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Performance optimization of an R410A air-conditioner with a dual evaporator ejector cycle based on cooling seasonal performance factor
Accepted Manuscript Performance optimization of an R410A air-conditioner with a dual evaporator ejector cycle based on cooling seasonal performance factor Sunjae Kim, Yongseok Jeon, Hyun Joon Chung, Yongchan Kim PII: DOI: Reference:
S1359-4311(17)33447-6 https://doi.org/10.1016/j.applthermaleng.2017.12.012 ATE 11533
To appear in:
Applied Thermal Engineering
Received Date: Revised Date: Accepted Date:
20 May 2017 19 October 2017 2 December 2017
Please cite this article as: S. Kim, Y. Jeon, H. Joon Chung, Y. Kim, Performance optimization of an R410A airconditioner with a dual evaporator ejector cycle based on cooling seasonal performance factor, Applied Thermal Engineering (2017), doi: https://doi.org/10.1016/j.applthermaleng.2017.12.012
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Performance optimization of an R410A air-conditioner with a dual evaporator ejector cycle based on cooling seasonal performance factor
Sunjae Kim, Yongseok Jeon, Hyun Joon Chung, Yongchan Kim* Department of Mechanical Engineering, Korea University Anam-Dong, Sungbuk-Ku, Seoul 136-713, Republic of Korea
ABSTRACT Even though a dual evaporator ejector cycle (DEEC) offers several advantages over a standard two-phase ejector cycle, few experimental investigations of the performance of the DEEC are available in the literature. This study presents the performance characteristics of an R410A air-conditioner adopted with a DEEC under various operating conditions and ejector geometries. The COP of the DEEC decreased with an increase in entrainment ratio (ER) due to the decrease in pressure lifting ratio. For the optimum ER, the effectiveness of the DEEC increased with an increase in compressor speed with a larger total mass flow rate. The optimum mixing section diameter was determined to be 5 mm based on the cooling seasonal performance factor (CSPF) and CSPFbin of the DEEC. The maximum allowable limit for the ER was also suggested to be 0.3. In addition, the CSPF of the DEEC was 6.3% higher than that of the baseline cycle at an ER of 0.1.
Keywords: Dual evaporator ejector cycle, Ejector, COP, CSPF
* Corresponding author. Tel.: +82-2-3290-3366; Fax: +82-2-921-5439 E-mail address: [email protected]
Nomenclature COP
coefficient of performance
COPi
COP ratio of the ejector cycle to the baseline cycle
CSPF
cooling seasonal performance factor
CSPFbin
cooling seasonal performance factor for a specific bin temperature
D
diameter (mm)
DEEC
dual evaporator ejector cycle
ER
entrainment ratio
L
length (mm)
MFR
mass flow meter mass flow rate (kg h-1)
n
number of temperature bins
nj
bin hours
P
pressure (kPa)
PLF
part load factor
PLR
pressure lifting ratio
SC
subcooling (°C)
SH
superheat (°C)
TA
bin outdoor temperature (°C)
Q
cooling capacity (W h) cooling capacity rate (W)
X
ratio of building load to cooling capacity at outdoor temperature
Subscripts c
cooling
high-temp
high temperature evaporator
low-temp
low temperature evaporator
m
mixing section
mot
motive
out
outlet
sec
secondary flow
1. Introduction Even though air-conditioning systems have contributed to human well-being in many aspects, they have a considerable impact in terms of various environmental and energy problems, such as ozone layer depletion and global warming [1]. In the air-conditioning industry, it is crucial to develop environmentally friendly and energy-efficient airconditioners [2–5]. Renewable energies have been introduced by many researchers to diversify the energy source of air-conditioning systems [6,7]. In addition, refrigerant injection technology has been studied to increase the COP and reliability of a heat pump system in severe outdoor conditions [8–10]. Furthermore, the expansion valve has the potential for further advances because it suffers a significant energy loss during the isenthalpic expansion process [11–13]. Therefore, an ejector has been introduced to retrieve the energy loss during the expansion process in conventional expansion devices [14,15]. Previous ejector studies [16–19] have focused on a standard two-phase ejector cycle that consists of a compressor, an ejector, a condenser, and a liquid-vapor separator. In contrast, Oshitani et al. [20] proposed a dual evaporator ejector cycle (DEEC) as a climate control system of the cabin in automobiles. They found that the DEEC had better operation stability and adaptability with low-pressure refrigerants compared to the standard two-phase ejector
cycle. As shown in Fig. 1, the DEEC consists of a compressor, an ejector, a condenser, and two evaporators each with different evaporating temperature. In the DEEC, the refrigerant at the condenser outlet is divided into two separate streams. One of them flows directly into the ejector passing through the motive nozzle, and the other stream (secondary flow loop) is isenthalpically expanded through the needle valve and is then vaporized in the low temperature evaporator. The refrigerant at the secondary loop is entrained into the secondary block and is then mixed with the motive flow in the mixing section. The mixed two-phase flow recovers pressure energy while passing through the diffuser section and then passes through the high temperature evaporator to superheat the refrigerant. Lawrence and Elbel [21,22] observed that a DEEC showed maximum COP improvements of 12% and 8% using R1234yf and R134a, respectively. In addition, Wang et al. [23] compared the performance of alternate ejector cycles, including the DEEC. A transcritical cycle using R744 (CO2) usually operates with extremely high condensing pressure causing not only large compression ratio but also large expansion loss. A two-phase ejector cycle has been adopted to overcome large expansion loss of the transcritical cycle using R744. Kornhauser [24] presented a one-dimensional model for an R12 refrigeration system with a two-phase ejector. Kornhauser’s model was utilized to study the COP improvement of R744 ejector cycles [25–27]. Takeuchi et al. [28] reported a COP improvement of 20% using an R744 ejector cycle for an automobile air-conditioning system. Elbel and Hrnjak [16] experimentally reported an improvement in COP and cooling capacity by up to 7% and 8%, respectively. More recently, Nakagawa et al. [17] observed a COP improvement of 26% by optimizing the mixing section length of an ejector in an R744 refrigeration system. In addition, the ejector cycles with low-pressure refrigerants can still offer noticeable COP improvements even though those have received relatively less attention due to a lower work recovery potential. Harrell and Kornhouser [18] reported COP
improvements of 4% to 8% in an R134a ejector cycle against a conventional vapor compression cycle. Disawas and Wongwises [19] presented an increase in COP and cooling capacity in a two-phase ejector cycle over a conventional R134a vapor compression cycle. Most recently, Pottker and Hrnjak [29] observed a COP improvement of 8.4% in an ejector cycle with a liquid fed evaporator system due to the work recovery effect. Even though the DEEC offers several advantages over the standard two-phase ejector cycle [21,22], it has been mainly adopted in refrigerators with two-evaporating temperatures. The DEEC becomes more beneficial with an increase in refrigerant mass flow rate, which can be used in air-conditioners with relatively higher cooling capacity. However, few experimental investigations on the performance characteristics of a DEEC-based air-conditioner have been presented in the literature. In particular, the effects of the ejector geometries and operating conditions on the performance of the DEEC have not been comprehensively analyzed. Furthermore, the ejector optimization of the DEEC based on the cooling seasonal performance factor (CSPF) is essential in order to cope with various cooling loads. However, the optimization of the DEEC based on the CSPF has been rarely presented in the literature. Thus, this study investigates the performance characteristics of a DEEC-based R410A airconditioner according to the operating conditions and ejector geometries. The performance of the DEEC was measured and optimized by varying the compressor speed, entrainment ratio (ER), outdoor temperature, and mixing section diameter and length. The COP and cooling capacity of the DEEC were compared to those of the baseline cycle using a conventional expansion device. In addition, the CSPF of the DEEC was compared to that of the baseline cycle in order to optimize the ejector geometries under various cooling loads.
2. Experimental setup and test conditions 2.1. Experimental setup
Fig. 2 shows a schematic diagram of the experimental setup used to measure the performance of the DEEC. The testing setup for the DEEC consisted of a rotary compressor with an inverter driver, a condenser, two-evaporators, an ejector, and a needle valve. A plate type heat exchanger was used as the condenser to control the condensing pressure and heat transfer rate, and a water chiller was connected to the condenser in order to remove the heating load. The needle valve was utilized as an expansion device in the secondary loop of the DEEC. The cooling load into evaporator #1 was controlled using cartridge heaters, adjusting the evaporating pressure and heat transfer rate. The cooling load into evaporator #2, which had a higher evaporating temperature than the evaporator #1, was also controlled using cartridge heaters. The DEEC can be used to provide the required cooling capacity for each air-conditioning space with different temperatures. In addition, two evaporators can be combined serially as one unit. In that case, the temperature gliding in the combined evaporator can be matched closely with that in the air stream, resulting in a higher heat transfer effectiveness. As shown in Fig. 3, the ejector consisted of a motive nozzle, a secondary block, a mixing section, and a diffuser. Each ejector component was manufactured separately, and then all pieces were assembled together. The ejector geometries varied by diversifying the selection of the modules, and the throat diameter of the motive nozzle was controlled by adjusting the ejector needle. The ejector needle and the needle valve on the secondary line were manually controlled to adjust the ER. To increase the ER, the ejector needle was moved forward to reduce the motive flow rate, while the needle valve was partially opened to increase the secondary flow. Geometrical specifications of the tested ejector are listed in Table 1. In addition, the DEEC performance was compared by constructing the baseline cycle using a conventional air-conditioner consisting of a rotary compressor with an inverter driver, a condenser, an evaporator, and a needle valve as an expansion device. The baseline cycle had a
single evaporating temperature to make a performance comparison of the novel cycle with the conventional air-conditioner cycle, as that proposed by Lawrence and Elbel [21,22]. Table 2 summarizes the component specifications of the DEEC and the baseline cycle. The refrigerant pressure was measured at various points using pressure transducers, and all pressure transducers had an accuracy of ±0.3%. The refrigerant temperature was also measured at various points using T-type thermocouples with an accuracy of ±0.2 °C. Coriolistype mass flow meters with an accuracy of ±0.1% were used to measure the refrigerant mass flow rates. Mass flow meter #1 was installed right after the condenser to measure the total mass flow rate, and a mass flow meter #2 was installed before the ejector to measure the mass flow rate through the motive nozzle. The power consumption in the compressor and the cooling capacity in the evaporators were measured using power meters with an accuracy of ±0.2%. Specifications of the measuring devices are listed in Table 4.
2.2. Test conditions The performance characteristics of the DEEC were compared with those of the baseline cycle when the operating conditions (compressor speed, operating pressures, subcooling, and superheat) were the same. The operating conditions for the DEEC and the baseline cycle are specified in Table 3. The compressor speed varied with 900, 2340, and 4200 rpm, which were respectively determined as the minimum, medium, and full capacity of the DEEC. The operating pressures in the condenser and low-temperature evaporator were determined according to the ISO 16358-1 Standard [30] for each compressor speed. At compressor speeds of 900, 2340, and 4200 rpm, the condensing/low-evaporating pressures were 2496/1465 kPa for the minimum capacity, 2763/1184 kPa for the medium capacity, and 3105/1027 kPa for the maximum capacity, respectively. The high-evaporating pressure in the DEEC can be determined based on the pressure at the diffuser outlet. The ER varied at 0, 0.1,
0.2, 0.4 and 0.6, while maintaining the diameter-to-length ratio of the mixing section at 10. The performance of the DEEC at an ER of 0 was measured to obtain a possible maximum value, even though this is not practical owing to the absence of the secondary flow. The mixing section diameter was varied at 4, 5, 7, and 9 mm. In addition, all experiments in the DEEC were conducted at the optimum refrigerant charge of 3.0 kg, which was determined based on preliminary tests to obtain the maximum COP for the standard testing conditions. The optimum refrigerant charge of the baseline cycle was also determined to be 2.0 kg based on preliminary tests.
2.3. Estimation parameters As given in Eq. (1), the ER was defined by the ratio of the entrained mass flow rate through the secondary nozzle to the mass flow rate through the motive nozzle. In addition, the pressure lifting ratio (PLR) was defined by Eq. (2). The pressure at the compressor inlet was the same as that at the ejector outlet. As the refrigerant passed through the ejector, the pressure lifting effect occurred due to pressure recovery in the diffuser, which can provide an energy saving potential in the compressor work. In other words, the work recovery effect of the ejector was estimated using the PLR in the DEEC.
ER msec
mmot
(1)
PLR
Pejector.out Psec 100 (%) Psec
(2)
As given in Eq. (3), COP in the DEEC was evaluated by the ratio of the total cooling capacity in the low and high temperature evaporators to the power consumption. In addition, as given in Eq. (4), COPi was defined by the COP ratio of the DEEC to the baseline cycle.
COP
Qlowtemp Qhightemp
COPi
(3)
Wcompressor
COPejector cycle COPbaseline cycle
(4)
The CSPF is a commonly used energy factor to estimate the overall performance of an airconditioner in the entire outdoor temperature range. The CSPF was estimated with the bin weather data based on the ISO 16358-1 [30]. The outdoor temperature (TA) ranged from 21 °C (TAmin) to 35 °C (TAmax) with 1 °C increments. As given in Eq. (5), the CSPF of the DEEC was defined as the ratio of the total cooling capacity in the cooling season ( the total power consumption in the cooling season (
P
C
Q
C
) to
). The CSPF for each outdoor
temperature, CSPFbin, was calculated using Eq. (6).
CSPF =
QC PC
CSPFbin =
(5)
Qc (TAj )
(6)
/
Pc (TAj ) PLF (TAj )
n
Q C X (TAj ) QC (TAj ) n j
(7)
j 1
n
X (TAj ) Pc (TAj ) n j
j 1
PLF (TAj )
PC
(8)
The cooling capacity, Qc (TAj ) , and power consumption at each outdoor temperature,
Pc (TAj ) , were obtained from a relationship between the measured data and building load calculated with the bin data (TA). Note that X(TAj) is the ratio of the building load to the
cooling capacity Based on each sensor’s accuracy, the estimated uncertainties [31] for the COP were ±2.6%, ±2.5%, and ±3.6% for 900, 2340, and 4200 rpm, respectively. In addition, the estimated uncertainty for the CSPF was within ±5.1%.
3. Results and discussion 3.1. Effects of mixing section geometry Nakagawa et al. [17] and Hu et al. [32] reported that the mixing section length and diameter had a significant influence on the performance of a standard two-phase ejector cycle. However, the effects of the mixing section geometry on the DEEC have not yet been studied. Figs. 4 and 5 show the PLR and COP of the DEEC according to the mixing section diameter for each compressor speed. At a compressor speed of 900 rpm, the PLR and COP for all ERs decreased constantly as the mixing section diameter increased. Therefore, at a lower compressor speed, a higher PLR and COP were obtained with a smaller mixing section diameter owing to less frictional pressure loss at the mixing section with a relatively lower mass flow rate through the mixing section [33,34]. At a compressor speed of 2340 rpm, the maximum PLR and COP for all ERs were observed at a mixing section diameter of 5 mm. As the mixing section diameter increased up to 5 mm, the reduced frictional pressure loss at the mixing section led to an increase in the PLR and COP due to the enlarged cross-sectional area of the mixing section. Note that the frictional pressure loss at the mixing section occurred owing to both the collision of the entrained flow with the inner wall of the ejector body before entering the mixing section and the throttling effect of the mixed two-phase flow along the mixing section. However, as the mixing section diameter increased beyond 5 mm, the PLR and COP decreased rapidly due to a decrease in the mixing efficiency with an insufficient mass flow rate through the ejector [33,34]. At a compressor speed of 4200 rpm,
the PLR and COP at ERs of 0.1 and 0.2 peaked at the mixing section diameter of 5 mm. However, at an ER of 0.6, the flow resistance became more substantial at a smaller mixing section diameter due to the increase in the secondary flow rate. Therefore, the optimum COP and PLR was observed at the mixing section diameter of 7 mm. As mentioned in Figs. 4 and 5, the optimum mixing section diameter of the ejector increased with an increase in compressor speed due to the increased mass flow rate through the mixing section. Particularly, at the high compressor speed, the optimum mixing section diameter increased from 5 mm to 7 mm with an increase in ER from 0.2 to 0.6. However, it is difficult to adjust the mixing section diameter during the operation of an air-conditioner under various cooling load conditions. Therefore, it may be reasonable to determine the optimum mixing section diameter based on the CSPF of the DEEC. Fig. 6 shows the CSPF of the DEEC according to the mixing section diameter at various ERs. For all ERs, the CSPF peaked at the mixing section diameter of 5 mm. At an ER of 0.1, the maximum CSPF was 6.81 at a mixing section diameter of 5 mm. Therefore, the optimum mixing section diameter can be determined to be 5 mm to ensure stability over the wide range of operating conditions.
3.2. Effects of entrainment ratio Fig. 7 shows the PLR, COP, COPi, compressor work, and total cooling capacity of the DEEC at a compressor speed of 4200 rpm according to the ER. As shown in Fig. 7(a), the PLR of the DEEC decreased with an increase in ER due to the reduction in energy conversion from pressure to kinetic energy with a decrease in mass flow rate through the motive nozzle. As the ER increased, the pressure at the compressor inlet decreased, resulting in a lower total mass flow rate through the compressor, even though the entrained mass flow rate increased. Furthermore, due to a large decrease in the total mass flow rate, the total cooling capacity of the DEEC decreased with an increase in ER. Specifically, as the ER increased, the cooling
capacity of evaporator #2 decreased significantly even though that of evaporator #1 increased slightly. In addition, the compression efficiency decreased with an increase in ER due to the increase in the compression ratio. As shown in Fig. 7(b), the COP and COPi of the DEEC decreased with an increase in ER. The COPi indicates the COP ratio between the DEEC and the baseline cycle. The COP of the baseline cycle at a compressor speed of 4200 rpm was 2.5. As shown in Fig. 7(c), the cooling capacity and compressor work decreased with an increase in ER due to the decreased total mass flow rate in the DEEC. However, the decrease in the cooling capacity was larger than that in the compressor work. As the ER increased beyond 0.4, the COP of the DEEC became lower than that of the baseline cycle. Therefore, the DEEC needs to be operated at ERs below 0.4 to improve the system performance over the baseline cycle. In addition, at a compressor speed of 4200 rpm with an optimized mixing section diameter of 5 mm, the DEEC showed an average COP enhancement of 12.3% over the baseline cycle for ERs between 0 and 0.4. Fig. 7(d) shows P-h diagrams of the DEEC and the baseline cycle at a compressor speed of 4200 rpm to provide clear overview of the comparison.
3.3. Effects of compressor speed Fig. 8 shows the PLR and total mass flow rate according to the compressor speed. For ERs of 0.1 and 0.2, the PLR increased with an increase in compressor speed, which was resulted from the increase in the mass flow rate through the motive nozzle with effective pressure recovery. However, for the ER of 0.6, the PLR rather decreased as the compressor speed increased due to extreme flow separation forwarding to the secondary line. For all ERs, the total mass flow rate increased with an increase in compressor speed. At ERs of 0.1 and 0.2, the DEEC showed a higher total mass flow rate than that of the baseline cycle due to the increased PLR. However, at an ER of 0.6, the DEEC showed a lower total mass flow rate
than the baseline cycle due to the decreased PLR. These trends became more substantial as the compressor speed increased beyond 2340 rpm. Fig. 9 shows the COP and COPi of the DEEC with respect to the compressor speed. As the compressor speed increased, the cooling capacity of the DEEC increased due to the increase in the total mass flow rate. However, at the same time, the power consumption increased with an increase in compressor speed due to the increased total mass flow rate and compression ratio. For all ERs, the COP of the DEEC decreased with an increase in compressor speed because the increase in the power consumption was larger than the increase in the cooling capacity. For ERs of 0.1 and 0.2, the COPi increased with an increase in compressor speed. At an ER of 0.1, the DEEC at compressor speeds of 900 and 4200 rpm showed performance improvements of 6% and 14%, respectively, over the baseline cycle. However, at an ER of 0.6, the COPi decreased with an increase in compressor speed due to the decrease in the PLR. In addition, the COPi at an ER of 0.6 was lower than 1.0 for all compressor speeds. Therefore, the effectiveness of the DEEC can be improved when the compressor operates at higher speeds with ERs below 0.4. Note that all experiments for the DEEC were conducted with specified condensing and evaporating pressures which can be matched with those of the baseline cycle.
3.4. Ejector optimization based on CSPF Fig. 10 shows the COP and CSPF of the DEEC and the baseline cycle at various compressor speeds. The CSPF of the DEEC was estimated by aggregating COPs for all experimented compressor speeds. The CSPF of the baseline cycle was 6.39, which is represented as a dotted line. The CSPF of the DEEC decreased with an increase in ER because the COP decreased according to the ER at all compressor speeds. The CSPF of the DEEC became lower than that of the baseline cycle as the ER increased beyond 0.3.
Therefore, the DEEC needs to be operated for ERs below 0.3 to attain a better CSPF over the baseline cycle. In addition, the improvement in CSPF of the DEEC over the baseline cycle was 6.3% at an ER of 0.1. Fig. 11 shows the CSPFbin of the DEEC according to the outdoor temperature. The time weighting factor indicates the ratio between the bin hours for each outdoor temperature to the total bin hours. The operation mode was classified based on the outdoor temperature, with Mode 1 for 21 °C to 24 °C, Mode 2 for 25 °C to 30 °C, and Mode 3 for 31°C to 35 °C. The compressor speed increased with an increase in cooling load to obtain a higher cooling capacity, resulting in a lower COP. For Mode 1, the compressor speed was fixed at 900 rpm with on-off operation because the cooling load was lower than the minimum cooling capacity. For Mode 2, the compressor speed varied from 900 to 2340 rpm, and for Mode 3, the compressor speed varied from 2340 to 4200 rpm according to the cooling load. The effects of the mixing section diameter on the CSPFbin of the DEEC were analyzed to determine the optimum mixing section diameter of the ejector. As shown in Fig. 11, for Mode 1 and early Mode 2, the CSPFbin was the highest at a mixing section diameter of 4 mm due to a higher mixing efficiency with a lower compressor speed. For late Mode 2 and Mode 3, the optimum mixing section diameter increased to 5 mm due to the increased mass flow rate according to the compressor speed. For Mode 3, the optimum mixing section diameter was observed at 7 mm due to the increased cooling load. As shown in Fig. 11, since the time weighting factor was relatively high in Modes 1 and 2, it is important to design an ejector to obtain a high efficiency in these modes. Considering the time weighting factor according to the outdoor temperature, the optimum mixing section diameter for the ejector was determined to be 5 mm to obtain a more stable improvement in the performance of the DEEC. Fig. 12 shows the CSPFbin of the DEEC according to the outdoor temperature at the optimum mixing section diameter of 5 mm. The baseline cycle was optimized at an outdoor
temperature of 31 °C. The CSPFbin.i was defined as the ratio of the CSPFbin of the DEEC over that of the baseline cycle. For Mode 1, the CSPFbin of the DEEC was constantly higher than that of the baseline cycle, showing an increase in the CSPFbin from 6.2% to 6.5%. For Mode 2, both the CSPFbin of the DEEC and the baseline cycle decreased with an increase in outdoor temperature. However, the DEEC showed a higher CSPFbin than that of the baseline cycle over the entire outdoor temperature range due to the larger pressure recovery potential with higher compressor speeds, showing the CSPFbin.i from 3.6% to 7.8%. For Mode 3, the CSPFbin of the DEEC at 31 °C was lower than that of the baseline cycle because the baseline cycle was optimized at this outdoor temperature. However, as the outdoor temperature increased toward 35 °C, the CSPFbin became significantly higher than that of the baseline cycle, showing an improvement of 40.0% in the CSPFbin.
4. Conclusions This study measured and analyzed the performance characteristics of an R410A airconditioner adopting the DEEC with respect to the operating conditions and ejector geometries, including compressor speed, ER, outdoor temperature, mixing section diameter and length. In addition, the optimum ejector geometries were determined to achieve the maximum CSPF of the DEEC according to the given cooling load. As the ER increased, the COP of the DEEC decreased due to the decrease in the PLR. The effectiveness of the DEEC was improved when the compressor operated at higher speeds with ERs below 0.4. The optimum mixing section diameter was determined to be 5 mm, showing the maximum CSPF of 6.81 at an ER of 0.1. The same optimum mixing section diameter was also obtained by analyzing the CSPFbin of the DEEC, achieving the most stable improvement in the DEEC performance. In addition, the maximum allowable limit of the ER was 0.3 because the CSPF of the DEEC became lower than that of the baseline cycle as the ER increased beyond 0.3.
For ERs between 0 and 0.4 at 4200 rpm, the DEEC with the optimized mixing section diameter showed an average improvement in COP of 12.3% over the baseline cycle, and the CSPF enhancement of the DEEC over the baseline cycle was 6.3% at an ER of 0.1.
Acknowledgements This work was supported by the Human Resources Program in Energy Technology (No. 20144010200770) and the Energy Technology Development Program (No. 20142010102660) of the Korea Institute of Energy Technology Evaluation and Planning (KETEP) grant financial resource from the Ministry of Trade, Industry & Energy, Republic of Korea.
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[25] F. Liu, E.A. Groll, D. Li, Investigation on performance of variable geometry ejectors for CO2 refrigeration cycles, Energy J. 45 (2012) 829–839. [26] D. Li, E.A. Groll, Transcritical CO2 refrigeration cycle with ejector-expansion device, Int. J. Refrig. 28 (2005) 766–773. [27] S. Elbel, P. Hrnjak, Flash gas bypass for improving the performance of transcritical R744 systems that use microchannel evaporators, Int. J. Refrig. 27 (2004) 724–735. [28] H. Takeuchi, H. Nishijima, T. Ikemoto, “World’s first high efficiency refrigeration cycle with two-phase ejector: “Ejector cycle”, In: 2004 SAE World Congress and Exhibition, Detroit (MI), paper 2004-01-0916. [29] G. Pottker, P. Hrnjak, Ejector in R410A vapor compression systems with experimental quantification of two major mechanisms of performance improvement: Work recovery and liquid feeding, Int. J. Refrig. 50 (2015) 184–192. [30] ISO 16358-1, Air-cooled air conditioners and air-to-air heat pumps-Testing and calculating methods for seasonal performance factors-Part 1: Cooling seasonal performance factor, International Standards Organization, 2013. [31] B.N. Taylor, C.E. Kuyatt, Guidelines for evaluating and expressing the uncertainty of NIST measurement results, Technology 1994. [32] J. Hu, J. Shi, Y. Liang, Z. Yang, J. Chen, Numerical and experimental investigation on nozzle parameters for R410A ejector air conditioning system, Int. J. Refrig. 40 (2014) 338–346. [33] N. Galanis, M. Sorin, Ejector design and performance prediction, Int. J. Therm. Sci. 104 (2016) 315–329. [34] M. Khennich, N. Galanis, M. Sorin, Effect of design conditions and irreversibilities on the dimensions of ejectors in refrigeration systems, Appl. Energy 179 (2016) 1020– 1031.
Fig. 1 Schematic and P-h diagrams of the DEEC.
Fig. 2 Schematic diagram of the experimental setup.
Fig. 3 Schematic diagram of the ejector assembly.
5.8
ER=0.1 ER=0.2 ER=0.6
(a)
5.6 5.4
4.8
SC = 3oC SH = 5oC
5.0 4.8
4.5
Compressor speed=2340 rpm
ER=0.1 ER=0.2 ER=0.6
(b)
4.0
SC = 3oC SH = 5oC
4.6
COP
5.2
COP
5.0
Compressor speed=900 rpm
4.4 4.2
Compressor speed = 4200 rpm
ER=0.1 ER=0.2 ER=0.6
(c)
SC = 3oC SH = 5oC
3.5
COP
6.0
3.0 2.5
4.6 4.4 4.2
4.0
2.0
3.8
1.5
3.6
1.0
4.0 3.8 4
5
6
7
8
9
4
5
6
7
8
9
4
5
6
7
Mixing section diameter (mm)
Fig. 4 COP according to mixing section diameter for different ERs.
8
9
ER=0.1 ER=0.2 ER=0.6
(a)
SC = 3oC SH = 5oC
10 8 6 4 2 0
12
20
Compressor speed=2340 rpm
ER=0.1 ER=0.2 ER=0.6
(b)
10
SC = 3oC SH = 5oC
8 6 4 2 0
Pressure lifting ratio (%)
12
14
Compressor speed=900 rpm
Pressure lifting ratio (%)
Pressure lifting ratio (%)
14
5
6
7
8
9
ER=0.1 ER=0.2 ER=0.6
(c)
SC = 3oC SH = 5oC
10 5 0 -5 -10 -15
-2 4
15
Compressor speed = 4200 rpm
4
5
6
7
8
9
4
5
6
7
8
Mixing section diameter (mm)
Fig. 5 Pressure lifting ratio according to mixing section diameter for different ERs.
9
7.0
SC = 3oC, SH = 5oC ER=0.1 ER=0.2 ER=0.6
6.8
CSPF
6.6 6.4 6.2 6.0 5.8 4
5
6
7
Mixing section diameter (mm)
Fig. 6 CSPF according to mixing section diameter for different ERs.
200 10 100 5
0
3.0
2.8 1.1 2.6
(b) 2.2
0 0.0
0.1
0.2
0.3
0.4
0.5
0.6
0.2
0.3
0.4
0.5
0.6
0.7
Entrainment ratio
o
SC = 3 C, SH = 5 C, Dm = 5.0 mm Compressor speed = 4200 rpm Cooling capacity Compressor power
11
0.1
3.65
3.55 10 3.45 9
Compressor power (kW)
Cooling capacity (kW)
o
0.9 0.0
0.7
Entrainment ratio
12
1.0
2.4
(a) -100
1.3 o SC = 3 C, SH = 5oC, Dm = 5.0 mm Compressor speed = 4200 rpm COP COPi 1.2
COPi
300
3.2
COP
20 o SC = 3oC, SH= 5 C, Dm = 5.0 mm Compressor speed = 4200 rpm PLR Total flow 15 Entrained flow
Pressure lifting ratio (%)
-1
Mass flow rate (kgh )
400
(c) 8
3.35 0.0
0.1
0.2
0.3
0.4
0.5
0.6
0.7
Entrainment ratio
Fig. 7 (a) Mass flow rate and pressure lifting ratio according to ER, (b) COP and COPi , (c) Cooling capacity and compressor work at a compressor speed of 4200 rpm, (d) P-h diagrams of the DEEC and baseline cycle at a compressor speed of 4200 rpm and ER of 0.2.
12
o
o
Pressure lifting ratio (%)
SC = 3 C, SH = 5 C, Dm = 5.0 mm 10
ER=0.1 ER=0.2 ER=0.6
8 6 4 2
(a)
0 1000
2000
3000
4000
5000
Compressor speed (rpm)
280
o
o
240
-1
Mass flow rate (kgh )
SC = 3 C, SH = 5 C
200
ER=0.1 ER=0.2 ER=0.6 Baseline cycle
160 120 80
(b) 40 1000
2000
3000
4000
5000
Compressor speed (rpm)
Fig. 8 Pressure lifting ratio and total mass flow rate according to compressor speed for different ERs.
1.4 COP at ER=0.1 COP at ER=0.2 COP at ER=0.6 COPi at ER=0.1
5
COPi at ER=0.2
COP
COPi at ER=0.6
4
1.3 1.2 1.1 1.0
3 0.9 o
2
o
SC = 3 C, SH = 5 C, Dm = 5.0 mm 1000
2000
3000
4000
0.8 5000
Compressor speed (rpm)
Fig. 9 COP and COPi according to compressor speed.
COPi
6
7.2 6
COP
5
7.0 6.8 6.6
4
6.4
CSPF
COP at 900 rpm COP at 2340 rpm COP at 4200 rpm CSPF, Ejector cycle CSPF, Baseline cycle
6.2 3 6.0 o
2
o
SC = 3 C, SH = 5 C, Dm = 5.0 mm 0.0
0.1
0.2
0.3
0.4
5.8 0.5
0.6
0.7
Entrainment ratio
Fig. 10 COP and CSPF according to ER for different compressor speeds.
0.14 Entrainment ratio = 0.2 Dm = 4.0 Dm = 5.0 Dm = 7.0 Time weighting factor
CSPFbin
10
0.12 0.10 0.08
8
0.06 6 0.04 4
Time weighting factor
12
0.02
2
0.00 20
22
24
26
28
30
32
34
36
Outdoor temperature (oC)
Fig. 11 CSPFbin and bin time weighting factor of the ejector cycle according to outdoor temperature for different mixing section diameters.
12
Baseline cycle Ejector cycle Ejector effect
7.0 14 12
Baseline cycle Ejector cycle Ejector effect
6.8 10
CSPFbin
10 6.6
8
12 14 10 12 8
Baseline cycle Ejector cycle Ejector effect
30 8
6
6
6.4
6 4
6.2 2 0
6.0 21 22 23 24
20 6
4
4
50 40
10
8
60
10 4
0
2
2
2
-10
0
0
0
-20
25 26 27 28 29 30
CSPFbin.i (%)
14
31 32 33 34 35
Outdoor temperature (oC)
Fig. 12 CSPFbin and CSPFbin.i of the ejector cycle and the baseline cycle according to outdoor temperature.
Figure captions
Fig. 1
Schematic and P-h diagrams of the DEEC.
Fig. 2
Schematic diagram of the experimental setup.
Fig. 3
Schematic diagram of the ejector assembly.
Fig. 4
COP according to mixing section diameter for different ERs.
Fig. 5
Pressure lifting ratio according to mixing section diameter for different ERs.
Fig. 6
CSPF according to mixing section diameter for different ERs.
Fig. 7
(a) Mass flow rate and pressure lifting ratio according to ER, (b) COP and COPi , (c) Cooling capacity and compressor work at a compressor speed of 4200 rpm, (d) P-h diagrams of the DEEC and baseline cycle at a compressor speed of 4200 rpm and ER of 0.2.
Fig. 8
Pressure lifting ratio and total mass flow rate according to compressor speed for different ERs.
Fig. 9
COP and COPi according to compressor speed.
Fig. 10 COP and CSPF according to ER for different compressor speeds. Fig. 11 CSPFbin and bin time weighting factor of the ejector cycle according to outdoor temperature for different mixing section diameters. Fig. 12 CSPFbin and CSPFbin.i of the ejector cycle and the baseline cycle according to outdoor temperature.
Table 1 Specifications of the tested ejector Item
Ejector
Part
Value
Length of the converging part of the motive nozzle
12 mm
Length of the diverging part of the motive nozzle
24 mm
Length between the motive nozzle and the inlet of the mixing section
0 mm
Diameter of the nozzle throat
2.4 mm
Diameter of the nozzle outlet
2.7 mm
Angle of the diverging part of the motive nozzle Angle of the diffuser
1.02° 5°
Table 2 Specifications of the DEEC and the baseline cycle. Item Compressor
Condenser Evaporator #1
Specifications Type Working fluid
DEEC Rotary R410A
Baseline cycle Rotary R410A
Unit
Charge amount Type Capacity Type
3 Plate heat exchanger 12 Cartridge heater
2 Plate heat exchanger 12 Cartridge heater
kg
Capacity
12
12
KW
Evaporator #2 Type Capacity Orifice Needle valve diameter
Cartridge heater 12 6.4
None 6.4
KW
KW mm
Table 3 Test conditions and ejector geometries. Parameter
Value
Compressor speed (rpm)
900, 2340, 4200
Entrainment ratio, ER
0, 0.1, 0.2, 0.4, 0.6
Condensing pressure (kPa)
3105
2763
2496
Low-evaporating pressure (kPa)
1027
1184
1465
Condensing temperature (°C)
51
46
41
Low-evaporating temperature (°C)
8
13
20
Subcooling (°C)
3
Superheat (°C) Mixing diameter (mm)
5 4
5
7
9
Mixing diameter length (mm)
40
50
70
90
Table 4 Specifications of the measuring devices. Range
Accuracy
Number of Measuring points
0–5 MPa
±0.3%.
10
Thermocouple (T-type)
-200–200 °C
±0.2 °C
10
Mass flow meter
0–300 kg h-1
±0.1%
2
0–300 V, 0–20 A
±0.2%
3
Device Pressure transducer
Power meter
Highlights
The performance of an R410A air-conditioner adopting a DEEC.
The effects of operating parameters on the COP and CSPF.
Optimized mixing section diameter of the ejector.
The maximum limit of the entrainment ratio (ER).
S1359-4311(17)33447-6 https://doi.org/10.1016/j.applthermaleng.2017.12.012 ATE 11533
To appear in:
Applied Thermal Engineering
Received Date: Revised Date: Accepted Date:
20 May 2017 19 October 2017 2 December 2017
Please cite this article as: S. Kim, Y. Jeon, H. Joon Chung, Y. Kim, Performance optimization of an R410A airconditioner with a dual evaporator ejector cycle based on cooling seasonal performance factor, Applied Thermal Engineering (2017), doi: https://doi.org/10.1016/j.applthermaleng.2017.12.012
This is a PDF file of an unedited manuscript that has been accepted for publication. As a service to our customers we are providing this early version of the manuscript. The manuscript will undergo copyediting, typesetting, and review of the resulting proof before it is published in its final form. Please note that during the production process errors may be discovered which could affect the content, and all legal disclaimers that apply to the journal pertain.
Performance optimization of an R410A air-conditioner with a dual evaporator ejector cycle based on cooling seasonal performance factor
Sunjae Kim, Yongseok Jeon, Hyun Joon Chung, Yongchan Kim* Department of Mechanical Engineering, Korea University Anam-Dong, Sungbuk-Ku, Seoul 136-713, Republic of Korea
ABSTRACT Even though a dual evaporator ejector cycle (DEEC) offers several advantages over a standard two-phase ejector cycle, few experimental investigations of the performance of the DEEC are available in the literature. This study presents the performance characteristics of an R410A air-conditioner adopted with a DEEC under various operating conditions and ejector geometries. The COP of the DEEC decreased with an increase in entrainment ratio (ER) due to the decrease in pressure lifting ratio. For the optimum ER, the effectiveness of the DEEC increased with an increase in compressor speed with a larger total mass flow rate. The optimum mixing section diameter was determined to be 5 mm based on the cooling seasonal performance factor (CSPF) and CSPFbin of the DEEC. The maximum allowable limit for the ER was also suggested to be 0.3. In addition, the CSPF of the DEEC was 6.3% higher than that of the baseline cycle at an ER of 0.1.
Keywords: Dual evaporator ejector cycle, Ejector, COP, CSPF
* Corresponding author. Tel.: +82-2-3290-3366; Fax: +82-2-921-5439 E-mail address: [email protected]
Nomenclature COP
coefficient of performance
COPi
COP ratio of the ejector cycle to the baseline cycle
CSPF
cooling seasonal performance factor
CSPFbin
cooling seasonal performance factor for a specific bin temperature
D
diameter (mm)
DEEC
dual evaporator ejector cycle
ER
entrainment ratio
L
length (mm)
MFR
mass flow meter mass flow rate (kg h-1)
n
number of temperature bins
nj
bin hours
P
pressure (kPa)
PLF
part load factor
PLR
pressure lifting ratio
SC
subcooling (°C)
SH
superheat (°C)
TA
bin outdoor temperature (°C)
Q
cooling capacity (W h) cooling capacity rate (W)
X
ratio of building load to cooling capacity at outdoor temperature
Subscripts c
cooling
high-temp
high temperature evaporator
low-temp
low temperature evaporator
m
mixing section
mot
motive
out
outlet
sec
secondary flow
1. Introduction Even though air-conditioning systems have contributed to human well-being in many aspects, they have a considerable impact in terms of various environmental and energy problems, such as ozone layer depletion and global warming [1]. In the air-conditioning industry, it is crucial to develop environmentally friendly and energy-efficient airconditioners [2–5]. Renewable energies have been introduced by many researchers to diversify the energy source of air-conditioning systems [6,7]. In addition, refrigerant injection technology has been studied to increase the COP and reliability of a heat pump system in severe outdoor conditions [8–10]. Furthermore, the expansion valve has the potential for further advances because it suffers a significant energy loss during the isenthalpic expansion process [11–13]. Therefore, an ejector has been introduced to retrieve the energy loss during the expansion process in conventional expansion devices [14,15]. Previous ejector studies [16–19] have focused on a standard two-phase ejector cycle that consists of a compressor, an ejector, a condenser, and a liquid-vapor separator. In contrast, Oshitani et al. [20] proposed a dual evaporator ejector cycle (DEEC) as a climate control system of the cabin in automobiles. They found that the DEEC had better operation stability and adaptability with low-pressure refrigerants compared to the standard two-phase ejector
cycle. As shown in Fig. 1, the DEEC consists of a compressor, an ejector, a condenser, and two evaporators each with different evaporating temperature. In the DEEC, the refrigerant at the condenser outlet is divided into two separate streams. One of them flows directly into the ejector passing through the motive nozzle, and the other stream (secondary flow loop) is isenthalpically expanded through the needle valve and is then vaporized in the low temperature evaporator. The refrigerant at the secondary loop is entrained into the secondary block and is then mixed with the motive flow in the mixing section. The mixed two-phase flow recovers pressure energy while passing through the diffuser section and then passes through the high temperature evaporator to superheat the refrigerant. Lawrence and Elbel [21,22] observed that a DEEC showed maximum COP improvements of 12% and 8% using R1234yf and R134a, respectively. In addition, Wang et al. [23] compared the performance of alternate ejector cycles, including the DEEC. A transcritical cycle using R744 (CO2) usually operates with extremely high condensing pressure causing not only large compression ratio but also large expansion loss. A two-phase ejector cycle has been adopted to overcome large expansion loss of the transcritical cycle using R744. Kornhauser [24] presented a one-dimensional model for an R12 refrigeration system with a two-phase ejector. Kornhauser’s model was utilized to study the COP improvement of R744 ejector cycles [25–27]. Takeuchi et al. [28] reported a COP improvement of 20% using an R744 ejector cycle for an automobile air-conditioning system. Elbel and Hrnjak [16] experimentally reported an improvement in COP and cooling capacity by up to 7% and 8%, respectively. More recently, Nakagawa et al. [17] observed a COP improvement of 26% by optimizing the mixing section length of an ejector in an R744 refrigeration system. In addition, the ejector cycles with low-pressure refrigerants can still offer noticeable COP improvements even though those have received relatively less attention due to a lower work recovery potential. Harrell and Kornhouser [18] reported COP
improvements of 4% to 8% in an R134a ejector cycle against a conventional vapor compression cycle. Disawas and Wongwises [19] presented an increase in COP and cooling capacity in a two-phase ejector cycle over a conventional R134a vapor compression cycle. Most recently, Pottker and Hrnjak [29] observed a COP improvement of 8.4% in an ejector cycle with a liquid fed evaporator system due to the work recovery effect. Even though the DEEC offers several advantages over the standard two-phase ejector cycle [21,22], it has been mainly adopted in refrigerators with two-evaporating temperatures. The DEEC becomes more beneficial with an increase in refrigerant mass flow rate, which can be used in air-conditioners with relatively higher cooling capacity. However, few experimental investigations on the performance characteristics of a DEEC-based air-conditioner have been presented in the literature. In particular, the effects of the ejector geometries and operating conditions on the performance of the DEEC have not been comprehensively analyzed. Furthermore, the ejector optimization of the DEEC based on the cooling seasonal performance factor (CSPF) is essential in order to cope with various cooling loads. However, the optimization of the DEEC based on the CSPF has been rarely presented in the literature. Thus, this study investigates the performance characteristics of a DEEC-based R410A airconditioner according to the operating conditions and ejector geometries. The performance of the DEEC was measured and optimized by varying the compressor speed, entrainment ratio (ER), outdoor temperature, and mixing section diameter and length. The COP and cooling capacity of the DEEC were compared to those of the baseline cycle using a conventional expansion device. In addition, the CSPF of the DEEC was compared to that of the baseline cycle in order to optimize the ejector geometries under various cooling loads.
2. Experimental setup and test conditions 2.1. Experimental setup
Fig. 2 shows a schematic diagram of the experimental setup used to measure the performance of the DEEC. The testing setup for the DEEC consisted of a rotary compressor with an inverter driver, a condenser, two-evaporators, an ejector, and a needle valve. A plate type heat exchanger was used as the condenser to control the condensing pressure and heat transfer rate, and a water chiller was connected to the condenser in order to remove the heating load. The needle valve was utilized as an expansion device in the secondary loop of the DEEC. The cooling load into evaporator #1 was controlled using cartridge heaters, adjusting the evaporating pressure and heat transfer rate. The cooling load into evaporator #2, which had a higher evaporating temperature than the evaporator #1, was also controlled using cartridge heaters. The DEEC can be used to provide the required cooling capacity for each air-conditioning space with different temperatures. In addition, two evaporators can be combined serially as one unit. In that case, the temperature gliding in the combined evaporator can be matched closely with that in the air stream, resulting in a higher heat transfer effectiveness. As shown in Fig. 3, the ejector consisted of a motive nozzle, a secondary block, a mixing section, and a diffuser. Each ejector component was manufactured separately, and then all pieces were assembled together. The ejector geometries varied by diversifying the selection of the modules, and the throat diameter of the motive nozzle was controlled by adjusting the ejector needle. The ejector needle and the needle valve on the secondary line were manually controlled to adjust the ER. To increase the ER, the ejector needle was moved forward to reduce the motive flow rate, while the needle valve was partially opened to increase the secondary flow. Geometrical specifications of the tested ejector are listed in Table 1. In addition, the DEEC performance was compared by constructing the baseline cycle using a conventional air-conditioner consisting of a rotary compressor with an inverter driver, a condenser, an evaporator, and a needle valve as an expansion device. The baseline cycle had a
single evaporating temperature to make a performance comparison of the novel cycle with the conventional air-conditioner cycle, as that proposed by Lawrence and Elbel [21,22]. Table 2 summarizes the component specifications of the DEEC and the baseline cycle. The refrigerant pressure was measured at various points using pressure transducers, and all pressure transducers had an accuracy of ±0.3%. The refrigerant temperature was also measured at various points using T-type thermocouples with an accuracy of ±0.2 °C. Coriolistype mass flow meters with an accuracy of ±0.1% were used to measure the refrigerant mass flow rates. Mass flow meter #1 was installed right after the condenser to measure the total mass flow rate, and a mass flow meter #2 was installed before the ejector to measure the mass flow rate through the motive nozzle. The power consumption in the compressor and the cooling capacity in the evaporators were measured using power meters with an accuracy of ±0.2%. Specifications of the measuring devices are listed in Table 4.
2.2. Test conditions The performance characteristics of the DEEC were compared with those of the baseline cycle when the operating conditions (compressor speed, operating pressures, subcooling, and superheat) were the same. The operating conditions for the DEEC and the baseline cycle are specified in Table 3. The compressor speed varied with 900, 2340, and 4200 rpm, which were respectively determined as the minimum, medium, and full capacity of the DEEC. The operating pressures in the condenser and low-temperature evaporator were determined according to the ISO 16358-1 Standard [30] for each compressor speed. At compressor speeds of 900, 2340, and 4200 rpm, the condensing/low-evaporating pressures were 2496/1465 kPa for the minimum capacity, 2763/1184 kPa for the medium capacity, and 3105/1027 kPa for the maximum capacity, respectively. The high-evaporating pressure in the DEEC can be determined based on the pressure at the diffuser outlet. The ER varied at 0, 0.1,
0.2, 0.4 and 0.6, while maintaining the diameter-to-length ratio of the mixing section at 10. The performance of the DEEC at an ER of 0 was measured to obtain a possible maximum value, even though this is not practical owing to the absence of the secondary flow. The mixing section diameter was varied at 4, 5, 7, and 9 mm. In addition, all experiments in the DEEC were conducted at the optimum refrigerant charge of 3.0 kg, which was determined based on preliminary tests to obtain the maximum COP for the standard testing conditions. The optimum refrigerant charge of the baseline cycle was also determined to be 2.0 kg based on preliminary tests.
2.3. Estimation parameters As given in Eq. (1), the ER was defined by the ratio of the entrained mass flow rate through the secondary nozzle to the mass flow rate through the motive nozzle. In addition, the pressure lifting ratio (PLR) was defined by Eq. (2). The pressure at the compressor inlet was the same as that at the ejector outlet. As the refrigerant passed through the ejector, the pressure lifting effect occurred due to pressure recovery in the diffuser, which can provide an energy saving potential in the compressor work. In other words, the work recovery effect of the ejector was estimated using the PLR in the DEEC.
ER msec
mmot
(1)
PLR
Pejector.out Psec 100 (%) Psec
(2)
As given in Eq. (3), COP in the DEEC was evaluated by the ratio of the total cooling capacity in the low and high temperature evaporators to the power consumption. In addition, as given in Eq. (4), COPi was defined by the COP ratio of the DEEC to the baseline cycle.
COP
Qlowtemp Qhightemp
COPi
(3)
Wcompressor
COPejector cycle COPbaseline cycle
(4)
The CSPF is a commonly used energy factor to estimate the overall performance of an airconditioner in the entire outdoor temperature range. The CSPF was estimated with the bin weather data based on the ISO 16358-1 [30]. The outdoor temperature (TA) ranged from 21 °C (TAmin) to 35 °C (TAmax) with 1 °C increments. As given in Eq. (5), the CSPF of the DEEC was defined as the ratio of the total cooling capacity in the cooling season ( the total power consumption in the cooling season (
P
C
Q
C
) to
). The CSPF for each outdoor
temperature, CSPFbin, was calculated using Eq. (6).
CSPF =
QC PC
CSPFbin =
(5)
Qc (TAj )
(6)
/
Pc (TAj ) PLF (TAj )
n
Q C X (TAj ) QC (TAj ) n j
(7)
j 1
n
X (TAj ) Pc (TAj ) n j
j 1
PLF (TAj )
PC
(8)
The cooling capacity, Qc (TAj ) , and power consumption at each outdoor temperature,
Pc (TAj ) , were obtained from a relationship between the measured data and building load calculated with the bin data (TA). Note that X(TAj) is the ratio of the building load to the
cooling capacity Based on each sensor’s accuracy, the estimated uncertainties [31] for the COP were ±2.6%, ±2.5%, and ±3.6% for 900, 2340, and 4200 rpm, respectively. In addition, the estimated uncertainty for the CSPF was within ±5.1%.
3. Results and discussion 3.1. Effects of mixing section geometry Nakagawa et al. [17] and Hu et al. [32] reported that the mixing section length and diameter had a significant influence on the performance of a standard two-phase ejector cycle. However, the effects of the mixing section geometry on the DEEC have not yet been studied. Figs. 4 and 5 show the PLR and COP of the DEEC according to the mixing section diameter for each compressor speed. At a compressor speed of 900 rpm, the PLR and COP for all ERs decreased constantly as the mixing section diameter increased. Therefore, at a lower compressor speed, a higher PLR and COP were obtained with a smaller mixing section diameter owing to less frictional pressure loss at the mixing section with a relatively lower mass flow rate through the mixing section [33,34]. At a compressor speed of 2340 rpm, the maximum PLR and COP for all ERs were observed at a mixing section diameter of 5 mm. As the mixing section diameter increased up to 5 mm, the reduced frictional pressure loss at the mixing section led to an increase in the PLR and COP due to the enlarged cross-sectional area of the mixing section. Note that the frictional pressure loss at the mixing section occurred owing to both the collision of the entrained flow with the inner wall of the ejector body before entering the mixing section and the throttling effect of the mixed two-phase flow along the mixing section. However, as the mixing section diameter increased beyond 5 mm, the PLR and COP decreased rapidly due to a decrease in the mixing efficiency with an insufficient mass flow rate through the ejector [33,34]. At a compressor speed of 4200 rpm,
the PLR and COP at ERs of 0.1 and 0.2 peaked at the mixing section diameter of 5 mm. However, at an ER of 0.6, the flow resistance became more substantial at a smaller mixing section diameter due to the increase in the secondary flow rate. Therefore, the optimum COP and PLR was observed at the mixing section diameter of 7 mm. As mentioned in Figs. 4 and 5, the optimum mixing section diameter of the ejector increased with an increase in compressor speed due to the increased mass flow rate through the mixing section. Particularly, at the high compressor speed, the optimum mixing section diameter increased from 5 mm to 7 mm with an increase in ER from 0.2 to 0.6. However, it is difficult to adjust the mixing section diameter during the operation of an air-conditioner under various cooling load conditions. Therefore, it may be reasonable to determine the optimum mixing section diameter based on the CSPF of the DEEC. Fig. 6 shows the CSPF of the DEEC according to the mixing section diameter at various ERs. For all ERs, the CSPF peaked at the mixing section diameter of 5 mm. At an ER of 0.1, the maximum CSPF was 6.81 at a mixing section diameter of 5 mm. Therefore, the optimum mixing section diameter can be determined to be 5 mm to ensure stability over the wide range of operating conditions.
3.2. Effects of entrainment ratio Fig. 7 shows the PLR, COP, COPi, compressor work, and total cooling capacity of the DEEC at a compressor speed of 4200 rpm according to the ER. As shown in Fig. 7(a), the PLR of the DEEC decreased with an increase in ER due to the reduction in energy conversion from pressure to kinetic energy with a decrease in mass flow rate through the motive nozzle. As the ER increased, the pressure at the compressor inlet decreased, resulting in a lower total mass flow rate through the compressor, even though the entrained mass flow rate increased. Furthermore, due to a large decrease in the total mass flow rate, the total cooling capacity of the DEEC decreased with an increase in ER. Specifically, as the ER increased, the cooling
capacity of evaporator #2 decreased significantly even though that of evaporator #1 increased slightly. In addition, the compression efficiency decreased with an increase in ER due to the increase in the compression ratio. As shown in Fig. 7(b), the COP and COPi of the DEEC decreased with an increase in ER. The COPi indicates the COP ratio between the DEEC and the baseline cycle. The COP of the baseline cycle at a compressor speed of 4200 rpm was 2.5. As shown in Fig. 7(c), the cooling capacity and compressor work decreased with an increase in ER due to the decreased total mass flow rate in the DEEC. However, the decrease in the cooling capacity was larger than that in the compressor work. As the ER increased beyond 0.4, the COP of the DEEC became lower than that of the baseline cycle. Therefore, the DEEC needs to be operated at ERs below 0.4 to improve the system performance over the baseline cycle. In addition, at a compressor speed of 4200 rpm with an optimized mixing section diameter of 5 mm, the DEEC showed an average COP enhancement of 12.3% over the baseline cycle for ERs between 0 and 0.4. Fig. 7(d) shows P-h diagrams of the DEEC and the baseline cycle at a compressor speed of 4200 rpm to provide clear overview of the comparison.
3.3. Effects of compressor speed Fig. 8 shows the PLR and total mass flow rate according to the compressor speed. For ERs of 0.1 and 0.2, the PLR increased with an increase in compressor speed, which was resulted from the increase in the mass flow rate through the motive nozzle with effective pressure recovery. However, for the ER of 0.6, the PLR rather decreased as the compressor speed increased due to extreme flow separation forwarding to the secondary line. For all ERs, the total mass flow rate increased with an increase in compressor speed. At ERs of 0.1 and 0.2, the DEEC showed a higher total mass flow rate than that of the baseline cycle due to the increased PLR. However, at an ER of 0.6, the DEEC showed a lower total mass flow rate
than the baseline cycle due to the decreased PLR. These trends became more substantial as the compressor speed increased beyond 2340 rpm. Fig. 9 shows the COP and COPi of the DEEC with respect to the compressor speed. As the compressor speed increased, the cooling capacity of the DEEC increased due to the increase in the total mass flow rate. However, at the same time, the power consumption increased with an increase in compressor speed due to the increased total mass flow rate and compression ratio. For all ERs, the COP of the DEEC decreased with an increase in compressor speed because the increase in the power consumption was larger than the increase in the cooling capacity. For ERs of 0.1 and 0.2, the COPi increased with an increase in compressor speed. At an ER of 0.1, the DEEC at compressor speeds of 900 and 4200 rpm showed performance improvements of 6% and 14%, respectively, over the baseline cycle. However, at an ER of 0.6, the COPi decreased with an increase in compressor speed due to the decrease in the PLR. In addition, the COPi at an ER of 0.6 was lower than 1.0 for all compressor speeds. Therefore, the effectiveness of the DEEC can be improved when the compressor operates at higher speeds with ERs below 0.4. Note that all experiments for the DEEC were conducted with specified condensing and evaporating pressures which can be matched with those of the baseline cycle.
3.4. Ejector optimization based on CSPF Fig. 10 shows the COP and CSPF of the DEEC and the baseline cycle at various compressor speeds. The CSPF of the DEEC was estimated by aggregating COPs for all experimented compressor speeds. The CSPF of the baseline cycle was 6.39, which is represented as a dotted line. The CSPF of the DEEC decreased with an increase in ER because the COP decreased according to the ER at all compressor speeds. The CSPF of the DEEC became lower than that of the baseline cycle as the ER increased beyond 0.3.
Therefore, the DEEC needs to be operated for ERs below 0.3 to attain a better CSPF over the baseline cycle. In addition, the improvement in CSPF of the DEEC over the baseline cycle was 6.3% at an ER of 0.1. Fig. 11 shows the CSPFbin of the DEEC according to the outdoor temperature. The time weighting factor indicates the ratio between the bin hours for each outdoor temperature to the total bin hours. The operation mode was classified based on the outdoor temperature, with Mode 1 for 21 °C to 24 °C, Mode 2 for 25 °C to 30 °C, and Mode 3 for 31°C to 35 °C. The compressor speed increased with an increase in cooling load to obtain a higher cooling capacity, resulting in a lower COP. For Mode 1, the compressor speed was fixed at 900 rpm with on-off operation because the cooling load was lower than the minimum cooling capacity. For Mode 2, the compressor speed varied from 900 to 2340 rpm, and for Mode 3, the compressor speed varied from 2340 to 4200 rpm according to the cooling load. The effects of the mixing section diameter on the CSPFbin of the DEEC were analyzed to determine the optimum mixing section diameter of the ejector. As shown in Fig. 11, for Mode 1 and early Mode 2, the CSPFbin was the highest at a mixing section diameter of 4 mm due to a higher mixing efficiency with a lower compressor speed. For late Mode 2 and Mode 3, the optimum mixing section diameter increased to 5 mm due to the increased mass flow rate according to the compressor speed. For Mode 3, the optimum mixing section diameter was observed at 7 mm due to the increased cooling load. As shown in Fig. 11, since the time weighting factor was relatively high in Modes 1 and 2, it is important to design an ejector to obtain a high efficiency in these modes. Considering the time weighting factor according to the outdoor temperature, the optimum mixing section diameter for the ejector was determined to be 5 mm to obtain a more stable improvement in the performance of the DEEC. Fig. 12 shows the CSPFbin of the DEEC according to the outdoor temperature at the optimum mixing section diameter of 5 mm. The baseline cycle was optimized at an outdoor
temperature of 31 °C. The CSPFbin.i was defined as the ratio of the CSPFbin of the DEEC over that of the baseline cycle. For Mode 1, the CSPFbin of the DEEC was constantly higher than that of the baseline cycle, showing an increase in the CSPFbin from 6.2% to 6.5%. For Mode 2, both the CSPFbin of the DEEC and the baseline cycle decreased with an increase in outdoor temperature. However, the DEEC showed a higher CSPFbin than that of the baseline cycle over the entire outdoor temperature range due to the larger pressure recovery potential with higher compressor speeds, showing the CSPFbin.i from 3.6% to 7.8%. For Mode 3, the CSPFbin of the DEEC at 31 °C was lower than that of the baseline cycle because the baseline cycle was optimized at this outdoor temperature. However, as the outdoor temperature increased toward 35 °C, the CSPFbin became significantly higher than that of the baseline cycle, showing an improvement of 40.0% in the CSPFbin.
4. Conclusions This study measured and analyzed the performance characteristics of an R410A airconditioner adopting the DEEC with respect to the operating conditions and ejector geometries, including compressor speed, ER, outdoor temperature, mixing section diameter and length. In addition, the optimum ejector geometries were determined to achieve the maximum CSPF of the DEEC according to the given cooling load. As the ER increased, the COP of the DEEC decreased due to the decrease in the PLR. The effectiveness of the DEEC was improved when the compressor operated at higher speeds with ERs below 0.4. The optimum mixing section diameter was determined to be 5 mm, showing the maximum CSPF of 6.81 at an ER of 0.1. The same optimum mixing section diameter was also obtained by analyzing the CSPFbin of the DEEC, achieving the most stable improvement in the DEEC performance. In addition, the maximum allowable limit of the ER was 0.3 because the CSPF of the DEEC became lower than that of the baseline cycle as the ER increased beyond 0.3.
For ERs between 0 and 0.4 at 4200 rpm, the DEEC with the optimized mixing section diameter showed an average improvement in COP of 12.3% over the baseline cycle, and the CSPF enhancement of the DEEC over the baseline cycle was 6.3% at an ER of 0.1.
Acknowledgements This work was supported by the Human Resources Program in Energy Technology (No. 20144010200770) and the Energy Technology Development Program (No. 20142010102660) of the Korea Institute of Energy Technology Evaluation and Planning (KETEP) grant financial resource from the Ministry of Trade, Industry & Energy, Republic of Korea.
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Fig. 1 Schematic and P-h diagrams of the DEEC.
Fig. 2 Schematic diagram of the experimental setup.
Fig. 3 Schematic diagram of the ejector assembly.
5.8
ER=0.1 ER=0.2 ER=0.6
(a)
5.6 5.4
4.8
SC = 3oC SH = 5oC
5.0 4.8
4.5
Compressor speed=2340 rpm
ER=0.1 ER=0.2 ER=0.6
(b)
4.0
SC = 3oC SH = 5oC
4.6
COP
5.2
COP
5.0
Compressor speed=900 rpm
4.4 4.2
Compressor speed = 4200 rpm
ER=0.1 ER=0.2 ER=0.6
(c)
SC = 3oC SH = 5oC
3.5
COP
6.0
3.0 2.5
4.6 4.4 4.2
4.0
2.0
3.8
1.5
3.6
1.0
4.0 3.8 4
5
6
7
8
9
4
5
6
7
8
9
4
5
6
7
Mixing section diameter (mm)
Fig. 4 COP according to mixing section diameter for different ERs.
8
9
ER=0.1 ER=0.2 ER=0.6
(a)
SC = 3oC SH = 5oC
10 8 6 4 2 0
12
20
Compressor speed=2340 rpm
ER=0.1 ER=0.2 ER=0.6
(b)
10
SC = 3oC SH = 5oC
8 6 4 2 0
Pressure lifting ratio (%)
12
14
Compressor speed=900 rpm
Pressure lifting ratio (%)
Pressure lifting ratio (%)
14
5
6
7
8
9
ER=0.1 ER=0.2 ER=0.6
(c)
SC = 3oC SH = 5oC
10 5 0 -5 -10 -15
-2 4
15
Compressor speed = 4200 rpm
4
5
6
7
8
9
4
5
6
7
8
Mixing section diameter (mm)
Fig. 5 Pressure lifting ratio according to mixing section diameter for different ERs.
9
7.0
SC = 3oC, SH = 5oC ER=0.1 ER=0.2 ER=0.6
6.8
CSPF
6.6 6.4 6.2 6.0 5.8 4
5
6
7
Mixing section diameter (mm)
Fig. 6 CSPF according to mixing section diameter for different ERs.
200 10 100 5
0
3.0
2.8 1.1 2.6
(b) 2.2
0 0.0
0.1
0.2
0.3
0.4
0.5
0.6
0.2
0.3
0.4
0.5
0.6
0.7
Entrainment ratio
o
SC = 3 C, SH = 5 C, Dm = 5.0 mm Compressor speed = 4200 rpm Cooling capacity Compressor power
11
0.1
3.65
3.55 10 3.45 9
Compressor power (kW)
Cooling capacity (kW)
o
0.9 0.0
0.7
Entrainment ratio
12
1.0
2.4
(a) -100
1.3 o SC = 3 C, SH = 5oC, Dm = 5.0 mm Compressor speed = 4200 rpm COP COPi 1.2
COPi
300
3.2
COP
20 o SC = 3oC, SH= 5 C, Dm = 5.0 mm Compressor speed = 4200 rpm PLR Total flow 15 Entrained flow
Pressure lifting ratio (%)
-1
Mass flow rate (kgh )
400
(c) 8
3.35 0.0
0.1
0.2
0.3
0.4
0.5
0.6
0.7
Entrainment ratio
Fig. 7 (a) Mass flow rate and pressure lifting ratio according to ER, (b) COP and COPi , (c) Cooling capacity and compressor work at a compressor speed of 4200 rpm, (d) P-h diagrams of the DEEC and baseline cycle at a compressor speed of 4200 rpm and ER of 0.2.
12
o
o
Pressure lifting ratio (%)
SC = 3 C, SH = 5 C, Dm = 5.0 mm 10
ER=0.1 ER=0.2 ER=0.6
8 6 4 2
(a)
0 1000
2000
3000
4000
5000
Compressor speed (rpm)
280
o
o
240
-1
Mass flow rate (kgh )
SC = 3 C, SH = 5 C
200
ER=0.1 ER=0.2 ER=0.6 Baseline cycle
160 120 80
(b) 40 1000
2000
3000
4000
5000
Compressor speed (rpm)
Fig. 8 Pressure lifting ratio and total mass flow rate according to compressor speed for different ERs.
1.4 COP at ER=0.1 COP at ER=0.2 COP at ER=0.6 COPi at ER=0.1
5
COPi at ER=0.2
COP
COPi at ER=0.6
4
1.3 1.2 1.1 1.0
3 0.9 o
2
o
SC = 3 C, SH = 5 C, Dm = 5.0 mm 1000
2000
3000
4000
0.8 5000
Compressor speed (rpm)
Fig. 9 COP and COPi according to compressor speed.
COPi
6
7.2 6
COP
5
7.0 6.8 6.6
4
6.4
CSPF
COP at 900 rpm COP at 2340 rpm COP at 4200 rpm CSPF, Ejector cycle CSPF, Baseline cycle
6.2 3 6.0 o
2
o
SC = 3 C, SH = 5 C, Dm = 5.0 mm 0.0
0.1
0.2
0.3
0.4
5.8 0.5
0.6
0.7
Entrainment ratio
Fig. 10 COP and CSPF according to ER for different compressor speeds.
0.14 Entrainment ratio = 0.2 Dm = 4.0 Dm = 5.0 Dm = 7.0 Time weighting factor
CSPFbin
10
0.12 0.10 0.08
8
0.06 6 0.04 4
Time weighting factor
12
0.02
2
0.00 20
22
24
26
28
30
32
34
36
Outdoor temperature (oC)
Fig. 11 CSPFbin and bin time weighting factor of the ejector cycle according to outdoor temperature for different mixing section diameters.
12
Baseline cycle Ejector cycle Ejector effect
7.0 14 12
Baseline cycle Ejector cycle Ejector effect
6.8 10
CSPFbin
10 6.6
8
12 14 10 12 8
Baseline cycle Ejector cycle Ejector effect
30 8
6
6
6.4
6 4
6.2 2 0
6.0 21 22 23 24
20 6
4
4
50 40
10
8
60
10 4
0
2
2
2
-10
0
0
0
-20
25 26 27 28 29 30
CSPFbin.i (%)
14
31 32 33 34 35
Outdoor temperature (oC)
Fig. 12 CSPFbin and CSPFbin.i of the ejector cycle and the baseline cycle according to outdoor temperature.
Figure captions
Fig. 1
Schematic and P-h diagrams of the DEEC.
Fig. 2
Schematic diagram of the experimental setup.
Fig. 3
Schematic diagram of the ejector assembly.
Fig. 4
COP according to mixing section diameter for different ERs.
Fig. 5
Pressure lifting ratio according to mixing section diameter for different ERs.
Fig. 6
CSPF according to mixing section diameter for different ERs.
Fig. 7
(a) Mass flow rate and pressure lifting ratio according to ER, (b) COP and COPi , (c) Cooling capacity and compressor work at a compressor speed of 4200 rpm, (d) P-h diagrams of the DEEC and baseline cycle at a compressor speed of 4200 rpm and ER of 0.2.
Fig. 8
Pressure lifting ratio and total mass flow rate according to compressor speed for different ERs.
Fig. 9
COP and COPi according to compressor speed.
Fig. 10 COP and CSPF according to ER for different compressor speeds. Fig. 11 CSPFbin and bin time weighting factor of the ejector cycle according to outdoor temperature for different mixing section diameters. Fig. 12 CSPFbin and CSPFbin.i of the ejector cycle and the baseline cycle according to outdoor temperature.
Table 1 Specifications of the tested ejector Item
Ejector
Part
Value
Length of the converging part of the motive nozzle
12 mm
Length of the diverging part of the motive nozzle
24 mm
Length between the motive nozzle and the inlet of the mixing section
0 mm
Diameter of the nozzle throat
2.4 mm
Diameter of the nozzle outlet
2.7 mm
Angle of the diverging part of the motive nozzle Angle of the diffuser
1.02° 5°
Table 2 Specifications of the DEEC and the baseline cycle. Item Compressor
Condenser Evaporator #1
Specifications Type Working fluid
DEEC Rotary R410A
Baseline cycle Rotary R410A
Unit
Charge amount Type Capacity Type
3 Plate heat exchanger 12 Cartridge heater
2 Plate heat exchanger 12 Cartridge heater
kg
Capacity
12
12
KW
Evaporator #2 Type Capacity Orifice Needle valve diameter
Cartridge heater 12 6.4
None 6.4
KW
KW mm
Table 3 Test conditions and ejector geometries. Parameter
Value
Compressor speed (rpm)
900, 2340, 4200
Entrainment ratio, ER
0, 0.1, 0.2, 0.4, 0.6
Condensing pressure (kPa)
3105
2763
2496
Low-evaporating pressure (kPa)
1027
1184
1465
Condensing temperature (°C)
51
46
41
Low-evaporating temperature (°C)
8
13
20
Subcooling (°C)
3
Superheat (°C) Mixing diameter (mm)
5 4
5
7
9
Mixing diameter length (mm)
40
50
70
90
Table 4 Specifications of the measuring devices. Range
Accuracy
Number of Measuring points
0–5 MPa
±0.3%.
10
Thermocouple (T-type)
-200–200 °C
±0.2 °C
10
Mass flow meter
0–300 kg h-1
±0.1%
2
0–300 V, 0–20 A
±0.2%
3
Device Pressure transducer
Power meter
Highlights
The performance of an R410A air-conditioner adopting a DEEC.
The effects of operating parameters on the COP and CSPF.
Optimized mixing section diameter of the ejector.
The maximum limit of the entrainment ratio (ER).