.M. Choi, Y.C. Kim / Energy 27 (2002) 391–404
pump will degrade performance and deteriorate system reliability [1,2]. Therefore, the heat pump
should be charged with an optimum amount of refrigerant in order to operate with high performance over its lifetime. However, it is difficult to determine the optimum charge due to its dependency on operating parameters and expansion devices of the heat pump.
Proper selection and operation of an expansion device is the most important factor from a
standpoint of capacity and system control. Expansion devices that have been widely used in heat
pumps such as capillary tubes, short tube orifices, and thermostatic expansion valves (TXV) are
being gradually replaced with electronic expansion valves (EEVs), due to an increasing focus on
comfort, energy conservation with environmentally safe refrigerants and application of a variable
speed compressor [3,4]. Each expansion device may regulate refrigerant flow differently under
improper or proper charging conditions. Thus, it is necessary to have selection and design guides
for the EEV to achieve proper system control and high performance.
Nowadays, most heat pump systems are designed to have a small receiver or even no receiver,
to make compact systems and reduce refrigerant charge amounts. Generally, a system with a small
receiver can cause flash gas in a liquid line, and often produces instability in system operation and
flow control [5]. Therefore, adequate matching between refrigerant charge and an expansion
device is essential.
Houcek and Thedford [1] conducted tests at three charging conditions: ⫺23%, nominal, and
+23% of nominal charge. They showed that the system capacity and energy efficiency ratio (EER)
were reduced at outrange of nominal charge conditions. Stoecker et al. [2] compared the performance of an air conditioner with a capillary tube and TXV when the system was charged based
on the manufacturer’s guidelines. The test results showed that the seasonal coefficient of performance (COP) with the TXV was higher than that with the capillary tube. Domingorena [6] demonstrated that the heating capacity and COP became lower with a reduction of refrigerant charge.
Farzard and O’Neal [7,8] also reported refrigerant charging effects on the performance of a heat
pump with a capillary tube, short tube orifice, and TXV. It was found that the TXV system
showed a small variation of the COP according to refrigerant charge but a strong dependence on
the outdoor temperature.
Most of the previous studies related to effects of refrigerant charge were focused on the heat
pump with capillary tubes, short tube orifices, and TXVs. Due to several benefits of EEV such
as a wide coverage of flow rate, precise control and reduction of superheat hunting, the EEV has
been widely applied in inverter type heat pumps as well as multi-type heat pump systems. However, studies on control characteristics of the EEV by considering the refrigerant charge are very
limited in the open literature. A comprehensive study on the effects of refrigerant charge in a
heat pump having an EEV as an expansion device is required, to enhance system performance
and to achieve proper capacity modulation.
In this study, a water-to-water heat pump is tested to investigate the effects of refrigerant charge
on performance in a steady state, cooling mode operation with expansion devices of capillary
tube and EEV. The characteristics of the heat pump with the EEV are compared with those with
the capillary tube in terms of refrigerant charge. In addition, the effects of outdoor conditions
are analyzed.
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2. Experimental setup and test procedure
The experimental setup was designed to measure the performance of the water-to-water heat
pump under variable operating conditions. A schematic of the experimental setup is shown in
Fig. 1. The test rig included heat pump and water flow loops. The nominal cooling capacity of
the tested heat pump was 3.5 kW, and the working fluid was R22. The heat pump consisted of
a scroll compressor, two double tube type heat exchangers (condenser and evaporator) and an
expansion device. The condenser and evaporator were double tube type heat exchangers. These
heat exchangers had a counter flow pattern between the refrigerant and water.
Either the EEV or capillary tube could be used as the expansion device in the heat pump system.
A stepping motor using 1–2 excitation method drove the EEV. Table 1 shows the specification of
the capillary tube and EEV. The control system for EEV driving included an A/D card, stepping
motor driver, and computer.
Water was selected as a heat source and sink for the heat pump system because of its simplicity
of capacity measurements. Water flow loops (secondary flow loops) for the evaporator and condenser were closed loop having a magnetic pump and constant temperature bath. Each water bath
was equipped with a refrigeration system and electric heater. An inverter driven pump and manual
needle valve controlled the water flow rate supplied to the condenser and evaporator to establish
test conditions based on [9–11].
Temperatures in the test setup were monitored at the selected locations using thermocouples
according to ASHRAE Standard 41.1 [12], and refrigerant pressures were also measured according
Fig. 1. Schematic of the experimental setup for a water-to-water heat pump.
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Table 1
Specification of expansion devices
Expansion devices
Specification
Capillary tube
Diameter(mm)
Length(mm)
Manufacturer
Model No.
Step
Orifice diameter(mm)
Orifice length(mm)
EEV
1.2
600, 750, 900, 1200
Saginomiya
DKV-14D13
0–480
1.4
3
to ASHRAE Standard 41.3 [13]. A mass flow meter to measure refrigerant flow rate was installed
between the condenser and expansion device. The pressure drop across the mass flow meter was
approximately 3.92 kPa, which was less than the value of 82.7 kPa generally allowed in ASHRAE
Standard 116 [14] at full charge condition. A volumetric flow meter was installed to measure
water flow rate in the secondary flow loop. Each sensor was calibrated to reduce experimental
uncertainties. The specification and accuracy of sensors are summarized in Table 2.
The first step of the test procedure was to determine full charge under the standard condition.
Initially, the refrigerant was charged at water temperatures of 34°C and 25°C entering the condenser and evaporator, respectively. The EEV regulates flow rate by varying flow area in an
orifice using a needle stem. For the heat pump with the capillary tube (called the “the capillary
tube system”), the refrigerant was added into the heat pump in 50 g increments until the maximum
COP was obtained. Based on the tests, it was found that the full charge for the capillary tube
system was 1350 g. Once the full charge was determined, the heat pump was evacuated. The
refrigerant charge was then varied from ⫺20% to +20% of full charge for each expansion device.
In this study, the tests for the heat pump with the EEV (called the “the EEV system”) were
performed by setting the same full charge as the capillary tube system in order to compare characteristics of each expansion device at the same basis of charge amount. The EEV opening was
controlled to attain a constant superheat at each operating condition.
Since the heat pump is normally charged in the cooling mode [7], the tests in this study were
Table 2
Specification of sensors
Item
Accuracy
Full scale
Model
Temperature
Pressure transducer
Mass flow meter (Coriolis
meter)
Turbine flow meter
Power meter
Electronic balance weight
Data acquisition unit
±0.1°C
±0.2% of full scale
⫺270 to 400°C
3447 kPa
Thermocouple T-type
Setra C206
±0.2% of full scale
5 kg/min
Micro Motion D012S-SS-200
±0.5% of full scale
±0.01% of full scale
±0.5 g
57 lpm
20 kW
41 kg
Invalco W30750
Yokogawa WT1030
AND HP-40K
Yokogawa DA100
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395
carried out in the cooling mode only. The water temperature entering the evaporator was kept at
25°C, and that entering the condenser was varied at 30°C, 34°C, 38°C, and 42°C. Water flow
rates through the condenser and evaporator were kept constant at 9 lpm and 7 lpm, respectively.
Temperature, pressure, mass flow rate and power input of the heat pump were monitored using
a computer data acquisition system. The test data were recorded continuously for 40 min with 2s intervals. Each test was repeated three times to confirm repeatability and accuracy of the data.
Cooling capacity was calculated using water flow rate and temperature difference between
evaporator inlet and outlet. To confirm the calculation, the capacity was also determined by
refrigerant flow rate and enthalpy difference between evaporator inlet and outlet [9–11]. The
maximum difference between the water side and the refrigerant side capacity was less than 5%
which was consistent with ARI Standard 320 [10]. Ninety percent of the data was within 2.7%.
The uncertainties of cooling capacity and COP estimated by the single-sample analysis according
to ASHRAE Guideline 2 [15] were approximately 3.1% and 3.2%, respectively.
3. Results and discussion
For overcharged conditions, power consumption of the heat pump increased due to a rise of
refrigerant flow rate and compression ratio, or wet compression. For undercharged conditions,
refrigerating capacity was reduced and compressor reliability may be degraded due to high discharge temperatures. Therefore, it is essential to have the optimum amount of charge in the heat
pump to ensure high performance operation and achieve high system reliability. In the present
study, the refrigerant charge varied to determine its effect on the performance of the heat pump.
For each charge, experimental data were taken by varying operating conditions and expansion
devices including the capillary tube and EEV.
3.1. Effects of refrigerant charge on cooling capacity
Fig. 2 shows the variations of the capacity for the capillary tube system as a function of refrigerant charge. When the water temperature entering the condenser was 34°C, the heat pump had the
maximum capacity at full charge condition. The slope of the capacity with a charge amount was
much steeper at undercharged conditions than that at overcharged conditions. For overcharged
conditions, the capacity was reduced due to a decrease of the temperature difference between the
refrigerant and the water with increasing refrigerant charge. For undercharged conditions, the
capacity dropped with decreasing refrigerant charge due to a reduction of refrigerant flow rate
and compressor efficiency resulting from an increase of suction temperature. In addition, since
the superheat at undercharged conditions was higher than that at overcharged conditions, the more
significant reduction of evaporator efficiency occurred at undercharged conditions. These trends
were also reported in previous studies by Houcek and Thedford [1] and Farzard and O’Neal [7].
When the water temperature entering the condenser was 34°C, the capacity decreased by 23.7%
as the charge amount was reduced from full charge to ⫺20% of full charge. However, the
reduction of the capacity was only by 4.5% as the charge amount increased from full charge to
+20% of full charge.
The capacity was also affected by water temperature entering the condenser. For overcharged
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Fig. 2. Capacity variation of the capillary tube system as a function of refrigerant charge.
conditions, the capacity reduction with an increase of charge amount was slightly larger as the
water temperature entering the condenser was raised. However, for undercharged conditions, it
was reduced with an increase of water temperature entering the condenser. For +20% of full
charge, the capacity dropped by 4.5% and 7.23% for 34°C and 42°C, respectively, while for ⫺
20% of full charge, it decreased by 23.7% and 19.1% for 34°C and 42°C, respectively.
Generally, as the water temperature entering the condenser increased, the increment of condensing pressure was higher than that of evaporating pressure as reported by Domanski and Didion
[16]. Besides, the refrigerant flow rate through the capillary tube was strongly dependent on
condensing pressure, while it was insensitive to evaporating pressure due to choking [17]. As the
water temperature entering the condenser increased, the mass flow rate passing through the capillary tube linearly increased because the pressure difference between capillary tube inlet and outlet
became greater (Figs. 3 and 4). These results were similar to the flow trends reported by Kuehl
and Goldschmidt [17] and the ASHRAE Handbook [18]. The increment of the mass flow rate
according to the water temperature entering the condenser at undercharged conditions was higher
than that at overcharged conditions. For ⫺10% of full charge, the mass flow rate increased by
13.7% as the water temperature was altered from 30°C to 42°C. However, for +5% and +10%
of full charge, it increased by 6.7% and 5.4%, respectively, the same increase of the water temperature entering the condenser. In addition, the refrigerant flow rate increased with an addition
of the charge amount at fixed water temperatures entering the condenser and evaporator.
Fig. 5 shows the variations of the capacity for the EEV system as a function of refrigerant
charge. The variation of the capacity for the EEV system with respect to refrigerant charge was
less pronounced than that for the capillary tube system. For the water temperatures of 30°C and
34°C with the EEV system, the variations of the capacity were almost negligible as refrigerant
charge was altered from ⫺10% to +20% of full charge. The maximum cooling capacity was
dependent on water temperature entering the condenser. For the water temperatures of 38°C and
42°C, the maximum cooling capacity was observed at +5% of full charge.
As the water temperature entering the condenser increases, the superheat at the compressor
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Fig. 3. Pressure difference between inlet and outlet of the capillary tube with refrigerant charge.
Fig. 4.
Mass flow rate of the capillary tube system with a variation of condensing pressure.
inlet decreases due to a rise of refrigerant flow rate, and condensing pressure increases at a given
charge amount. Therefore, it is necessary to decrease the EEV opening to keep the superheat at
an optimum level and provide high evaporator performance. In addition, the EEV opening needs
to be reduced with the addition of refrigerant charge, because the superheat decreases due to a
rise of mass flow rate.
As the water temperature entering the condenser increased, the pressure difference between
inlet and outlet of the expansion device increased (Fig. 6). In this case, the EEV opening was
reduced to maintain the constant superheat. The EEV showed a greater slope of pressure difference
vs refrigerant charge as compared with the capillary tube as shown in Fig. 3. For overcharged
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Fig. 5. Capacity variation of the EEV system as a function of refrigerant charge.
Fig. 6. Pressure difference between inlet and outlet of the EEV with refrigerant charge.
conditions, the pressure difference across the EEV was higher than that of the capillary tube due
to increased restriction on the EEV, while for undercharged conditions, it was lower than that of
the capillary tube. Therefore, the optimum charge amount of the EEV system needs to be increased
with a rise of water temperature entering the condenser.
Fig. 7 shows refrigerant flow rate as a function of condensing pressure for different charge
amounts. For all tests with the EEV system, the maximum difference of refrigerant flow rate was
2.3 kg/h when the water temperature entering the condenser varied from 30°C to 42°C. Even
though both the pressure difference between inlet and outlet of the EEV and condensing pressure
increased with an addition of charge, the refrigerant flow rate was maintained nearly constant
with a given water temperature entering the condenser. It is due to the reduction of the EEV
opening to maintain a constant level of superheat.
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Fig. 7.
399
Mass flow rate of the EEV system with a variation of condensing pressure.
3.2. Effects of refrigerant charge on COP
Fig. 8 shows the COP of the capillary tube system as a function of water temperature entering
the condenser and refrigerant charge. As the water temperature entering the condenser increased,
the COP was significantly dropped at all charge amounts due to a higher power consumption and
lower cooling capacity. For the capillary tube system, as the charge amount deviated from the
full charge the reduction of COP was more significant at undercharged conditions than that at
Fig. 8.
COP of the capillary tube system as a function of refrigerant charge.
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overcharged conditions. This trend was also reported by Farzard and O’Neal [7]. The COP was
reduced by 16.1% and 4.8% at ⫺20% and +20% of full charge, respectively, at a water temperature of 34°C.
For the EEV system, the COP was strongly dependent on water temperature entering the condenser, but it was relatively insensitive to refrigerant charge as shown in Fig. 9. When the water
temperature entering the condenser was 34°C, the maximum difference of COP was 0.11 at all
refrigerant charges. The effects of refrigerant charge on the COP of the EEV system were considerably different from that of the capillary tube system. The EEV system showed less mass
flow rate at overcharged conditions than that of the capillary tube system. In addition, it had a
lower pressure difference between the condenser and evaporator at undercharged conditions.
Therefore, power consumption of the EEV system was less than that of the capillary tube system
for all charge conditions. Generally, the EEV system showed less degradation of the COP than
the capillary tube system at off-design charges.
3.3. Variations of superheat and subcooling
The degree of subcooling is defined as the difference between refrigerant temperature and
refrigerant saturation temperature corresponding to the pressure at the condenser exit. The subcooling at the condenser exit, which strongly affects cooling capacity and refrigerant flow rate, can
be increased by three methods: (1) an enhancement of condenser capacity, (2) an addition of
refrigerant charge, and (3) a rise of restriction on expansion devices. Degree of superheat is
defined as the difference between refrigerant temperature at the evaporator exit and evaporating
temperature. The superheat has been used as a controlling parameter to adjust refrigerant flow
rate through a variable expansion device and to ensure that only superheated vapor enters the compressor.
Fig. 9.
COP of the EEV system as a function of refrigerant charge.
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Fig. 10. Variations of subcooling with refrigerant charge in the capillary tube system.
Fig. 10 shows the subcooling of the capillary tube system as a function of refrigerant charge
at four water temperatures entering the condenser. As the refrigerant charge increased, the condensing pressure increased due to an accumulation of refrigerant in the high-pressure side, and
the subcooling became high. These trends were also observed by Stoecker et al. [2] and Farzard
and O’Neal [7]. The subcooling was reduced when the water temperature entering the condenser
increased at all charge conditions except for ⫺20% of full charge. When the water temperature
entering the condenser increased, the mean temperature difference between the water and the
refrigerant decreased and less heat was rejected in the condenser. For +20% of full charge, the
subcooling decreased from 10.4°C to 7.7°C as the water temperature entering the condenser varied
from 30°C to 42°C.
Fig. 11 shows the subcooling of the EEV system as a function of refrigerant charge. For under-
Fig. 11.
Variations of subcooling with refrigerant charge in the EEV system.
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charged conditions, the subcooling gradually increased with a rise of refrigerant charge due to an
accumulation of refrigerant in the condenser, while for overcharged conditions it rapidly increased.
The water temperature entering the condenser was another factor affecting the subcooling at all
charge conditions. Generally, the subcooling of a heat pump with a capillary tube and short tube
tended to decrease as the water temperature entering the condenser increased [2,7]. However, the
variation of the subcooling for the EEV system was nearly negligible with an increase of water
temperature entering the condenser at a given charge amount. For the EEV system, the EEV
opening was controlled in terms of the superheat. As the water temperature entering the condenser
increased, the EEV opening was reduced to maintain a constant superheat at the evaporator exit.
Therefore, the mass flow rate through the EEV did not vary much even though the condensing
pressure increased. In addition, the mean temperature difference between the refrigerant and the
water in the condenser was nearly constant according to an increase of water temperature due to
a rise of condensing temperature, resulting from a reduction of EEV opening.
Fig. 12 represents the superheat at the evaporator exit as a function of refrigerant charge. The
superheat was reduced as the refrigerant charge increased due to a rise of the mass flow rate
through the evaporator. In addition, the superheat decreased with increasing the water temperature
entering the condenser. The superheat was kept at nearly zero for +10% and +20% of full charge.
For the EEV system, the superheat was nearly constant for all water temperatures entering the
condenser and charge conditions as shown in Fig. 13. Therefore, it can be concluded that maintaining a constant superheat can optimize the performance of the heat pump at off-design conditions.
4. Conclusions
To characterize the effects of refrigerant charge on the performance of the heat pump, the
experiments were carried out by varying the refrigerant charge from ⫺20% to +20% of full charge
Fig. 12. Variations of superheat with refrigerant charge in the capillary tube system.
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Fig. 13. Variations of superheat with refrigerant charge in the EEV system.
with the capillary tube and EEV. As a result, it was found that the capillary tube system was
relatively sensitive to refrigerant charge and outdoor load conditions. In general, for the capillary
tube system the degradation of the performance was higher at undercharged conditions than that
at overcharged conditions. The capacity and COP of the EEV system had little dependence on
the refrigerant charge, while those were strongly dependent on outdoor load conditions. The EEV
system showed much higher system performance as compared with the capillary tube system.
One of the major contributions of this study was to provide a control variable to obtain optimum
cycle with the EEV. Maintaining a constant superheat by controlling the EEV opening can optimize the performance of the heat pump at off-design conditions.
One of the primary limitations of the present study was that the data were taken for only one
heat pump system. The values obtained from this study may not represent general characteristics
expected in many other air conditioners or heat pumps currently in use due to the differences in
system configuration.
Acknowledgements
This work was supported by a grant No. 99-E-ID03-P-01 from R&D Management Center for
Energy and Resources of the Korea Energy Management Corporation Foundation.
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