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Optimal refrigerant charge and energy efficiency of an oil-free refrigeration system using R134a
Applied Thermal Engineering 164 (2020) 114473
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Optimal refrigerant charge and energy efficiency of an oil-free refrigeration system using R134a Zhaohua Lia,b, Hanying Jiangb, Xinwen Chena, Kun Lianga,b, a b
T
⁎
Department of Mechanical Engineering, Yangzhou University, Yangzhou 225012, PR China Department of Engineering and Design, University of Sussex, Falmer, Brighton BN1 9QT, UK
H I GH L IG H T S
charge and energy efficiency of an oil-free VCR system were investigated. • Refrigerant increases with the refrigerant charge while subcooling shows a reverse trend. • Superheat capacity and EER reach maximum value at the optimal refrigerant charge. • Cooling capacity decreases and EER increases with condenser temperature. • Cooling • Higher compressor stroke has a higher optimal refrigerant charge.
A R T I C LE I N FO
A B S T R A C T
Keywords: Optimal refrigerant charge Energy efficiency Oil-free refrigeration R134a Volumetric efficiency
Growing number of refrigerator and air conditioner units in recent years has rapidly increased the carbon emissions due to refrigerant leakage and energy demand. Operating a vapour compression refrigeration system with an optimal refrigerant charge can effectively increase energy efficiency and capacity, and minimise refrigerant leakage thus reduce global warming effect. This paper experimentally studies the optimal refrigerant charge and energy efficiency of a small oil-free refrigeration system for a wide range of operating conditions using R134a. The oil-free refrigeration system consists of two balanced linear compressors, an off-the-shelf condenser, an expansion valve and an evaporator with an electric heater. The condenser temperature, pressure ratio and compressor stroke were controlled allowing comparable test conditions for evaluating the system performance under various refrigerant charges. The experimental results show that at a condenser temperature of 40 °C, a pressure ratio of 3.5 and a compressor stroke of 13 mm, the oil-free refrigeration system can achieve an energy efficiency ratio (EER) of 9.0 with a refrigerant charge of 280 g. The maximum cooling capacity is 218.5 W at a compressor stroke of 13 mm and a pressure ratio of 3.5 with a refrigerant charge of 300 g. By increasing 20% condenser temperature or 25% compressor stroke, the optimal refrigerant charge of the oil-free refrigeration system is increased by 11%.
1. Introduction The Montreal Protocol emphasized the importance of phasing down the use of refrigerants which have greenhouse effect. Koronaki et al. [1] mentioned that the refrigeration and air-conditioning industry in the UK consume around 16% of all UK electricity and are responsible for up to 10% of all the UK greenhouse gas emission. Refrigeration unit contributes to greenhouse emission through refrigerant leakage and energy consumption. Refrigerant charge has significant impact on the capacity, efficiency, reliability, and stability of vapour compression refrigeration (VCR) systems. However, very limited works are found in the literature
⁎
related to refrigerant charge optimisation. Rossi et al. [2] tested the system performance of a long-period food storage vertical freezer. The results indicate that higher refrigerant charge provides higher mass flow rate resulting in higher electrical consumption. Kim and Braun [3] evaluated the impacts of refrigerant charge on air conditioner and heat pump performance. The results show that refrigerant undercharging in the range of 25% can lead to an average reduction of 20% in cooling capacity and 15% in energy efficiency. Lee and Yoo [4] developed a computer program for performance analysis of condensing capacity with different refrigerant charges and condenser sizes. Deymi-Dashtebayz et al. [5] carried out experiments to evaluate the performance of
Corresponding author at: Department of Engineering and Design, University of Sussex, Falmer, Brighton BN1 9QT, UK. E-mail address: [email protected] (K. Liang).
https://doi.org/10.1016/j.applthermaleng.2019.114473 Received 4 May 2019; Received in revised form 25 September 2019; Accepted 30 September 2019 Available online 03 October 2019 1359-4311/ © 2019 Elsevier Ltd. All rights reserved.
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Nomenclature CoP D DAQ EER f h I LVDT ṁ P PID PWM Q̇ R S T t V
VCR Ẇ
coefficient of performance piston diameter data acquisition energy efficiency ratio frequency (Hz) enthalpy (kJ/kg) current (A) linear variable differential transformer mass flow rate (g/s) pressure (bar) proportional-integral-derivative pulse-width-modulation cooling capacity (W) specific gas constant (J/kg/K) stroke (mm) temperature (°C) time (s) voltage (V)
vapour compression refrigeration power (W)
Greek symbol η
efficiency
Subscripts 1 2 c cond dis g sat suc sub sup V
evaporator inlet suction cooling condenser discharge gas saturation suction subcooling superheat volumetric
performance, including works by Palm [10] and Ghoubail et al. [11]. For an overcharged system, over 50% refrigerant is contained in the condenser and the quantity of the refrigerant in other components remains unchanged. Rossi et al. [2] and Macchi et al. [12] mentioned that over 30% refrigerant is contained in the liquid pipe. Poggi et al. [13] reviewed the refrigerant charge in refrigeration systems and the strategies of charge reduction, and concluded that an overcharge involves the condenser flooding resulting in a rise of condenser pressure thus volumetric efficiency and CoP. Experimental works mentioned above all involve oil lubricant for compressors. Youbi-Idrissi et al. [14] reported that over 20% refrigerant can be dissolved in oil lubricant at a low superheat condition. The oil not only inevitably affects the heat transfer in evaporator and condenser, but also increases the refrigerant consumption of refrigeration system. Most studies mentioned above mainly evaluated the effect of refrigerant charge on cooling capacity and energy efficiency. Moreover, due to the limitation of experimental setup, the pressure ratio, and condenser temperature were not controlled. The VCR system driven by linear compressor provides significant advantages in terms of
an air-conditioning system with different refrigerant charges and ambient air temperatures. The results indicate that the highest cooling capacity of 3.2 kW and energy efficiency ratio (EER) of 2.5 can be obtained at a refrigerant charge of 640 g. Afshar et al. [6] tested R22, R404a and R134a with different refrigerant charges. The results indicate that with the increase of refrigerant charge, the condensing pressure increases. The refrigerant property is a significant parameter for working condition of a heat pump. Belman-Flores et al. [7] compared the coefficient of performance (CoP) of a domestic refrigeration system using R1234yf with R134a. Vjacheslav et al. [8] developed a model to predict the optimal refrigerant charge of a traditional VCR system. The calculated results reveal similar trends to those of experimental data. The maximum CoP is achieved at a refrigerant charge of 400 g. Pisano et al. [9] studied the combined effects of capillary tube diameter and refrigerant charge on the system performance for a light freezer. The results show that once the optimal design has been selected, if the ambient temperature changes, the optimal refrigerant charge also changes. Several researchers focused on the refrigerant distribution to evaluate the impacts of refrigerant charge on system
V1 AC Power Amplifier
Panic Box
f1, Vamp f1, f2, Vamp, Duty Cycle, LDAQ P, T, V, I1, m, LVDT1
T1
Capacitor
I1 LVDT1
PC, LabVIEW P1, P2, P3, P4, v, I1, I2, HDAQ with PID LVDT1 &2 Controllers
Capacitor
f2, Duty Cycle Solenoid Valve
COMP B
COMP Comp A A
Comp A
I2 LVDT2
Bleed Flow
PWM Driver
T6
m2 T5
P3
m1
R134a T2 P2
T3 Condenser
Evaporator Insulated
P1
Discharge
Suction T4
Variac Heater Controller
P4
Expansion Valve
Water
Fig. 1. Schematic of the oil-free VCR system and instrumentation (COMP: compressor, P: pressure transducer, T: thermocouple, m : mass flow meter, V: voltage sensor, I: current transducer, LVDT: displacement transducer). 2
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refrigerant charge reduction, flexible operating condition and system efficiency improvement. A number of works have been conducted to study of the performance of linear compressors and the system performance with different refrigerants. Zou et al. [15] proposed an improved algorithm of frequency adjustment to ensure the reliability and stability of the linear compressor. Jiang et al. [16] developed a Finite Element Analysis model to study the static motor force and magnet spring. Schmidt et al. [17] developed a thermodynamic model of a hotgas bypass cycle which can be used to select the components and tubing sizes. Jomde et al. [18] developed a novel oil-free moving coil linear compressor. The oil-free linear compressor for refrigeration application was tested with R134a. The experimental results show that the system CoP with linear compressor is 18.6% higher than the reciprocating compressor. Li et al. [19] experimental studied the performance of a VCR system using oil-free linear compressor with R1234yf and R134a. The results show that at a condenser temperature of 40 °C, the CoP of R1234yf is 5–20% lower than R134a depending on the pressure ratio and compressor stroke. It is acknowledged that the optimal refrigerant charge may also vary with compressor stroke for linear compressors which are capable of capacity modulation. This study experimentally evaluated the effect of refrigerant charge on the energy efficiency of an oil-free VCR system with variable compressor strokes using R134a. This study aims to provide optimal refrigerant charge data for oil-free VCR systems which are mostly operated at part load. Oil lubricant that inevitably recirculates with refrigerant has negative effect on the heat transfer and pressure drop within evaporator and condenser. Besides, due to variation of the dissolvability of the refrigerant in the oil lubricant under different conditions, it is difficult to quantify the optimal refrigerant charge across the operating conditions. The elimination of oil lubricant in this work allows effective comparison among different refrigerant charges and optimisation.
generated as analogue output in the LabVIEW to drive the compressor. A PID (proportional-integral-derivative) controller was developed to control the compressor stroke with displacement signal from the LVDT (linear variable differential transformer) displacement transducer. A VXA-2000 power amplifier was used to amplify the analogue signal to drive the compressors. Two 150 µF capacitors were adopted to reduce the voltage of the compressors. A second PID controller adjusted the duty cycle of the solenoid valve to keep the piston offset at zero. A Variac heater controller was used to adjust the heat into the evaporator. A power meter measured the real power of the electrical heater to be compared with cooling capacity that can be calculated from pressureenthalpy diagram (as shown in Fig. 3). Parameters including pressures (discharge, suction, body, evaporator inlet), temperatures (discharge, condenser, evaporator inlet/outlet, suction, body), and mass flow rates (main flow and bleed flow) were collected as low-speed data acquisition (LDAQ). High-speed data acquisition (HDAQ) collected data of pressures, voltage, currents, and displacements. Table 2 lists the specifications and accuracies of the instruments for the experimental apparatus. Fig. 2 shows the complete experimental apparatus for the oil-free VCR system using R134a. The oil-free VCR system mainly consists of two balanced oil-free linear compressors, an off-the-shelf water cooled coaxial condenser, an expansion valve and an evaporator with an electric heater. The specifications of components for the oil-free VCR system are listed in Table 3. 3. Data reduction The oil-free linear compressor was operated at resonance for the range of test conditions. The resonant frequency f can be calculated as:
f = 2π
k m
(1)
where k is the total stiffness and m is the total mass of compressor piston. The total stiffness k of the linear compressor is comprised by mechanical spring stiffness k m and effective gas spring stiffnesskg [20]. For low pressure ratios, the gas spring stiffness can be linearized. Thus, the effective gas spring in the compressor chamber and total stiffness can be calculate as
2. Experimental apparatus Fig. 1 shows the schematic of the oil-free VCR system and instrumentation using R134a. The pressure difference between refrigerant cylinder and oil-free VCR system was used to charge the refrigerant to the system. The amount of refrigerant charge was measured by an electronic scale. Two hundred and five steady-state measurements were conducted for the system using R134a with various refrigerant charges. The measurements were carried out at an ambient temperature of 22 ± 1 °C. Before the measurements, the compressors were preheated by resistors to over 35 °C so that no refrigerant condenses in the compressor. During the measurements, compressed hot refrigerant vapour is released from the compressor to the condenser. The cooled refrigerant liquid flowed to the evaporator via a needle valve. The pressure ratio was manually controlled by adjusting the valve lift of the needle valve. The refrigerant vapour then flowed back to the compressors. A bleed flow loop with a solenoid valve was added to control the piston offset at zero. Operating frequency was set to be 30 Hz and manually adjusted for resonance according to the calculated resonant frequency in the LabVIEW. Considering the maximum compressor stroke of 14 mm and the limitation of the displacement controller accuracy, the range of compressor stroke was set from 9 mm to 13 mm during experiments. The experiments were carried out with pressure ratios lower than 4.0. The design pressure ratio of the linear compressor is 3.0. According to the Nyquist–Shannon sampling theorem, the sampling rate of the data acquisition system was 5 kS/s which is 130 times the highest operating frequency of compressor (38 Hz). The measurements of the oil-free VCR system using R134a for each refrigerant charge were carried out at different compressor strokes, pressure ratios, frequencies, and condenser temperatures as shown in Table 1. Two NI USB-6341 data acquisition cards were used for both system control and data logging in LabVIEW. A sinusoidal waveform signal was
kg =
(Pdis − Psuc) π D2 4S
(2)
k = kg + k m
(3)
where D is the piston diameter and S is the compressor stroke. The total power input of the oil-free VCR system can be expressed as
1 Ẇ = t
∫0
t
VIdt
(4)
where t is the period, V is the voltage, and I is the current. The subcooling and superheat can be calculated as
Tsub = Tsat − Tcond
(4)
Tsup = Tsuc − Tsat
(5)
Table 1 Test conditions of the oil-free VCR system using R134a. Charge (g) Pressure ratio Compressor stroke (mm) Condenser temperature (°C) Operating frequency (Hz) Suction temperature (°C) Compressor body temperature (°C) Ambient temperature (°C) Sampling rate (kS/s)
3
220, 250, 280, 300, 330 2.0, 2.5, 3.0, 3.5, 4.0 9, 10, 11, 12, 13 40, 45, 50 32–38 20–30 > 35 22 5
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Table 2 List of instruments for the oil-free VCR system. Instruments
Model
Quantity
Accuracy (refer to value)
Data acquisition card Current transducer Voltage attenuator Isolation amplifier LVDT LVDT signal conditioner Pressure transducer Capacitor AC power amplifier
NI USB-6341 LA LEM 25-NP Fylde 261HVA HV Fylde 4600A Lucas Schaevitz ATA-101
2 2 1 1 2 2
N/A ± 0.5% ± 0.5% ± 0.5% ± 0.025 mm N/A
DRUCK PMP1400 EPCOS B32361 Vonyx VXA-2000 (class A) K-type Hastings HFM-201
4 2 1
± 0.15% ± 5% ± 1.2 dB
8 1
± 1.5 °C ± 1%
Tylan FM-360 RS Pro IDS1000 Series IDS1072AU Electronic Wattmeter EW604 RCS-7040
1 3
± 1% N/A
1
< 2.5% (50 Hz, unity power factor, 25 °C) ± 0.05%
Thermocouple Main mass flow meter Bleed flow meter Oscilloscope Power meter (heater) Electric charge scale
1
Fig. 3. Pressure-enthalpy (p-h) diagram of the oil-free VCR system with various refrigerant charges for a pressure ratio of 3.5, a compressor stroke of 13 mm and a condenser temperature of 40 °C.
refrigerant charge due to the increase of discharge and suction pressures. The enthalpy difference with a refrigerant charge of 330 g is 4% higher than the refrigerant charge of 220 g. This will lead to high cooling capacity if the mass flow rate is the same. However, the highest cooling capacity of the oil-free VCR system for given conditions is achieved with a refrigerant charge of 280 g. This is mainly due to the variation of the mass flow rate with refrigerant charge. Fig. 4 shows the discharge and suction pressures against refrigerant charge with different condenser temperatures. For a fixed condenser temperature, both discharge and suction pressures increase with refrigerant charge. For a fixed refrigerant charge, discharge and suction pressures increase with condenser temperature. Higher suction and discharge pressures lead to a higher in-cylinder pressure resulting in a higher shaft force thus higher power input and copper loss. For a condenser temperature of 40 °C, the discharge and suction pressure is 9.8 bar and 2.8 bar, respectively, at a refrigerant charge of 220 g which is 23% lower than that of 330 g (12.7 bar and 3.6 bar). For a refrigerant charge of 220 g, both discharge and suction pressures at condenser temperature of 40 °C are 25% lower than 50 °C. Fig. 5 shows the condenser inlet and evaporator inlet temperatures against refrigerant charge with various condense temperatures at a pressure ratio of 3.5 and a compressor stroke of 13 mm. Both condenser and evaporator inlet temperatures increase with refrigerant charge. This is because higher refrigerant charge leads to higher discharge and suction pressures resulting in higher discharge temperature and evaporator inlet pressure thus higher condenser and evaporator inlet temperatures. With a refrigerant charge of 220 g, the oil-free VCR system can achieve an evaporator inlet temperature of 3.7 °C. With the increase of condenser temperature, both condenser and evaporator inlet temperatures present a growing trend. It is worth mentioning that the increase rate of evaporator inlet temperature against the increase of refrigerant charge is higher at low condenser temperatures. This is due
The volumetric efficiency was calculated according to:
ηV =
̇ g Tsuc 4mR πSfPsuc D2
(6)
where f is the operating frequency, Tsuc is the temperature at the compressor inlet, Psuc is the pressure at the compressor inlet, Rg is the specific gas constant and ṁ is the mass flow rate. Entropy and enthalpy for the refrigerant during refrigerant cycle were determined using CoolProp [21] based on the condenser outlet temperature, suction temperature, discharge pressure and suction pressure. Thus, the cooling capacity can be calculated as
Qċ = ṁ (h2 − h1)
(7)
where h1 is the enthalpy of the refrigerant at evaporator inlet and h2 is the enthalpy of the refrigerant at suction. Thus, energy efficiency ratio (EER) can be determined as
EER = 3.41
Ẇ Pin
(8)
4. Results and discussions Fig. 3 shows the pressure-enthalpy (p-h) diagram of the oil-free VCR system with various refrigerant charges for a pressure ratio of 3.5, a compressor stroke of 13 mm, a condenser temperature of 40 °C and an evaporator inlet temperature of 5 °C. The enthalpy difference between the evaporator inlet and outlet increases slightly with the increase of
Heater Power Meter Power Amplifier
Solenoid Valve
Linear Compressor
NI DAQ Card
Variac Heater Controller
Evaporator with Electric Heater
Expansion Valve
Condenser
Fig. 2. Complete experimental apparatus for the oil-free VCR system using R134a. 4
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Table 3 List of components for the oil-free VCR system. Components
Specifications
Compressor
Two identical oil-free linear compressors working in opposite, piston diameter of 19 mm, maximum compressor stroke of 14 mm, rated power of 100 W for each Coaxial water-cooled, copper, off-the-shelf, coolant connection diameter of 12.7 mm, refrigerant connection diameter of 16 mm Needle valve, stainless steel medium flow high pressure Copper, electric heater (resistance of 50 Ω), length of 128 cm , inner diameter of 7.9 mm, outer diameter of 12.7 mm R134a from BOC
Condenser Expansion valve Evaporator Refrigerant
condensation. The reduction of heat transfer efficiency decreases the quality of refrigerant circulating in the system. For a condenser temperatures of 40 °C and 45 °C the highest mass flow rates are achieved at a refrigerant charge of 280 g. For a condenser temperature of 50 °C, the highest mass flow rate is achieved at a refrigerant charge of 300 g. High operating frequency due to high suction and discharge pressures could also contribute to the increase of mass flow rate at high condenser temperature. The highest mass flow rate is 1.51 g/s at a condenser temperature of 50 °C with a refrigerant charge of 300 g while the lowest is 0.93 g/s at a condenser temperature of 40 °C with a refrigerant charge of 220 g. Fig. 9 shows the variation of volumetric efficiency with refrigerant charge at a pressure ratio of 3.5, a compressor stroke of 13 mm and condenser temperatures of 40 °C, 45 °C, and 50 °C. For a fixed condenser temperature, the variation of the volumetric efficiency has a similar trend to mass flow rate shown in Fig. 8. The oil-free VCR system can achieve a volumetric efficiency of 35.5% at a condenser temperature of 40 °C. For a fixed refrigerant charge, higher mass flow rate does not lead to higher volumetric efficiency. Mass flow rate is increased by 16% with the increase of condenser temperature from 40° to 50° while suction pressure is increased by 22%. The lower increment of the mass flow rate compared with the suction pressure lead to a decrease of volumetric efficiency with the increase of condenser temperature. Fig. 10 shows the variation of cooling capacity with refrigerant charge at a pressure ratio of 3.5, a compressor stroke of 13 mm and condenser temperatures of 40 °C, 45 °C, and 50 °C. The cooling capacity shows a similar trend with mass flow rate shown in Fig. 8. Both enthalpy difference and mass flow rate affect the cooling capacity. For the measurements shown in Fig. 3, the enthalpy difference only decreases by about 3% by increasing refrigerant charge from 220 g to 280 g while mass flow rate is increased by 25%. The cooling capacity reaches the maximum value while the system provides the highest mass flow rate. For the undercharged condition, the insufficient refrigerant in the system resulting in a small mass flow rate is the main reason for the low cooling capacity. For the overcharged system, the excessive liquid refrigerant contained in the condenser which reduces the heat transfer
to the small increment of saturation pressure at low temperatures. Fig. 6 shows the superheat and subcooling varying with refrigerant charge at a pressure ratio of 3.5, a compressor stroke of 13 mm and condenser temperature of 40 °C, 45 °C, and 50 °C. For a fixed condenser temperature, superheat decreases with the increase of refrigerant charge as the increase of suction pressure results in a rise in suction saturation temperature. For a fixed refrigerant charge, higher condenser temperature tends to have lower superheat due to the increase of discharge and suction pressures. Subcooling shows a reverse trend compared with superheat. The maximum superheat and subcooling is 26.5 K (at the refrigerant charge of 220 g) and 8.5 K (at the refrigerant charge of 330 g), respectively. Several researchers [5,22,23] pointed out that the subcooling of a system is an indicator of the quantity of the refrigerant in the condenser. Higher subcooling represents greater length of subcooled two-phase region thus higher quantity of refrigerant in the condenser. The variation of power input with refrigerant charge at a pressure ratio of 3.5, a compressor stroke of 13 mm and condenser temperatures of 40 °C, 45 °C, and 50 °C is shown in Fig. 7. At a fixed condenser temperature, the power input increases with the refrigerant charge due to the requirement of higher shaft force caused by the increasing incylinder pressure. The highest power input of the oil-free VCR system occurs at a refrigerant charge of 330 g and a condenser temperature of 50 °C. High refrigerant charge could lead to higher gas spring stiffness thus higher resonant frequency as well. Higher power input also cause higher copper loss and relatively lower motor efficiency. Fig. 8 shows the mass flow rate against refrigerant charge with various condenser temperatures at a compressor stroke of 13 mm and a pressure ratio of 3.5. For a fixed refrigerant charge, the mass flow rate decreases as the condenser temperature increases. This is because high condensation temperature causes more refrigerant vapour to circulate in the VCR system rather than condensing in the condenser. For undercharged system, the mass flow rate increases with refrigerant charge until reaching the optimal refrigerant charge. With the increase of the refrigerant charge, the increasing subcooling region occupies a larger portion of the condenser resulting in a reduction of the area for
Fig. 4. Discharge and suction pressure against refrigerant charge with various condenser temperatures at a pressure ratio of 3.5 and a compressor stroke of 13 mm. 5
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Fig. 5. Condenser inlet and evaporator inlet temperatures against refrigerant charge with various condenser temperatures at a pressure ratio of 3.5 and a compressor stroke of 13 mm.
Fig. 7. Variation of power input with refrigerant charge at a pressure ratio of 3.5, a compressor stroke of 13 mm and condenser temperatures of 40 °C, 45 °C, and 50 °C.
Fig. 6. Superheat (a) and subcooling (b) against refrigerant charge at a pressure ratio of 3.5, a compressor stroke of 13 mm and condenser temperatures of 40 °C, 45 °C, and 50 °C.
Fig. 8. Variation of mass flow rate with refrigerant charge at a pressure ratio of 3.5, a compressor stroke of 13 mm and condenser temperatures of 40 °C, 45 °C, and 50 °C.
area of the refrigerant vapour is the main reason for the cooling capacity deterioration. The highest cooling capacity of the oil-free system is 218.5 W at a condenser temperature 50 °C, a compressor stroke of 13 mm, a pressure ratio of 3.5 and a refrigerant charge of 300 g. Fig. 11 shows the variation of EER with refrigerant charge at a pressure ratio of 3.5, a compressor stroke of 13 mm and condenser temperatures of 40 °C, 45 °C, and 50 °C. For a condenser temperature of 40 °C, the EER increases with refrigerant charge before reaching the
optimal charge (280 g). With further increased refrigerant charge, the condenser flooding and high in-cylinder pressure result in a decrease of mass flow rate and high power input thus a decrease of EER. For a fixed refrigerant charge, the EER decreases with the increase of condenser temperature. Though higher condenser temperature tends to have higher cooling capacity, the excessive power input results in a decrease of EER. The maximum EER is 9.0 with a refrigerant charge of 280 g and 6
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Fig. 9. Volumetric efficiency against refrigerant charge for the oil-free refrigeration system at a pressure ratio of 3.5, a compressor stroke of 13 mm and condenser temperatures of 40 °C, 45 °C, and 50 °C.
Fig. 12. EER against refrigerant charge for the oil-free refrigeration system with various compressor strokes at a pressure ratio of 2.5and condenser temperatures of 40 °C, 45 °C, and 50 °C.
Fig. 10. Variation of cooling capacity with refrigerant charge at a pressure ratio of 3.5, a compressor stroke of 13 mm and condenser temperatures of 40 °C, 45 °C, and 50 °C.
Fig. 13. Variation of the optimal refrigerant charge for the oil-free refrigeration system with various pressure ratio and compressor stroke at a fixed condenser temperature of 40 °C.
Table 4 The optimal charge of the oil-free VCR system for various compressor stroke and condenser temperature at a pressure ratio of 2.5. Tcond (°C)
40 45 50
Stroke (mm) 9
10
11
12
13
250 None None
250 280 None
250 280 280
280 280 280
None None 300
of 2.5 and condenser temperatures of 40 °C, 45 °C, and 50 °C. For a fixed refrigerant charge, the EER increases with compressor stroke due to the improvement of the volumetric efficiency. For a fixed compressor stroke, the EER increases firstly with the refrigerant charge until reaching the optimal charge. With an increase of the compressor stroke, the optimal charge of the oil-free VCR system increases. Fig. 13 shows the optimal charge of the oil-free VCR system for various operating conditions at a condenser temperature of 40 °C. For a fixed compressor stroke, the optimal charge decreases with the pressure ratio. For a fixed pressure ratio, the optimal charge increases with the compressor stroke. This is because a higher compressor stroke and lower pressure ratio tend to have a higher volumetric efficiency resulting in an increase of mass flow rate. Table 4 lists the optimal charge of the oil-free VCR system for various compressor strokes and condenser temperatures at a pressure ratio of 2.5. At a condenser temperature of
Fig. 11. EER against refrigerant charge for the oil-free refrigeration system at a pressure ratio of 3.5, a compressor stroke of 13 mm, and condenser temperatures of 40 °C, 45 °C, and 50 °C.
a condenser temperature of 40 °C. By reducing the refrigerant charge by 11.3%, the EER of the system is reduced by 7.5%. Thus, to operate the system under a slightly undercharged condition may be considered to reduce the refrigerant charge thus leakage. Fig. 12 shows the EER against refrigerant charge for the oil-free refrigeration system with various compressor strokes at a pressure ratio 7
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40 °C, by increasing compressor stroke by 25%, the optimal refrigerant charge is increased by 11%. The same amount of the increment of optimal refrigerant charge can be achieved by increasing 20% (40–50 °C) of condenser temperature. The optimal refrigerant charge is more sensitive on condenser temperature than compressor stroke. A higher condenser capacity tends to have a higher optimal refrigerant charge. Ghoubali et al. [11] reported that using microchannel heat exchanger as condenser can reduce 13% of the total refrigerant charge concentrating in the condenser. Moreover, more than 18% refrigerant can be contained in liquid line [2]. To minimize the diameter of the liquid line also can be a potential choice for the refrigerant charge reduction.
(UA016-05) for supporting the instrumentation of the VCR system is acknowledged. The authors are also grateful to Prof Richard Stone, Dr Mike Dadd and Paul Bailey at the University of Oxford for the prototype oil-free linear compressor. The authors are grateful to Andy White and Alex Burns at the University of Sussex for the lab support.
5. Experimental uncertainty
[1] I. Koronaki, et al., Refrigerant emissions and leakage prevention across EuropeResults from the RealSkillsEurope project, Energy 45 (1) (2012) 71–80. [2] F. De Rossi, A. Mauro, M. Musto, G.P. Vanoli, Long-period food storage household vertical freezer: refrigerant charge influence on working conditions during steady operation, Int. J. Refrig. 34 (5) (2011) 1305–1314. [3] H.C. Kim, G.A. Keoleian, Y.A. Horie, Optimal household refrigerator replacement policy for life cycle energy, greenhouse gas emissions, and cost, Energy Policy 34 (15) (2006) 2310–2323. [4] G. Lee, J. Yoo, Performance analysis and simulation of automobile air conditioning system, Int. J. Refrig. 23 (3) (2000) 243–254. [5] M. Deymi-Dashtebayaz, M. Farahnak, M. Moraffa, A. Ghalami, N. Mohammadi, Experimental evaluation of refrigerant mass charge and ambient air temperature effects on performance of air-conditioning systems, Heat Mass Transf. 54 (3) (2018) 803–812. [6] F. Afshari, O. Comakli, N. Adiguzel, S. Karagoz, Optimal charge amount for different refrigerants in air-to-water heat pumps, Iran. J. Sci. Technol. Trans. Mech. Eng. 40 (4) (2016) 325–335. [7] J. Belman-Flores, V. Rangel-Hernandez, S. Uson, C. Rubio-Maya, Energy and exergy analysis of R1234yf as drop-in replacement for R134a in a domestic refrigeration system, Energy 132 (2017) 116–125. [8] N. Vjacheslav, A. Rozhentsev, C.-C. Wang, Rationally based model for evaluating the optimal refrigerant mass charge in refrigerating machines, Energy Convers. Manage. 42 (18) (2001) 2083–2095. [9] A. Pisano, S. Martínez-Ballester, J.M. Corberán, A.W. Mauro, Optimal design of a light commercial freezer through the analysis of the combined effects of capillary tube diameter and refrigerant charge on the performance, Int. J. Refrig. 52 (2015) 1–10. [10] B. Palm, Refrigeration systems with minimum charge of refrigerant, Appl. Therm. Eng. 27 (10) (2007) 1693–1701. [11] R. Ghoubali, P. Byrne, F. Bazantay, Refrigerant charge optimisation for propane heat pump water heaters, Int. J. Refrig. 76 (2017) 230–244. [12] H. Macchi, J. Guilpart, A. Mahungu, Reduction de charge: comparaison entre detente directe, recirculation et réfrigération indirecte, Journée Française du FroidInterclima (1999) 47–63. [13] F. Poggi, H. Macchi-Tejeda, D. Leducq, A. Bontemps, Refrigerant charge in refrigerating systems and strategies of charge reduction, Int. J. Refrig. 31 (3) (2008) 353–370. [14] M. Youbi-Idrissi, J. Bonjour, C. Marvillet, F. Meunier, Impact of refrigerant–oil solubility on an evaporator performances working with R-407C, Int. J. Refrig. 26 (3) (2003) 284–292. [15] H. Zou, M. Wang, M. Tang, C. Tian, X. Chen, Study on stroke unstable fluctuation of linear compressor and start-up control optimization with frequency tracking, Int. J. Refrig. 96 (2018) 100–105. [16] H. Jiang, K. Liang, Z. Li, Characteristics of a novel moving magnet linear motor for linear compressor, Mech. Syst. Sig. Process. 121 (2019) 828–840. [17] D. Schmidt, J. Singleton, C.R. Bradshaw, Development of a light-commercial compressor load stand to measure compressor performance using low-GWP refrigerants, Int. J. Refrig. 100 (2019) 443–453. [18] A. Jomde, et al., Modeling and measurement of a moving coil oil-free linear compressor performance for refrigeration application using R134a, Int. J. Refrig. 88 (2018) 182–194. [19] Z. Li, K. Liang, H. Jiang, Experimental study of R1234yf as a drop-in replacement for R134a in an oil-free refrigeration system, Appl. Therm. Eng. 153 (2019) 646–654. [20] K. Liang, R. Stone, M. Dadd, P. Bailey, Piston position sensing and control in a linear compressor using a search coil, Int. J. Refrig. 66 (2016) 32–40. [21] Property calculator. Available: http://platform.sysmoltd.com/ThermoFluids/ FluidPropsCalculatorView (2019, 19 March). [22] E. Björk, B. Palm, Refrigerant mass charge distribution in a domestic refrigerator. Part II: steady state conditions, Appl. Therm. Eng. 26 (8–9) (2006) 866–871. [23] G. Pottker, P. Hrnjak, Effect of the condenser subcooling on the performance of vapor compression systems, Int. J. Refrig. 50 (2015) 156–164.
Appendix A. Supplementary material Supplementary data to this article can be found online at https:// doi.org/10.1016/j.applthermaleng.2019.114473. References
Pressures, temperature, stroke, current, voltage, and mass flow rate were measured during the experiments. Typically, a set of readings was taken every 15 min to allow time for thermal equilibrium to be attained. The measurements of pressure, stroke, and temperature have absolute uncertainties of 0.015 bar, 0.025 mm, and 0.5 °C. The current transducer, voltage transducers, and mass flow rate have accuracies of 0.5%, 0.5%, and 1% respectively. The combined uncertainties of the calculated values were calculated using a 95% confidence interval. The cooling capacity, EER, and volumetric efficiency have relative uncertainty of 1%, 1.3%, and 0.3%, respectively. 6. Conclusions Operating a VCR system with an optimal refrigerant charge can reduce the refrigerant leakage without losing capacity and energy efficiency. In this study, the optimal refrigerant charge and energy efficiency of an oil-free refrigeration system using R134a were investigated for a wide range of operating conditions. The key findings are shown below: (1) For given pressure ratio and stroke, and compressor stroke, the discharge and suction pressure increase with the increase of the refrigerant charge. This leads to an increasing power input and copper loss. (2) At a fixed condenser temperature, compressor stroke, and pressure ratio, the superheat increases with the refrigerant charge while subcooling shows a reverse trend due to the increase of the suction and discharge pressures. (3) At a fixed condenser temperature, the mass flow rate increases first with the refrigerant charge, then decreases due to the flooding in the condenser and the increase of the in-cylinder pressure. (4) For a fixed compressor stroke and a fixed pressure ratio, both cooling capacity and EER reach maximum value at the optimal refrigerant charge. The cooling capacity decreases with condenser temperature while EER increases with condenser temperature. (5) The optimal charge varies with operating conditions. A higher compressor stroke and a lower pressure ratio tend to have a higher optimal refrigerant charge. Declaration of Competing Interest The authors declared that there is no conflict of interest. Acknowledgement The Research Development Fund from the University of Sussex
8
Contents lists available at ScienceDirect
Applied Thermal Engineering journal homepage: www.elsevier.com/locate/apthermeng
Optimal refrigerant charge and energy efficiency of an oil-free refrigeration system using R134a Zhaohua Lia,b, Hanying Jiangb, Xinwen Chena, Kun Lianga,b, a b
T
⁎
Department of Mechanical Engineering, Yangzhou University, Yangzhou 225012, PR China Department of Engineering and Design, University of Sussex, Falmer, Brighton BN1 9QT, UK
H I GH L IG H T S
charge and energy efficiency of an oil-free VCR system were investigated. • Refrigerant increases with the refrigerant charge while subcooling shows a reverse trend. • Superheat capacity and EER reach maximum value at the optimal refrigerant charge. • Cooling capacity decreases and EER increases with condenser temperature. • Cooling • Higher compressor stroke has a higher optimal refrigerant charge.
A R T I C LE I N FO
A B S T R A C T
Keywords: Optimal refrigerant charge Energy efficiency Oil-free refrigeration R134a Volumetric efficiency
Growing number of refrigerator and air conditioner units in recent years has rapidly increased the carbon emissions due to refrigerant leakage and energy demand. Operating a vapour compression refrigeration system with an optimal refrigerant charge can effectively increase energy efficiency and capacity, and minimise refrigerant leakage thus reduce global warming effect. This paper experimentally studies the optimal refrigerant charge and energy efficiency of a small oil-free refrigeration system for a wide range of operating conditions using R134a. The oil-free refrigeration system consists of two balanced linear compressors, an off-the-shelf condenser, an expansion valve and an evaporator with an electric heater. The condenser temperature, pressure ratio and compressor stroke were controlled allowing comparable test conditions for evaluating the system performance under various refrigerant charges. The experimental results show that at a condenser temperature of 40 °C, a pressure ratio of 3.5 and a compressor stroke of 13 mm, the oil-free refrigeration system can achieve an energy efficiency ratio (EER) of 9.0 with a refrigerant charge of 280 g. The maximum cooling capacity is 218.5 W at a compressor stroke of 13 mm and a pressure ratio of 3.5 with a refrigerant charge of 300 g. By increasing 20% condenser temperature or 25% compressor stroke, the optimal refrigerant charge of the oil-free refrigeration system is increased by 11%.
1. Introduction The Montreal Protocol emphasized the importance of phasing down the use of refrigerants which have greenhouse effect. Koronaki et al. [1] mentioned that the refrigeration and air-conditioning industry in the UK consume around 16% of all UK electricity and are responsible for up to 10% of all the UK greenhouse gas emission. Refrigeration unit contributes to greenhouse emission through refrigerant leakage and energy consumption. Refrigerant charge has significant impact on the capacity, efficiency, reliability, and stability of vapour compression refrigeration (VCR) systems. However, very limited works are found in the literature
⁎
related to refrigerant charge optimisation. Rossi et al. [2] tested the system performance of a long-period food storage vertical freezer. The results indicate that higher refrigerant charge provides higher mass flow rate resulting in higher electrical consumption. Kim and Braun [3] evaluated the impacts of refrigerant charge on air conditioner and heat pump performance. The results show that refrigerant undercharging in the range of 25% can lead to an average reduction of 20% in cooling capacity and 15% in energy efficiency. Lee and Yoo [4] developed a computer program for performance analysis of condensing capacity with different refrigerant charges and condenser sizes. Deymi-Dashtebayz et al. [5] carried out experiments to evaluate the performance of
Corresponding author at: Department of Engineering and Design, University of Sussex, Falmer, Brighton BN1 9QT, UK. E-mail address: [email protected] (K. Liang).
https://doi.org/10.1016/j.applthermaleng.2019.114473 Received 4 May 2019; Received in revised form 25 September 2019; Accepted 30 September 2019 Available online 03 October 2019 1359-4311/ © 2019 Elsevier Ltd. All rights reserved.
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Nomenclature CoP D DAQ EER f h I LVDT ṁ P PID PWM Q̇ R S T t V
VCR Ẇ
coefficient of performance piston diameter data acquisition energy efficiency ratio frequency (Hz) enthalpy (kJ/kg) current (A) linear variable differential transformer mass flow rate (g/s) pressure (bar) proportional-integral-derivative pulse-width-modulation cooling capacity (W) specific gas constant (J/kg/K) stroke (mm) temperature (°C) time (s) voltage (V)
vapour compression refrigeration power (W)
Greek symbol η
efficiency
Subscripts 1 2 c cond dis g sat suc sub sup V
evaporator inlet suction cooling condenser discharge gas saturation suction subcooling superheat volumetric
performance, including works by Palm [10] and Ghoubail et al. [11]. For an overcharged system, over 50% refrigerant is contained in the condenser and the quantity of the refrigerant in other components remains unchanged. Rossi et al. [2] and Macchi et al. [12] mentioned that over 30% refrigerant is contained in the liquid pipe. Poggi et al. [13] reviewed the refrigerant charge in refrigeration systems and the strategies of charge reduction, and concluded that an overcharge involves the condenser flooding resulting in a rise of condenser pressure thus volumetric efficiency and CoP. Experimental works mentioned above all involve oil lubricant for compressors. Youbi-Idrissi et al. [14] reported that over 20% refrigerant can be dissolved in oil lubricant at a low superheat condition. The oil not only inevitably affects the heat transfer in evaporator and condenser, but also increases the refrigerant consumption of refrigeration system. Most studies mentioned above mainly evaluated the effect of refrigerant charge on cooling capacity and energy efficiency. Moreover, due to the limitation of experimental setup, the pressure ratio, and condenser temperature were not controlled. The VCR system driven by linear compressor provides significant advantages in terms of
an air-conditioning system with different refrigerant charges and ambient air temperatures. The results indicate that the highest cooling capacity of 3.2 kW and energy efficiency ratio (EER) of 2.5 can be obtained at a refrigerant charge of 640 g. Afshar et al. [6] tested R22, R404a and R134a with different refrigerant charges. The results indicate that with the increase of refrigerant charge, the condensing pressure increases. The refrigerant property is a significant parameter for working condition of a heat pump. Belman-Flores et al. [7] compared the coefficient of performance (CoP) of a domestic refrigeration system using R1234yf with R134a. Vjacheslav et al. [8] developed a model to predict the optimal refrigerant charge of a traditional VCR system. The calculated results reveal similar trends to those of experimental data. The maximum CoP is achieved at a refrigerant charge of 400 g. Pisano et al. [9] studied the combined effects of capillary tube diameter and refrigerant charge on the system performance for a light freezer. The results show that once the optimal design has been selected, if the ambient temperature changes, the optimal refrigerant charge also changes. Several researchers focused on the refrigerant distribution to evaluate the impacts of refrigerant charge on system
V1 AC Power Amplifier
Panic Box
f1, Vamp f1, f2, Vamp, Duty Cycle, LDAQ P, T, V, I1, m, LVDT1
T1
Capacitor
I1 LVDT1
PC, LabVIEW P1, P2, P3, P4, v, I1, I2, HDAQ with PID LVDT1 &2 Controllers
Capacitor
f2, Duty Cycle Solenoid Valve
COMP B
COMP Comp A A
Comp A
I2 LVDT2
Bleed Flow
PWM Driver
T6
m2 T5
P3
m1
R134a T2 P2
T3 Condenser
Evaporator Insulated
P1
Discharge
Suction T4
Variac Heater Controller
P4
Expansion Valve
Water
Fig. 1. Schematic of the oil-free VCR system and instrumentation (COMP: compressor, P: pressure transducer, T: thermocouple, m : mass flow meter, V: voltage sensor, I: current transducer, LVDT: displacement transducer). 2
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refrigerant charge reduction, flexible operating condition and system efficiency improvement. A number of works have been conducted to study of the performance of linear compressors and the system performance with different refrigerants. Zou et al. [15] proposed an improved algorithm of frequency adjustment to ensure the reliability and stability of the linear compressor. Jiang et al. [16] developed a Finite Element Analysis model to study the static motor force and magnet spring. Schmidt et al. [17] developed a thermodynamic model of a hotgas bypass cycle which can be used to select the components and tubing sizes. Jomde et al. [18] developed a novel oil-free moving coil linear compressor. The oil-free linear compressor for refrigeration application was tested with R134a. The experimental results show that the system CoP with linear compressor is 18.6% higher than the reciprocating compressor. Li et al. [19] experimental studied the performance of a VCR system using oil-free linear compressor with R1234yf and R134a. The results show that at a condenser temperature of 40 °C, the CoP of R1234yf is 5–20% lower than R134a depending on the pressure ratio and compressor stroke. It is acknowledged that the optimal refrigerant charge may also vary with compressor stroke for linear compressors which are capable of capacity modulation. This study experimentally evaluated the effect of refrigerant charge on the energy efficiency of an oil-free VCR system with variable compressor strokes using R134a. This study aims to provide optimal refrigerant charge data for oil-free VCR systems which are mostly operated at part load. Oil lubricant that inevitably recirculates with refrigerant has negative effect on the heat transfer and pressure drop within evaporator and condenser. Besides, due to variation of the dissolvability of the refrigerant in the oil lubricant under different conditions, it is difficult to quantify the optimal refrigerant charge across the operating conditions. The elimination of oil lubricant in this work allows effective comparison among different refrigerant charges and optimisation.
generated as analogue output in the LabVIEW to drive the compressor. A PID (proportional-integral-derivative) controller was developed to control the compressor stroke with displacement signal from the LVDT (linear variable differential transformer) displacement transducer. A VXA-2000 power amplifier was used to amplify the analogue signal to drive the compressors. Two 150 µF capacitors were adopted to reduce the voltage of the compressors. A second PID controller adjusted the duty cycle of the solenoid valve to keep the piston offset at zero. A Variac heater controller was used to adjust the heat into the evaporator. A power meter measured the real power of the electrical heater to be compared with cooling capacity that can be calculated from pressureenthalpy diagram (as shown in Fig. 3). Parameters including pressures (discharge, suction, body, evaporator inlet), temperatures (discharge, condenser, evaporator inlet/outlet, suction, body), and mass flow rates (main flow and bleed flow) were collected as low-speed data acquisition (LDAQ). High-speed data acquisition (HDAQ) collected data of pressures, voltage, currents, and displacements. Table 2 lists the specifications and accuracies of the instruments for the experimental apparatus. Fig. 2 shows the complete experimental apparatus for the oil-free VCR system using R134a. The oil-free VCR system mainly consists of two balanced oil-free linear compressors, an off-the-shelf water cooled coaxial condenser, an expansion valve and an evaporator with an electric heater. The specifications of components for the oil-free VCR system are listed in Table 3. 3. Data reduction The oil-free linear compressor was operated at resonance for the range of test conditions. The resonant frequency f can be calculated as:
f = 2π
k m
(1)
where k is the total stiffness and m is the total mass of compressor piston. The total stiffness k of the linear compressor is comprised by mechanical spring stiffness k m and effective gas spring stiffnesskg [20]. For low pressure ratios, the gas spring stiffness can be linearized. Thus, the effective gas spring in the compressor chamber and total stiffness can be calculate as
2. Experimental apparatus Fig. 1 shows the schematic of the oil-free VCR system and instrumentation using R134a. The pressure difference between refrigerant cylinder and oil-free VCR system was used to charge the refrigerant to the system. The amount of refrigerant charge was measured by an electronic scale. Two hundred and five steady-state measurements were conducted for the system using R134a with various refrigerant charges. The measurements were carried out at an ambient temperature of 22 ± 1 °C. Before the measurements, the compressors were preheated by resistors to over 35 °C so that no refrigerant condenses in the compressor. During the measurements, compressed hot refrigerant vapour is released from the compressor to the condenser. The cooled refrigerant liquid flowed to the evaporator via a needle valve. The pressure ratio was manually controlled by adjusting the valve lift of the needle valve. The refrigerant vapour then flowed back to the compressors. A bleed flow loop with a solenoid valve was added to control the piston offset at zero. Operating frequency was set to be 30 Hz and manually adjusted for resonance according to the calculated resonant frequency in the LabVIEW. Considering the maximum compressor stroke of 14 mm and the limitation of the displacement controller accuracy, the range of compressor stroke was set from 9 mm to 13 mm during experiments. The experiments were carried out with pressure ratios lower than 4.0. The design pressure ratio of the linear compressor is 3.0. According to the Nyquist–Shannon sampling theorem, the sampling rate of the data acquisition system was 5 kS/s which is 130 times the highest operating frequency of compressor (38 Hz). The measurements of the oil-free VCR system using R134a for each refrigerant charge were carried out at different compressor strokes, pressure ratios, frequencies, and condenser temperatures as shown in Table 1. Two NI USB-6341 data acquisition cards were used for both system control and data logging in LabVIEW. A sinusoidal waveform signal was
kg =
(Pdis − Psuc) π D2 4S
(2)
k = kg + k m
(3)
where D is the piston diameter and S is the compressor stroke. The total power input of the oil-free VCR system can be expressed as
1 Ẇ = t
∫0
t
VIdt
(4)
where t is the period, V is the voltage, and I is the current. The subcooling and superheat can be calculated as
Tsub = Tsat − Tcond
(4)
Tsup = Tsuc − Tsat
(5)
Table 1 Test conditions of the oil-free VCR system using R134a. Charge (g) Pressure ratio Compressor stroke (mm) Condenser temperature (°C) Operating frequency (Hz) Suction temperature (°C) Compressor body temperature (°C) Ambient temperature (°C) Sampling rate (kS/s)
3
220, 250, 280, 300, 330 2.0, 2.5, 3.0, 3.5, 4.0 9, 10, 11, 12, 13 40, 45, 50 32–38 20–30 > 35 22 5
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Table 2 List of instruments for the oil-free VCR system. Instruments
Model
Quantity
Accuracy (refer to value)
Data acquisition card Current transducer Voltage attenuator Isolation amplifier LVDT LVDT signal conditioner Pressure transducer Capacitor AC power amplifier
NI USB-6341 LA LEM 25-NP Fylde 261HVA HV Fylde 4600A Lucas Schaevitz ATA-101
2 2 1 1 2 2
N/A ± 0.5% ± 0.5% ± 0.5% ± 0.025 mm N/A
DRUCK PMP1400 EPCOS B32361 Vonyx VXA-2000 (class A) K-type Hastings HFM-201
4 2 1
± 0.15% ± 5% ± 1.2 dB
8 1
± 1.5 °C ± 1%
Tylan FM-360 RS Pro IDS1000 Series IDS1072AU Electronic Wattmeter EW604 RCS-7040
1 3
± 1% N/A
1
< 2.5% (50 Hz, unity power factor, 25 °C) ± 0.05%
Thermocouple Main mass flow meter Bleed flow meter Oscilloscope Power meter (heater) Electric charge scale
1
Fig. 3. Pressure-enthalpy (p-h) diagram of the oil-free VCR system with various refrigerant charges for a pressure ratio of 3.5, a compressor stroke of 13 mm and a condenser temperature of 40 °C.
refrigerant charge due to the increase of discharge and suction pressures. The enthalpy difference with a refrigerant charge of 330 g is 4% higher than the refrigerant charge of 220 g. This will lead to high cooling capacity if the mass flow rate is the same. However, the highest cooling capacity of the oil-free VCR system for given conditions is achieved with a refrigerant charge of 280 g. This is mainly due to the variation of the mass flow rate with refrigerant charge. Fig. 4 shows the discharge and suction pressures against refrigerant charge with different condenser temperatures. For a fixed condenser temperature, both discharge and suction pressures increase with refrigerant charge. For a fixed refrigerant charge, discharge and suction pressures increase with condenser temperature. Higher suction and discharge pressures lead to a higher in-cylinder pressure resulting in a higher shaft force thus higher power input and copper loss. For a condenser temperature of 40 °C, the discharge and suction pressure is 9.8 bar and 2.8 bar, respectively, at a refrigerant charge of 220 g which is 23% lower than that of 330 g (12.7 bar and 3.6 bar). For a refrigerant charge of 220 g, both discharge and suction pressures at condenser temperature of 40 °C are 25% lower than 50 °C. Fig. 5 shows the condenser inlet and evaporator inlet temperatures against refrigerant charge with various condense temperatures at a pressure ratio of 3.5 and a compressor stroke of 13 mm. Both condenser and evaporator inlet temperatures increase with refrigerant charge. This is because higher refrigerant charge leads to higher discharge and suction pressures resulting in higher discharge temperature and evaporator inlet pressure thus higher condenser and evaporator inlet temperatures. With a refrigerant charge of 220 g, the oil-free VCR system can achieve an evaporator inlet temperature of 3.7 °C. With the increase of condenser temperature, both condenser and evaporator inlet temperatures present a growing trend. It is worth mentioning that the increase rate of evaporator inlet temperature against the increase of refrigerant charge is higher at low condenser temperatures. This is due
The volumetric efficiency was calculated according to:
ηV =
̇ g Tsuc 4mR πSfPsuc D2
(6)
where f is the operating frequency, Tsuc is the temperature at the compressor inlet, Psuc is the pressure at the compressor inlet, Rg is the specific gas constant and ṁ is the mass flow rate. Entropy and enthalpy for the refrigerant during refrigerant cycle were determined using CoolProp [21] based on the condenser outlet temperature, suction temperature, discharge pressure and suction pressure. Thus, the cooling capacity can be calculated as
Qċ = ṁ (h2 − h1)
(7)
where h1 is the enthalpy of the refrigerant at evaporator inlet and h2 is the enthalpy of the refrigerant at suction. Thus, energy efficiency ratio (EER) can be determined as
EER = 3.41
Ẇ Pin
(8)
4. Results and discussions Fig. 3 shows the pressure-enthalpy (p-h) diagram of the oil-free VCR system with various refrigerant charges for a pressure ratio of 3.5, a compressor stroke of 13 mm, a condenser temperature of 40 °C and an evaporator inlet temperature of 5 °C. The enthalpy difference between the evaporator inlet and outlet increases slightly with the increase of
Heater Power Meter Power Amplifier
Solenoid Valve
Linear Compressor
NI DAQ Card
Variac Heater Controller
Evaporator with Electric Heater
Expansion Valve
Condenser
Fig. 2. Complete experimental apparatus for the oil-free VCR system using R134a. 4
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Table 3 List of components for the oil-free VCR system. Components
Specifications
Compressor
Two identical oil-free linear compressors working in opposite, piston diameter of 19 mm, maximum compressor stroke of 14 mm, rated power of 100 W for each Coaxial water-cooled, copper, off-the-shelf, coolant connection diameter of 12.7 mm, refrigerant connection diameter of 16 mm Needle valve, stainless steel medium flow high pressure Copper, electric heater (resistance of 50 Ω), length of 128 cm , inner diameter of 7.9 mm, outer diameter of 12.7 mm R134a from BOC
Condenser Expansion valve Evaporator Refrigerant
condensation. The reduction of heat transfer efficiency decreases the quality of refrigerant circulating in the system. For a condenser temperatures of 40 °C and 45 °C the highest mass flow rates are achieved at a refrigerant charge of 280 g. For a condenser temperature of 50 °C, the highest mass flow rate is achieved at a refrigerant charge of 300 g. High operating frequency due to high suction and discharge pressures could also contribute to the increase of mass flow rate at high condenser temperature. The highest mass flow rate is 1.51 g/s at a condenser temperature of 50 °C with a refrigerant charge of 300 g while the lowest is 0.93 g/s at a condenser temperature of 40 °C with a refrigerant charge of 220 g. Fig. 9 shows the variation of volumetric efficiency with refrigerant charge at a pressure ratio of 3.5, a compressor stroke of 13 mm and condenser temperatures of 40 °C, 45 °C, and 50 °C. For a fixed condenser temperature, the variation of the volumetric efficiency has a similar trend to mass flow rate shown in Fig. 8. The oil-free VCR system can achieve a volumetric efficiency of 35.5% at a condenser temperature of 40 °C. For a fixed refrigerant charge, higher mass flow rate does not lead to higher volumetric efficiency. Mass flow rate is increased by 16% with the increase of condenser temperature from 40° to 50° while suction pressure is increased by 22%. The lower increment of the mass flow rate compared with the suction pressure lead to a decrease of volumetric efficiency with the increase of condenser temperature. Fig. 10 shows the variation of cooling capacity with refrigerant charge at a pressure ratio of 3.5, a compressor stroke of 13 mm and condenser temperatures of 40 °C, 45 °C, and 50 °C. The cooling capacity shows a similar trend with mass flow rate shown in Fig. 8. Both enthalpy difference and mass flow rate affect the cooling capacity. For the measurements shown in Fig. 3, the enthalpy difference only decreases by about 3% by increasing refrigerant charge from 220 g to 280 g while mass flow rate is increased by 25%. The cooling capacity reaches the maximum value while the system provides the highest mass flow rate. For the undercharged condition, the insufficient refrigerant in the system resulting in a small mass flow rate is the main reason for the low cooling capacity. For the overcharged system, the excessive liquid refrigerant contained in the condenser which reduces the heat transfer
to the small increment of saturation pressure at low temperatures. Fig. 6 shows the superheat and subcooling varying with refrigerant charge at a pressure ratio of 3.5, a compressor stroke of 13 mm and condenser temperature of 40 °C, 45 °C, and 50 °C. For a fixed condenser temperature, superheat decreases with the increase of refrigerant charge as the increase of suction pressure results in a rise in suction saturation temperature. For a fixed refrigerant charge, higher condenser temperature tends to have lower superheat due to the increase of discharge and suction pressures. Subcooling shows a reverse trend compared with superheat. The maximum superheat and subcooling is 26.5 K (at the refrigerant charge of 220 g) and 8.5 K (at the refrigerant charge of 330 g), respectively. Several researchers [5,22,23] pointed out that the subcooling of a system is an indicator of the quantity of the refrigerant in the condenser. Higher subcooling represents greater length of subcooled two-phase region thus higher quantity of refrigerant in the condenser. The variation of power input with refrigerant charge at a pressure ratio of 3.5, a compressor stroke of 13 mm and condenser temperatures of 40 °C, 45 °C, and 50 °C is shown in Fig. 7. At a fixed condenser temperature, the power input increases with the refrigerant charge due to the requirement of higher shaft force caused by the increasing incylinder pressure. The highest power input of the oil-free VCR system occurs at a refrigerant charge of 330 g and a condenser temperature of 50 °C. High refrigerant charge could lead to higher gas spring stiffness thus higher resonant frequency as well. Higher power input also cause higher copper loss and relatively lower motor efficiency. Fig. 8 shows the mass flow rate against refrigerant charge with various condenser temperatures at a compressor stroke of 13 mm and a pressure ratio of 3.5. For a fixed refrigerant charge, the mass flow rate decreases as the condenser temperature increases. This is because high condensation temperature causes more refrigerant vapour to circulate in the VCR system rather than condensing in the condenser. For undercharged system, the mass flow rate increases with refrigerant charge until reaching the optimal refrigerant charge. With the increase of the refrigerant charge, the increasing subcooling region occupies a larger portion of the condenser resulting in a reduction of the area for
Fig. 4. Discharge and suction pressure against refrigerant charge with various condenser temperatures at a pressure ratio of 3.5 and a compressor stroke of 13 mm. 5
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Fig. 5. Condenser inlet and evaporator inlet temperatures against refrigerant charge with various condenser temperatures at a pressure ratio of 3.5 and a compressor stroke of 13 mm.
Fig. 7. Variation of power input with refrigerant charge at a pressure ratio of 3.5, a compressor stroke of 13 mm and condenser temperatures of 40 °C, 45 °C, and 50 °C.
Fig. 6. Superheat (a) and subcooling (b) against refrigerant charge at a pressure ratio of 3.5, a compressor stroke of 13 mm and condenser temperatures of 40 °C, 45 °C, and 50 °C.
Fig. 8. Variation of mass flow rate with refrigerant charge at a pressure ratio of 3.5, a compressor stroke of 13 mm and condenser temperatures of 40 °C, 45 °C, and 50 °C.
area of the refrigerant vapour is the main reason for the cooling capacity deterioration. The highest cooling capacity of the oil-free system is 218.5 W at a condenser temperature 50 °C, a compressor stroke of 13 mm, a pressure ratio of 3.5 and a refrigerant charge of 300 g. Fig. 11 shows the variation of EER with refrigerant charge at a pressure ratio of 3.5, a compressor stroke of 13 mm and condenser temperatures of 40 °C, 45 °C, and 50 °C. For a condenser temperature of 40 °C, the EER increases with refrigerant charge before reaching the
optimal charge (280 g). With further increased refrigerant charge, the condenser flooding and high in-cylinder pressure result in a decrease of mass flow rate and high power input thus a decrease of EER. For a fixed refrigerant charge, the EER decreases with the increase of condenser temperature. Though higher condenser temperature tends to have higher cooling capacity, the excessive power input results in a decrease of EER. The maximum EER is 9.0 with a refrigerant charge of 280 g and 6
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Fig. 9. Volumetric efficiency against refrigerant charge for the oil-free refrigeration system at a pressure ratio of 3.5, a compressor stroke of 13 mm and condenser temperatures of 40 °C, 45 °C, and 50 °C.
Fig. 12. EER against refrigerant charge for the oil-free refrigeration system with various compressor strokes at a pressure ratio of 2.5and condenser temperatures of 40 °C, 45 °C, and 50 °C.
Fig. 10. Variation of cooling capacity with refrigerant charge at a pressure ratio of 3.5, a compressor stroke of 13 mm and condenser temperatures of 40 °C, 45 °C, and 50 °C.
Fig. 13. Variation of the optimal refrigerant charge for the oil-free refrigeration system with various pressure ratio and compressor stroke at a fixed condenser temperature of 40 °C.
Table 4 The optimal charge of the oil-free VCR system for various compressor stroke and condenser temperature at a pressure ratio of 2.5. Tcond (°C)
40 45 50
Stroke (mm) 9
10
11
12
13
250 None None
250 280 None
250 280 280
280 280 280
None None 300
of 2.5 and condenser temperatures of 40 °C, 45 °C, and 50 °C. For a fixed refrigerant charge, the EER increases with compressor stroke due to the improvement of the volumetric efficiency. For a fixed compressor stroke, the EER increases firstly with the refrigerant charge until reaching the optimal charge. With an increase of the compressor stroke, the optimal charge of the oil-free VCR system increases. Fig. 13 shows the optimal charge of the oil-free VCR system for various operating conditions at a condenser temperature of 40 °C. For a fixed compressor stroke, the optimal charge decreases with the pressure ratio. For a fixed pressure ratio, the optimal charge increases with the compressor stroke. This is because a higher compressor stroke and lower pressure ratio tend to have a higher volumetric efficiency resulting in an increase of mass flow rate. Table 4 lists the optimal charge of the oil-free VCR system for various compressor strokes and condenser temperatures at a pressure ratio of 2.5. At a condenser temperature of
Fig. 11. EER against refrigerant charge for the oil-free refrigeration system at a pressure ratio of 3.5, a compressor stroke of 13 mm, and condenser temperatures of 40 °C, 45 °C, and 50 °C.
a condenser temperature of 40 °C. By reducing the refrigerant charge by 11.3%, the EER of the system is reduced by 7.5%. Thus, to operate the system under a slightly undercharged condition may be considered to reduce the refrigerant charge thus leakage. Fig. 12 shows the EER against refrigerant charge for the oil-free refrigeration system with various compressor strokes at a pressure ratio 7
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40 °C, by increasing compressor stroke by 25%, the optimal refrigerant charge is increased by 11%. The same amount of the increment of optimal refrigerant charge can be achieved by increasing 20% (40–50 °C) of condenser temperature. The optimal refrigerant charge is more sensitive on condenser temperature than compressor stroke. A higher condenser capacity tends to have a higher optimal refrigerant charge. Ghoubali et al. [11] reported that using microchannel heat exchanger as condenser can reduce 13% of the total refrigerant charge concentrating in the condenser. Moreover, more than 18% refrigerant can be contained in liquid line [2]. To minimize the diameter of the liquid line also can be a potential choice for the refrigerant charge reduction.
(UA016-05) for supporting the instrumentation of the VCR system is acknowledged. The authors are also grateful to Prof Richard Stone, Dr Mike Dadd and Paul Bailey at the University of Oxford for the prototype oil-free linear compressor. The authors are grateful to Andy White and Alex Burns at the University of Sussex for the lab support.
5. Experimental uncertainty
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Appendix A. Supplementary material Supplementary data to this article can be found online at https:// doi.org/10.1016/j.applthermaleng.2019.114473. References
Pressures, temperature, stroke, current, voltage, and mass flow rate were measured during the experiments. Typically, a set of readings was taken every 15 min to allow time for thermal equilibrium to be attained. The measurements of pressure, stroke, and temperature have absolute uncertainties of 0.015 bar, 0.025 mm, and 0.5 °C. The current transducer, voltage transducers, and mass flow rate have accuracies of 0.5%, 0.5%, and 1% respectively. The combined uncertainties of the calculated values were calculated using a 95% confidence interval. The cooling capacity, EER, and volumetric efficiency have relative uncertainty of 1%, 1.3%, and 0.3%, respectively. 6. Conclusions Operating a VCR system with an optimal refrigerant charge can reduce the refrigerant leakage without losing capacity and energy efficiency. In this study, the optimal refrigerant charge and energy efficiency of an oil-free refrigeration system using R134a were investigated for a wide range of operating conditions. The key findings are shown below: (1) For given pressure ratio and stroke, and compressor stroke, the discharge and suction pressure increase with the increase of the refrigerant charge. This leads to an increasing power input and copper loss. (2) At a fixed condenser temperature, compressor stroke, and pressure ratio, the superheat increases with the refrigerant charge while subcooling shows a reverse trend due to the increase of the suction and discharge pressures. (3) At a fixed condenser temperature, the mass flow rate increases first with the refrigerant charge, then decreases due to the flooding in the condenser and the increase of the in-cylinder pressure. (4) For a fixed compressor stroke and a fixed pressure ratio, both cooling capacity and EER reach maximum value at the optimal refrigerant charge. The cooling capacity decreases with condenser temperature while EER increases with condenser temperature. (5) The optimal charge varies with operating conditions. A higher compressor stroke and a lower pressure ratio tend to have a higher optimal refrigerant charge. Declaration of Competing Interest The authors declared that there is no conflict of interest. Acknowledgement The Research Development Fund from the University of Sussex
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