• Proceedings of the SAREK Conference
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• 2008.11a
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• pp.119-124
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• 2008

Recently, focus has been drawn on natural refrigerants due to increasing concern on global warming. As a consequence, CO2 systems such as a heat pump water heater using CO2 as a refrigerant are rapidly growing on the market. Currently, rolling piston rotary compressors are widely used for CO2 heating and/or refrigeration systems. There are several ways of realizing gas compression structure. They are single stage compression with single cylinder, single stage compression with two cylinders, and two stage compression with two cylinders. In this paper, computer simulation program which was validated for a single stage rotary compressor with one cylinder has been extended for a single stage, two cylinder rotary type. Numerical investigation has been made on optimal design for the cylinder configuration using the extended simulation program. For a single stage two cylinder rotary compressor having a displacement volume of 4 cc for each cylinder, compressor efficiency has been found to be maximum when the cylinder radius and height are 31mm and 10mm, respectively.

• Proceedings of the SAREK Conference
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• 2005.11a
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• pp.398-403
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• 2005

The cooling performance of a transcritical $CO_2$ cycle varies significantly with a variation of refrigerant charge amount. In this study, the performance of the $CO_2$ system was measured and analyzed by varying refrigerant charge amount with a change of cycle option. The applied cycle options are the single-stage compression system, two-stage compression with 1-EEV system, and two-stage compression with 2- EEV system. The optimum normalized charge were 0.363, 0.297, and 0.282 for the two-stage compression with 2-EEV system, two-stage compression with 1-EEV system, and single-stage compression system, respectively.

• Progress in Superconductivity and Cryogenics
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• v.15 no.1
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• pp.45-50
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• 2013

Closed-loop J-T (Joule-Thomson) refrigeration cycle is advantageous compared to common open loop $N_2$ decompression system in terms of nitrogen consumption. In this study, two closed-loop pure $N_2$ J-T refrigeration systems with sub-atmospheric device for cooling High Temperature Superconductor (HTS) power cable are investigated. J-T cooling systems include 2-stage compressor, 2-stage precooling cycle, J-T valve and a cold compressor or an auxiliary vacuum pump at the room temperature. The cold compressor and the vacuum pump are installed after the J-T valve to create sub-atmospheric condition. The temperature of 67 K is possible by lowering the pressure up to 24 kPa at the cold part. The optimized hydrocarbon mixed refrigerant (MR) J-T system is applied for precooling stage. The cold head of precooling MR J-T have the temperature from 120 K to 150 K. The various characteristics of cold compressor are invstigated and applied to design parameter of the cold compressor. The Carnot efficiency of cold compressor system is calculated as 16.7% and that of vacuum pump system as 16.4%. The efficiency difference between the cold compressor system and the vacuum pump system is due to difference of enthalpy change at cryogenic temperature, enthalpy change at room temperature and different work load at the pre-cooling cycle. The efficiency of neon-nitrogen MR J-T system is also presented for comparison with the sub-atmospheric devices. These systems have several pros and cons in comparison to typical MR J-T systems such as vacuum line maintainability, system's COP and etc. In this paper, the detailed design of the subcooled $N_2$ J-T systems are examined and some practical issues of the sub-atmospheric devices are discussed.

• Korean Journal of Air-Conditioning and Refrigeration Engineering
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• v.24 no.7
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• pp.585-590
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• 2012

The objective of this study is to investigate the performance of a two-stage compression heat pump system for district heating. The experimental setup of heat pump consists of compressor, condenser, evaporator, expansion device, intercooler, flash tank, oil separator and accumulator. The experimental evaluations on the two-stage compression cycle were carried out under various operating conditions which were heat source temperature, the degree of compressor inlet superheat, and intermediate pressure. The temperature ranges of unutilized energy as the heat source were used in the test conditions. As the heat source temperature increased from $10^{\circ}C$ to $30^{\circ}C$, the COP and heating capacity of the heat pump system increased by 22.6% and 45.8%, respectively. The performance of the two-stage heat pump system increased by 5.2% with the variation of the intermediate pressure in the same heat source temperature conditions.

• Progress in Superconductivity and Cryogenics
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• v.15 no.2
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• pp.1-8
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• 2013

This paper reviews the development status of magnetic refrigeration system for hydrogen liquefaction. There is no doubt that hydrogen is one of most important energy sources in the near future. In particular, liquid hydrogen can be utilized for infrastructure construction consisting of storage and transportation. Liquid hydrogen is in cryogenic temperatures and therefore high efficient liquefaction method must be studied. Magnetic refrigeration which uses the magneto-caloric effect has potential to realize not only the higher liquefaction efficiency > 50 %, but also to be environmentally friendly and cost effective. Our hydrogen magnetic refrigeration system consists of Carnot cycle for liquefaction stage and AMR (active magnetic regenerator) cycle for precooling stages. For the Carnot cycle, we develop the high efficient system > 80 % liquefaction efficiency by using the heat pipe. For the AMR cycle, we studied two kinds of displacer systems, which transferred the working fluid. We confirmed the AMR effect with the cooling temperature span of 12 K for 1.8 T of the magnetic field and 6 second of the cycle. By using the simulation, we estimate the total efficiency of the hydrogen liquefaction plant for 10 kg/day. A FOM of 0.47 is obtained in the magnetic refrigeration system operation temperature between 20 K and 77 K including LN2 work input.

• Proceedings of the SAREK Conference
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• 2008.06a
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• pp.304-308
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• 2008

The optimal design of two stage evaporation & absorption system which is related to the large temperature difference system was investigated numerically in the absorption refrigeration system. The concentrations at inlet & oulet of absorber are 62.9% and 56.9%, but in two stage absorption system the values are 62.2% and 56.2%. Therefore strong solution & weak solution became diluted than the standard value. The amount of weak solution circulation can be reduced in absorption refrigeration system, and the sensible heat load is more reduced to enhance the COP of system. As UAR is increased, COP becomes larger, and this means the role of top section is more important than bottom section in two stage evaporation & absorption system. But the increase of COP becomes slower at 0.7 of UAR ratio. The performance of Type2 is higher than Type1 in COP with the flow direction of cooling waters. This phenomena is due to the active absorption of vapor -absorption & lower temp. cooling water is more effective. The pressure at bottom section becomes higher & that at top section becomes lower and therefore the circulation rate can be diminished more.

• Proceedings of the SAREK Conference
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• 2008.11a
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• pp.203-211
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• 2008

The optimal design of two stage evaporation & absorption system which is related to the large temperature difference system was investigated numerically in the absorption refrigeration system. The concentrations at inlet & oulet of absorber are 62.9% and 56.9%, but in two stage absorption system the values are 62.2% and 56.2%. Therefore strong solution & weak solution became diluted than the standard value. The amount of weak solution circulation can be reduced in absorption refrigeration system, and the sensible heat load is more reduced to enhance the COP of system. As UAR is increased, COP becomes larger, and this means the role of top section is more important than bottom section in two stage evaporation & absorption system. But the increase of COP becomes slower at 0.7 of UAR ratio. The performance of Type2 is higher than Type1 in COP with the flow direction of cooling waters. This phenomena is due to the active absorption of vapor -absorption & lower temp. cooling water is more effective. The pressure at bottom section becomes higher & that at top section becomes lower and therefore the circulation rate can be diminished more.

• International Journal of Air-Conditioning and Refrigeration
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• v.7
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• pp.55-68
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• 1999

In this study, computer simulation programs were developed for single-stage, two-stage, and three-stage condensation heat pumps and their performance with CFC11, HCFC123, HCFC141b was examined under the same external conditions. The results showed that the coefficient of performance(COP) of an optimized 'non-split type' three-stage condensation heat pump is 25-42% higher than that of a conventional single-stage heat pump. The increase in COP, however, differed among the fluids tested. The improvement in COP is largely due to the decrease in average LMTDs in condensers, which results in the decrease in thermodynamic irreversibility in heat exchange process. For the three-stage heat pump, the highest COP is achieved when the total condenser area is evenly distributed among the three condensers. For the two-stage heat pump, however, the optimum distribution of the total condenser area varies with an individual working fluid. For the three-stage system, 'splitting the condenser cooling water'for the use of intermediate and high pressure subcoolers helps increase the COP further. When the individual cooling water entering the intermediate and high pressure subcoolers is roughly 10% of the total condenser cooling water, the maximum COP is achieved showing roughly an 11% increase in COP as compared to that of the 'non-split type' heat pump.

• Korean Journal of Air-Conditioning and Refrigeration Engineering
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• v.18 no.8
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• pp.641-648
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• 2006

In this study, a two-stage phase-separate cycle was investigated analytically to improve the performance of the $CO_2$ system in the cooling mode. The simulation results were verified with the measured data. The predictions using the simulation model were consistent with the measured data within ${\pm}20%$ deviations. The performance of the modified $CO_2$ system with the two-stage phase-separated cycle was analyzed with the variations of outdoor temperature and EEV opening. The cooling COP decreased with the increase of compressor frequency. The highest COP was 2.7 at compressor frequencies of 30 Hz and 30 Hz for the first and second compressors, respectively. In addition, the cooling COP increased by 9.3% with an application of optimum control of the first and second-stage EEV openings.

• Korean Journal of Air-Conditioning and Refrigeration Engineering
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• v.11 no.3
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• pp.349-358
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• 1999

Thermodynamic cycle analysis has been performed to maximize the liquid amount for various hydrogen liquefaction systems using GM(Gifford-McMahon) refrigerator. Since the present authors' previous experiments showed that the liquefaction rate was approximately 5.1mg/s in a direct contact with a commercial GM refrigerator, the purpose of this study is to predict how much the liquefaction rate can be increased in different configurations and with improved heat exchanger performance. The optimal operating conditions have been analytically sought with real properties of normal hydrogen for the single-stage GM precooled L-H(Linde-Hampson) system, the two-stage GM direct contact system, the two-stage GM precooled L-H system and the two-stage helium GM-JT (Joule-Thomson) system. The maximum liquefaction rate has been predicted to be only about 7 times greater than the previous experiment, when the two-stage precooling is employed and the effectiveness of heat exchangers approaches to 99.0%. It is concluded that the liquefaction rate is limited mainly by the cooling capacity of the current GM refrigerators and a larger scale of hydrogen liquefaction is possible with a greater capacity of cryocooler at 60-70 K range.