Heat Mass Transfer DOI 10.1007/s00231-017-2173-6
ORIGINAL
Experimental evaluation of refrigerant mass charge and ambient air temperature effects on performance of air-conditioning systems Mahdi Deymi-Dashtebayaz 1 & Mehdi Farahnak 2 Arash Ghalami 3 & Nima Mohammadi 3
&
Mojtaba Moraffa 1 &
Received: 24 March 2017 / Accepted: 20 September 2017 # Springer-Verlag GmbH Germany 2017
Abstract In this paper the effects of refrigerant charge amount and ambient air temperature on performance and thermodynamic condition of refrigerating cycle in the split type airconditioner have been investigated. Optimum mass charge is the point at which the energy efficiency ratio (EER) of refrigeration cycle becomes the maximum. Experiments have been conducted over a range of refrigerant mass charge from 540 to 840 g and a range of ambient temperature from 27 to 45 °C, in a 12,000 Btu/h split air-conditioner as case study. The various parameters have been considered to evaluate the cooling rate, energy efficiency ratio (EER), mass charge effect and thermodynamic cycle of refrigeration system with R22 refrigerant gas. Results confirmed that the lack of appropriate refrigerant mass charge causes the refrigeration system not to reach its maximum cooling capacity. The highest cooling capacity achieved was 3.2 kW (11,000 Btu/h). The optimum mass charge and
* Mehdi Farahnak
[email protected] Mahdi Deymi-Dashtebayaz
[email protected] Mojtaba Moraffa
[email protected] Arash Ghalami
[email protected] Nima Mohammadi
[email protected] 1
Department of Mechanical Engineering, Hakim Sabzevari University, Sabzevar, Iran
2
Young Researchers and Elite Club, Rasht Branch, Islamic Azad University, Rasht, Iran
3
Iran Energy Efficiency Organization (IEEO-SABA), Tehran, Iran
corresponding EER of studied system have been obtained about 640 g and 2.5, respectively. Also, it is observed that EER decreases by 30% as ambient temperature increases from 27 °C to 45 °C. By optimization of the refrigerant mass charge in refrigerating systems, about 785 GWh per year of electric energy can be saved in Iran’s residential sector.
1 Introduction All air-conditioners and refrigerators rely on the correct charge or amount of refrigerant gas in their systems, to work correctly. Air-conditioning units and refrigerators are designed to operate correctly with a predetermined charge of refrigerant charge. Today, one of the main goals of researchers and designers is the improvement of efficiency of industrial equipment. Accordingly, efficiency improvement and power consumption reduction are the main concerns in refrigeration industry. Determination of optimum refrigerant mass charge has significant impacts on power consumption and performance of compression refrigeration cycles. As it is known, the lack of enough refrigerant mass charge causes the refrigeration system not to reach its maximum capacity [1–3]. In this condition, the mass flow rate in the system is not sufficient and therefore the cooling effect is not high. A further increase of mass charge increases the cooling capacity and EER reaches its maximum value [4–6]. Again, increase of the mass charge leads to more power consumption of compressor, and as a result EER decreases. Many studies have been reported the reduction of refrigerant charge and the impact of it on the coefficient of performance (COP) and on the cooling capacity in refrigerating systems, as reviewed by F. Poggi et al. [7]. Melo and Boeng [8] determined the optimum mass charge of R600a in a compression refrigeration cycle. Their study showed that by
Heat Mass Transfer
increasing the refrigerant charge, the required work always increases linearly. Furthermore, the COP and cooling capacity of refrigeration cycle first increase and then decrease, with increase of the mass charge. Sarntichartsak et al. [9] investigated the optimum performance of refrigeration system for different capillary tubes and refrigerant mass charges. Their simulation modeling with R-22 and R − 407C had good agreement with experimental results. In another study, Vjacheslav et al. [10] evaluated the optimal refrigerant charge into refrigerating systems considering a rationally based algorithm. Their results of modeling using R410A as a refrigerant showed that the performance of the system is highly dependent on refrigerant charge. The results reported by Padalkar et al. [11] indicated the performance of R290 as an efficient substitute refrigerant to R22 in a split air-conditioner for different operating conditions. Improving the COP of refrigeration systems has always attracted interests of researchers in different aspects of this area of study [12–16]. One of these approaches is the performance study and optimization of heat exchangers [17–19]. F.W. Yu and K.T. Chan [20, 21, and] developed a thermodynamic model for controlling the condenser fans in air-cooled centrifugal chillers in order to maximize the COP under various operating strategies and used an experimental work verify modeling results. Utilization of evaporatively cooled air condenser as an alternative to air-cooled condenser was specially studied by E. Hajidavalloo and H. Eghtedari [22] and their experimental tests at different ambient air temperatures showed remarkable improvement in performance of split type air-conditioners by reducing the condenser temperature, which leads to decrease the compressor work and higher cooling capacity. Besides, a considerable amount of work has been devoted to the study of design of capillary tubes and flow characteristics of refrigerants flowing through capillary tubes [23–28]. In addition, some authors included specially the effect of ambient temperature on performance of refrigeration systems in their studies [29–31]. HCFC-22 (also known as R22) is still widely used in residential heat pump, residential, commercial and industrial airconditioning and refrigeration systems in Iran and new gases cannot be used in many of the older air-conditioning units. Some manufacturers claim that using modern refrigerant in converted air-conditioning systems can lead to poorer performance and so higher energy costs. Therefore, the R22 phase-out involves replacing all parts of air-conditioning infrastructure, including outdoor and indoor units, pipework and electrical wiring which increases initial capital outlay and installation time. So the task of phasing out R22 can be considered to be a major challenge but government of Iran is planning for R22 phase-out now. Although, there are some studied reported the alternative refrigerants without making main changes in the refrigeration systems considering cooling capacity, COP, exergy destruction, refrigerants thermodynamic properties, etc. [29, 32–35].
In present experimental study, the effect of variation in refrigerant mass charge and ambient air temperature on performance of refrigeration system has been investigated. A 12000 Btu/h air-conditioner has been chosen as case study which is a conventional single split air-conditioning system in residential sector of Iran. The refrigeration system has been charged with R22 refrigerant varying from 540 g to 840 g. The experimental tests consisted of controlling and measuring the system and component operating conditions. In order to do this, the refrigeration system was properly instrumented and monitored to record the steady-state behavior of different components of the system. Also, the ambient temperature effect tests were conducted at different temperatures of ambient air (27 °C to 45 °C) while the refrigerant mass charge was fixed to the optimum value obtained from the previous tests. Key measured parameters were used to evaluate the cooling capacity and energy efficiency ratio (EER) of the system and etc.
2 Experimentation 2.1 Experimental set-up and test condition The experimental consists of compressor, fan cooled condenser and evaporator, and expansion device. An existing single split air-conditioner with a rated cooling capacity of 12,000 Btu/h (3.5 kW) was used in experimental tests. The refrigeration system was equipped with a rotary type compressor which operates with R22, and a capillary tube type expansion device with tube length and diameter of 55 cm and 3 mm, respectively. The condenser and the evaporator were finned-tube heat exchangers with louvered fins. The lengths of the condenser and evaporator tubes were 16 m and 14 m, respectively. The outer and inner diameters of condenser and evaporator were similarly equal to 7 mm and 4.6 mm, respectively. All the connected pipes were thermally insulated by foam. A schematic diagram of indoor and outdoor chamber of the experimental setup is shown in Fig. 1. The refrigeration system was run at the standard test conditions. The dry bulb and wet bulb temperatures of the outdoor air chamber were kept at 35 °C and 24 °C respectively, while the dry bulb and wet bulb temperatures of the indoor air chamber were kept at 27 °C and 19 °C respectively, by control unit. For testing the effect of refrigerant mass charge, experiments were conducted with different amounts of refrigerant mass charge at a constant ambient air temperature of 35 °C. In order to test the effect of ambient temperature on performance of refrigeration system, while the optimum mass charge obtained from the previous test results was charged into system, the ambient air temperature was varied. The tested mass charges and ambient temperatures are summarized in Table 1.
Heat Mass Transfer Fig. 1 Schematic diagram of experimental set-up
Outdoor Chamber
DBT=35 °C WBT=24 °C
Indoor Chamber
DBT=27 °C WBT=19 °C
Condesnser
4 2
3 Capillary Tube
1 Accumulator Data Logger Evaporator
Compressor
The performance factors evaluated in this study were cooling capacity, energy efficiency ratio, power consumption and refrigerant mass charge inside the system. The cooling capacity measurement was based on air enthalpy difference calculated from the measured dry bulb and wet bulb temperatures of entering and leaving air flowrates. Measurement of air volume flow rate was carried out by a nozzle type air flow measuring device. Figure 2 shows the test facility of Fig. 1 in use. 2.2 Measurements The pressure and temperature of the refrigerant-side at four points were measured by four pressure sensors with an accuracy of ±1% and four Pt100 temperature sensors with an accuracy of ±0.5 °C, respectively. Two pressure sensors and two temperature sensors were mounted before each of the evaporator and condenser, and two pressure sensors and two temperature sensors were mounted after each of the evaporator and condenser. HC2-S Rotronic temperature and relative humidity probe was used for the air-side temperature and humidity measurements with the accuracies of ±0.1 °C and ±0.8% respectively. Two temperature sensors were placed both at inlet and outlet air stream of condenser. Also, four temperature sensors were located at inlet air stream of evaporator and three temperature sensors were located at outlet air stream of it. The locations of temperature and pressure sensors in the test facility are shown in Fig. 1. Power measurements including fan,
compressor, and total power were measured by a digital wattmeter with an accuracy of ±0.1% in order to measure the input power, voltage and electrical current. Air volume flow rate was measured by a nozzle type air flow measuring device with accuracy of 1%. The characteristics of the measurement devices used are summarized in Table 2. The experimental data including the pressure and temperature readings of refrigerant and air, as well as air volume flow rate and wattmeter reading were recorded continuously. Data were sampled over the period of 30 min with 20 s intervals, under steady state condition (after 20 min). A data logger was used to collect data from the testing apparatus, which was connected to a computer where the data were visually displayed during experimentation. 2.3 Experimental procedure In this experiment, the data sets were collected for different quantities of refrigerant charges and ambient temperature, taken after 30 min operation of the system. The procedure for testing the effect of refrigerant charge is based on performance test which is made with the system loaded with various refrigerant charges at ambient temperature of 35 °C. Initially, the system was completely evacuated from any refrigerant or moisture in the lines and then R22 refrigerant was carefully weighed and added to the system. For testing the effects of refrigerant overcharge and undercharge, the tests were conducted by changing the refrigerant mass charge from 540 to
Table 1 Test conditions Type of test
Refrigerant charge, m (g) Ambient air temperature, Tamb (°C)
Mass charge effect
Ambient temperature effect
540, 610, 640, 670, 740, 840 35
640 27, 30, 35, 40, 45
Heat Mass Transfer Fig. 2 Indoor/outdoor test facility (SABA Lab. [36])
840 g. At first, the system was loaded and experimented with an initial refrigerant mass charge of 540 g, and then following this initial test, the above experimental procedure and measurement were repeated for 610, 640, 670, 740 and 840 g of refrigerant, respectively. The ambient temperature effect tests were carried out at different temperatures of ambient air while the refrigerant mass charge was fixed to the optimum value obtained from the previous tests. The experiments were performed at four ambient air temperatures: 27, 30, 35, 40 and 45 °C.
function of air flow rate, enthalpy of air entering and leaving the indoor side, specific volume and specific humidity of air. Q˙ tc ¼
P1 þ
W˙ t ¼ V t I t cosφ
ρV 21 ρV 2 ¼ P2 þ 2 2 2
ð3Þ
Assuming uniform air velocity profiles, the continuity equation can be expressed as: V˙ a ¼ V 1 A1 ¼ V 2 A2
ð4Þ
Combining Eqs. (3) and (4), air flow rate will be: vffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi u 2ðP −P Þ u 1 2 V˙ a ¼ A2 u 2 t ρ 1− AA21
ð5Þ
ð1Þ
Cooling capacity of the refrigerating system which is the heat transfer rate in the evaporator can be obtained by Eq. (2). It is calculated using air-enthalpy Indoor method [37] as a
Energy efficiency ratio (EER) of the refrigeration system is calculated as a ratio between cooling capacity and total power usage as below, EER ¼
Table 2
ð2Þ
As it is mentioned, the measurement of air volume flow rate was conducted using a nozzle type air flow measuring equipment. The nozzle flow rate meters use the Bernoulli equation to calculate fluid flow rate using pressure difference through obstructions in the flow. The Bernoulli equation for a horizontal air flow is as follow:
3 Data reduction A computerized data-logging system recorded the measurement data from the sensors at predefined time intervals (Fig. 3). The temperature, pressure and flow rate of refrigerant and air collected at specified locations in the refrigeration system by sensors, are used to calculate and evaluate system performance at undercharged/overcharged condition by using following equation: The total power consumption of the refrigeration system is calculated as Eq. (1), which is the sum of the power consumptions of all components including compressor and fans.
V˙ a ðhai −hao Þ ν a ð1 þ ωa Þ
˙ Q tc W˙ t
ð6Þ
Characteristics of the measurement devices
Instrument
Measuring range
Accuracy
HC2-S rotronic temperature and relative humidity probe Temperature sensor (Pt100) Pressure sensor wattmeter
-40–60 °C 0–100% RH −50–150 °C 0–40 bar 0–6000 W
±0.1 °C ±0.8% ±0.5 °C ±1% ±0.1%
4 Results and discussion 4.1 Effect of refrigerant mass charge In order to study the impact of refrigerant overcharging and undercharging on performance of the refrigeration system, the experimental procedure above-mentioned was carried out for
Heat Mass Transfer Fig. 3 Data acquisition and monitoring system (SABA Lab. [36])
different values of mass charge. The thermodynamic properties of R22 required for the calculations were obtained through Engineering Equation Solver (EES) [38]. The thermodynamic cycle of experimental results under test condition is shown on the Pressure-Enthalpy diagram in Fig. 4. Points 1, 2, 3 and 4 in Fig. 4 correspond to the same points indicated in Fig. 1. As shown in Fig. 4, by increasing the refrigerant mass charge the pressure and temperature in condenser and evaporator reach the desired values. At low mass charges, a considerable part of evaporator is filled by
superheated vapor refrigerant and also the mass flow rate is not sufficient which together lead to a poor cooling effect. When refrigerant mass charge increases, the liquid-phase content in condenser will increase and the vapor-phase content will decrease and the outlet of the evaporator approaches the saturated vapor line. Therefore, the promotion of operational performance can be achieved by the desired values of temperature and pressure. Although the mass flow rate of the refrigerant and cooling effect are higher at the overcharged conditions, but the higher pressures of condenser and evaporator
Heat Mass Transfer
Fig. 4 P-h diagram of refrigeration system for different refrigerant mass charges at ambient temperature 35 °C
will increase the compressor power consumption and as a result the efficiency of the system will decrease. In Fig. 5, the suction and discharge temperatures of compressor for various mass charges are demonstrated. It is obvious that suction and discharge temperatures decrease with the increase of mass charge. This is due to the more liquid charge at evaporator, which reduces the superheating of vapors at the suction of the compressor and as a result reduces the temperature of superheated vapors entering the condenser. As mentioned before, by increasing the refrigerant mass charge, the amount of subcooling at condenser outlet will increase. Condenser subcooling is an indicator of how much refrigerant charge is in the refrigeration system. The lower the refrigerant charge, the lower the subcooling, and vice versa. This effect can be seen in Fig. 6. As the refrigerant mass charge increases, the condensation pressure increases due to an accumulation of refrigerant and the degree of subcooling at condenser rises. The effect of refrigerant charge on suction (evaporator) and discharge (condenser) pressures as well as pressure ratio is shown in Fig. 7. As it is shown in the figure, compressor inlet and exit pressure both increase with increasing the refrigerant mass charge. However, the change in exit pressure is less than inlet
Fig. 5 Variation of suction and discharge temperatures with refrigerant mass charge at ambient temperature 35 °C
Fig. 6 Effect of mass charge on condenser subcooling at ambient temperature 35 °C
pressure which causes the pressure ratio to drop. As known, Lower pressure ratio leads to higher system performance in terms of compressor volumetric efficiency [29]. Figure 8 shows the cooling capacity of the refrigeration system as a function of refrigerant mass charge under test conditions. It can be seen that the cooling ability increases with mass charge until a maximum value is reached, and then it decreases. Lack of enough refrigerant will cause high superheat in evaporator and poor cooling effect. Refrigerant overcharging will increase the system operating pressure and temperature and will decrease the cooling capacity of the system. Overcharging of the system results in overflowing of the condenser. The condensation and evaporation pressures will increase and as shown in Fig. 4, the enthalpy difference between evaporator inlet and exit (Q˙ tc ¼ h1 −h4 ) will decrease. The impact of refrigerant mass charge on power input is displayed in Fig. 9. As it is shown, higher is the refrigerant mass charge, higher is the compressor power consumption. This is because the increasing charge results in an increasing pressure of condenser and evaporator (decreasing of pressure ratio) which increases the volumetric efficiency of the compressor, and therefore the electric power consumption of compressor increases.
Fig. 7 Variation of suction and discharge pressures (left axis), and pressure ratio (right axis) with refrigerant mass charge at ambient temperature 35 °C
Heat Mass Transfer
Fig. 8 Variation of cooling capacity with refrigerant mass charge at ambient temperature 35 °C
Fig. 10 Variation of energy efficiency ratio with refrigerant mass charge at ambient temperature 35 °C
Figure 10 shows the variation of energy efficiency ratio (EER) of air-conditioning system with respect to the mass charge, where a maximum EER of 2.48 was obtained at a charge mass of 640 g. increase and decrease of the EER before and after maximum amount are caused by rise of cooling capacity and compressor work, respectively. At the left of the optimum mass charge, a slight rise of the EER can be observed which is related to effective cooling capacity. The sharp drop of EER after optimum mass charge is due to overcharging of the system and overflowing of the condenser. All the obtained results correspond to the results reported by other authors [1, 23, 25].
and performance parameters were calculated. Figure 11 shows the test results on the Pressure-Enthalpy diagram in terms of ambient air temperatures at optimum refrigerant charge. When ambient temperature increases, the temperature and pressure of condenser and evaporator as well as pressure ratio increase. Therefore as shown in Fig. 12, the compressor work will increases 0.017 kW per 1 °C, linearly. In lower ambient temperatures, the cooling effect (enthalpy difference between inlet and outlet of the evaporator) is larger, as can be seen in Fig. 11. Variation of cooling capacity with respect to the ambient air temperature is also demonstrated in Fig. 12. The cooling capacity decreases by about 0.022 kW for every 1 °C temperature rise. Decreasing cooling capacity together with increasing power consumption will result in energy efficiency loss of refrigeration system. As it can be seen in Fig. 13, by increasing the ambient temperature from 27 to 45 °C, energy efficiency ratio (EER) decreases by 30%. Similar results were reported in literature by several researchers based on their studies [22, 29–31].
4.2 Effect of ambient air temperature In this study, the tests were performed with different ambient temperatures while the refrigerant mass charge was fixed at optimum value of 640 g obtained from previous test results. According to the ambient temperature test results, thermodynamic properties of the refrigerant at indicated points of the cycle were obtained
Fig. 9 Variation of total power input with refrigerant mass charge at ambient temperature 35 °C
Fig. 11 P-h diagram of refrigeration system for different ambient air temperatures at optimum refrigerant charge (640 g)
Heat Mass Transfer
Fig. 12 Variation of cooling capacity (left axis) and total power input (right axis) with ambient temperature at refrigerant charge of 640 g
4.3 Power saving Based on these results as shown, it can be concluded that large savings potentials exist through optimization of refrigeration systems. According to reports of Iran Energy Efficiency Organization (IEEO-SABA) [36], there are 5,000,000 air-
Fig. 13 Variation of energy efficiency ratio with ambient temperature at refrigerant charge of 640 g
conditioners in Iran’s residential sector. Assuming equal distribution for all the capacities, by improving EER from 2 to 2.5 through optimization of refrigerant mass charge in only 20% of installed air-conditioners, about 785 GWh per year of electric energy can be saved, averagely. It can be calculated from Eqs. (7) and (8) and Fig. 14, as below:
ΔW t ¼ 4ðmonths require coolingÞ 30ðdays of a monthÞ 12ðaverage cooling hoursÞ kW power saving ΔW˙ t ; Fig:14
ð7Þ
P ðΔW t9000 þ ΔW t12000 þ ΔW t18000 þ ΔW t24000 þ ΔW t30000 Þ → ΔW t ¼ 5 20% 5000000 ð20% of installed airconditionersÞ ð8Þ
5 Conclusions
Fig. 14 Amount of kW power saving by improving EER from 2 to 2.5 through optimization of refrigeration systems, for different conventional cooling capacity of air-conditioners in Iran’s residential sector
The performance of a 12,000 Btu/h split air-conditioning system was experimentally investigated over a range of refrigerant mass charge from 540 to 840 g and a range of ambient temperature from 27 to 45 °C. The experimental results show that the optimum mass charge has a substantial effect on performance parameters of the refrigeration system, such as cooling capacity, power consumption and energy efficiency ratio (EER). It is observed that there is an optimum value of EER when refrigerant charge changes. The highest cooling capacity and EER were obtained were 3.2 kW (11,000 Btu/h) and 2.5 respectively, in 640 g of refrigerant mass charge. Also, with increase of ambient air temperature
Heat Mass Transfer
from 27 °C to 45 °C, EER decreases by 30%. The results from this study agreed well with previous studies reported by other researchers. It is found that about 785 GWh per year of energy saving can be achieved by optimization of refrigerant mass charge in air-conditioners installed in Iran’s residential sector. We strongly believe that the first step towards the standardization of refrigeration industry -in Iran and every other developing country- is to retrofit the current infrastructures and use the potential savings towards modernization of refrigeration industry, as the next step. Acknowledgements The authors would like to thank Iran Energy Efficiency Organization (IEEO-SABA) for providing air-conditioning laboratory services and their technical and financial support during this project. Nomenclature ΔEER, EER improvement; ΔTSC, condenser subcooling; ΔW ̅t, average energy saving per year for 20% of installed air-conditioners with different capacities (kWh/yr); ΔWti, energy saving per year for airconditioner with cooling capacity i (kWh/yr); ΔẆti, power saving for airconditioner with cooling capacity i (kW); ρ, density (kg/m3); ωa, specific humidity of air, kg/kg of dry air; A, area (m2); COP, coefficient of performance; cos φ, power factor; DBT, dry bulb temperature (°C); EER, energy Efficiency Ratio; h, kJ/kg; hai, enthalpy of air entering the indoor side, kJ/kg of dry air; hao, enthalpy of air leaving the indoor side, kJ/kg of dry air; It, total current; m, refrigerant mass charge (m); P,, pressure (bar); Q̇ tc, total cooling capacity; T, temperature (°C); Tamb, ambient air temperature; V, velocity (m/s); V̇ a, Indoor air-flow rate, m3/s; va, specific volume of air at point of measurement of air-water vapour mixture, m3/kg; Vt, total voltage; WBT, wet bulb temperature (°C); Ẇt, total input power.
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