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Potential energy benefits of integrated refrigeration system with microturbine and absorption chiller
International Journal of Refrigeration 27 (2004) 816–829 www.elsevier.com/locate/ijrefrig
Potential energy benefits of integrated refrigeration system with microturbine and absorption chiller Yunho Hwang* Department of Mechanical Engineering, Center for Environmental Energy Engineering, University of Maryland, Martin Hall Building, Rm 3135, College Park, MD 20742-3035, USA Received 14 April 2003; received in revised form 8 September 2003; accepted 7 January 2004
Abstract This paper presents and analyzes the performance potential of a refrigeration system that is integrated with a microturbine and an absorption chiller (RMA). The waste heat from the microturbine operates the absorption chiller, which provides additional cooling. This additional cooling capacity can be utilized either to subcool the liquid exiting the condenser of the refrigeration system or to precool the air entering the condenser in the refrigeration system. Moreover, any surplus cooling capacity not utilized in the subcooler can be utilized to precool the microturbine intake air. The additional assistance to the refrigeration system enhances the efficiency of the refrigeration cycle, which in turn reduces the required microturbine size. The smaller size of the microturbine enhances the part load efficiency, especially in lower ambient temperatures. With increased microturbine efficiency, RMA with subcooler, RMA with subcooler and microturbine intake air precooler, and RMA with condenser air precooler can reduce the annual energy consumption by 12, 19, and 3%, respectively, as compared to a refrigeration system operating without any waste heat utilization from the microturbine. Therefore, RMA with subcooler and microturbine intake air precooler has the best potential of energy savings. The payback period of RMA with subcooler and microturbine intake air precooler is estimated in 3 years, which facilitates it as an economically feasible solution among the options investigated. q 2004 Published by Elsevier Ltd and IIR. Keywords: Design; Refrigerating system; Cogeneration; Generation; Electricity; Absorption system; Performance
Syste`me frigorifique muni d’une microturbine et d’un refroidisseur a` absorption: avantages e´nerge´tiques potentielles Mots-cle´s: Conception; Syste`me frigorifique; Coge´ne´ration; Ge´ne´ration; Electricite´; Syste`me a` absorption; Performance
1. Introduction 1.1. Environmentally friendly energy use Greenhouse warming occurs when carbon dioxide, released mostly from the burning of fossil fuels (oil, natural * Corresponding author. Tel.: þ1-301-405-5247; fax: þ 1-301405-2025. E-mail address: [email protected] (Y. Hwang).
gas and coal), and other gases such as methane, nitrous oxide, ozone, CFCs, HCFCs, and water vapor, build up in the atmosphere, increasing the Earth’s natural greenhouse effect. This buildup is a concern that has resulted in the Kyoto Protocol [1], which sets greenhouse gas emission objectives for each developed country. The Kyoto Protocol aims at a 5% reduction in greenhouse gas emissions from 1990 levels over the 2008–2012 period. Policies to reduce global warming require industries to develop technologies that will reduce energy consumption and increase energy efficiency.
0140-7007/$ - see front matter q 2004 Published by Elsevier Ltd and IIR. doi:10.1016/j.ijrefrig.2004.01.007
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Nomenclature CEEE Center for Environmental Energy Engineering COP Coefficients of performance EES Engineering Equation Solver EPRI Electric Power Research Institute Phigh High pressure level in the absorption cycle Plow Low pressure level in the absorption cycle PR Pressure ratio Qmt;precooler Cooling capacity used to precool the microturbine intake air Qsubcool Cooling capacity available from the subcooler Qsurplus Surplus cooling capacity available after the subcooler is satisfied RMA Refrigeration system integrated with microturbine and absorption chiller Rsubcooling Amount of cooling capacity used for the subcooling as compared to the maximum cooling capacity available from the absorption chiller Tamb Ambient temperature hise Isentropic efficiency hvol Volumetric efficiency
1.2. Demands for on-site power generation The electricity consumption in the US during 1990s has increased by approximately 2% annually [2]. However, the electric grid is becoming increasingly susceptible to service interruptions due to peak demands. Grid reliability becomes a critical issue in a deregulated market as clearly illustrated by the recent power outages in California. These power outages have drawn attention to on-site power generation. Increased power demands by the cooling systems and IT systems of the existing old buildings, which may have only limited power supply also requires on-site power generation in addition to the existing power supply. Therefore, the onsite power generation becomes more important, not only to ease Electric Peak load but also as means to satisfy new demands independently from the grid.
1.3. On-site power generation For on-site power generation, engines and combustion turbines have been available for many years in a wide range of capacities. In the late 1990s, microturbines and fuel cells emerged as new on-site power generation technologies, targeted especially for small capacity ranges between 1 and 400 kW [3]. Microturbines are small-scale electric generators that are based on the Brayton cycle whose electrical output ranges from 30 to 500 kW. Recent successful development of microturbines enables commercialization by achieving maintenance intervals longer than 8000 h in advance of fuel cells. Microturbines customarily use natural gas as their energy source. High efficiencies of 27 – 30% are reached by using recuperation.
2. Refrigeration system integrated with microturbine and absorption chiller Similar to the integration of the cooling, heating, and power generation to commercial buildings, a refrigeration system can be integrated with a power generation system (microturbine) and an absorption chiller. Hereafter, such a system is referred to as RMA. Electricity in an RMA is provided by the microturbine. The waste heat from the microturbine is utilized to operate the absorption chiller. Then, the additional cooling capacity by the absorption chiller can be utilized in the refrigeration system in the following ways: † Direct refrigeration (2 20 8C) which requires an NH3/H2O absorption chiller † Subcooling the condenser outlet refrigerant of the refrigeration system † Precooling the air entering the microturbine † Precooling the air entering the condenser of the refrigeration system In this paper the latter three options were investigated in detail in anticipation of integrating an air-cooled H2O/LiBr absorption chiller, as will be discussed later. An energy flow diagram and a schematic diagram of an RMA are found in Figs. 1 and 2. As illustrated, the microturbine produces electricity; its exhaust, ranging from 232 to 260 8C drives the absorption chiller. The electricity produced by the microturbine is used to drive the vapor compression refrigeration system, and the chilled water produced by the absorption chiller is used to: further subcool the refrigerant from the condenser (extended solid line from the absorption chiller in Fig. 2); first subcool the refrigerant from the condenser and precool the microturbine intake air with any surplus capacity (extended dashed line from the
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Fig. 1. Energy flow in RMA.
subcooler in Fig. 2) and; precool the cooling air entering the condenser (extended dashed line from the absorption chiller in Fig. 2). The design aspects and performance potential of these three RMAs were theoretically investigated in detail as described next. 2.1. Absorption chiller An absorption chiller is a piece of heat-operated refrigeration equipment, whose cooling capacity is produced using waste heat instead of electricity as the
primary energy source. This ability provides an energy saving opportunity in cases where the cooling demand coincides with surplus heat. The cooling capacity of the current study ranges from 10 to 70 kW. The Gas Research Institute lists eight major manufacturers of absorption chillers in the US [4]. However, only two manufacturers produce absorption chillers in the capacity range of interest herein. One manufacturer produces 4 – 18 kW capacity absorption chillers with cooling tower [5] and the other manufacture produces 4–25 kW capacity absorption chillers with air-cooled condenser/absorber [6].
Fig. 2. Schematic diagram of RMA.
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2.1.1. Working fluids In the absorption cycle, H2O/LiBr and NH3/H2O are the two preferred refrigerant- and absorbent-pairs. While H2O/LiBr absorption systems operate at vacuum pressures between 0.8 and 20 kPa, NH3/H2O absorption systems operate at much higher pressures, between 400 and 2500 kPa. A disadvantage of H2O/LiBr is its narrow solution field, which is limited by crystallization, resulting in an absorber temperature limitation of approximately 40 8C thus requiring a cooling tower. NH3/H2O can operate at higher temperatures but the toxicity of NH3 requires careful attention to system design, installation, and maintenance. H2O/LiBr is used for small to large capacity applications (residential to industrial installations from 10 to 26,000 kW) while NH3/H2O is used for either very small (small refrigerator and residential application less than 70 kW) or large capacity installations. The cooling capacity of the current study ranges from 10 to 70 kW. However, the results generated in this study can be scaled to any capacity. Since H2O/LiBr is used in the capacity range of interest herein and also non-toxic, it was chosen as the working fluid in this study. 2.1.2. Modeling of single-stage H2O/LiBr absorption chiller A single-stage H 2O/LiBr absorption system was modeled using the computer software Engineering Equation Solver [7]. The following assumptions were used in the model: † Evaporator: † Refrigerant is pure water. † Stream exiting the evaporator is saturated vapor. † Pressure level is Plow : † Condenser: † Refrigerant is pure water. † Saturated liquid leaves the condenser. † Air-cooled condenser is used. The approach temperature of the condenser varies depending upon the performance of the condenser and boundary conditions. To make an economically feasible assumption on the performance of the condenser, the approach temperature of the condenser is based on the approach temperature of the commercially available air-cooled condenser, approximately 10 K [8]. † Pressure level is Phigh as determined by the condenser temperature. † Absorber: † Saturated liquid leaves the absorber at the condensing temperature. † Air-cooled absorber is used. Approach temperature of the absorber is 10 K as described for the condenser. The concentration difference between the rich and poor streams is 4%. † Pressure level is Plow as determined through the evaporator temperature.
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† Solution pump: † Efficiency is 100%. † Saturated liquid enters the pump from the absorber. † Pressure level increases from Plow to Phigh : † Solution heat exchanger: † Exiting rich solution does not enter the two-phase region (maximum enthalpy is for saturated liquid). † Entering poor solution is saturated liquid (at generator conditions). † An approach temperature is used in the heat transfer between the exiting poor solution and the entering rich solution ðDT ¼ 10 KÞ: (Here, ‘poor solution’ refers to a solution with a relatively low mass fraction of sorbent (LiBr) and accordingly a relatively high mass fraction of refrigerant (H2O); ‘rich solution’ refers to a solution with a relatively high mass fraction of sorbent and low mass fraction of refrigerant.) † Pressure level is Phigh : † Generator: † The generator heat is supplied from the microturbine waste heat. † Saturated water vapor leaves the generator at the same temperature as the entering poor solution. † Saturated liquid leaves the generator as the rich solution. † Pressure level is Phigh : † Operating conditions: † For water-cooled system, the cooling water temperature from the cooling tower is assumed to be 5.6 K higher than the wet-bulb temperature that is determined by the ambient temperature and relative humidity. † The relative humidity of the ambient air is constant at 40% RH for all ambient temperatures. 2.2. Microturbine There are only limited numbers of microturbine manufactures in the US. These manufactures produce 30 – 250 kW capacity microturbines [9– 11]. The efficiencies of these microturbines are 26 – 28% at ISO conditions (15 8C at sea level) based on natural gas low heating value. Three important characteristics of the microturbine performance should be considered: the dependence of the microturbine efficiency on the ambient temperature, the dependence of the power factor on the ambient temperature, and the part load efficiency degradation. Here, the power factor is defined as the ratio of the microturbine power output to its rated power output. Characteristics of typical microturbines are illustrated in Fig. 3 [12,13]. As the ambient temperature increases the efficiency and power factor of the microturbine both decrease (Fig. 3(a)). The main contribution to this performance degradation is the reduction in the intake air and fuel flow rate as the air and fuel density decrease. The
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Fig. 3. Characteristics of start-of-the-art microturbine. (a) Performance of microturbine vs. ambient temperature [12]. (b) Part load efficiency of microturbine vs. part load factor [13].
power factor decreases as the ambient temperature increases greater than 26.7 8C although the turbine speed and fuel gas compressor speed are maintained at their maximum. However, the power factor can be maintained by adjusting the turbine speed and fuel gas compressor speed when the ambient temperature decreases below 26.7 8C. It should be noted that this performance dependence on the ambient temperature can be changed through a different design approach. During the partial load operation, the turbine speed and fuel supply are adjusted. With a significant turbine speed reduction, the microturbine efficiency decreases during the partial load operation (Fig. 3(b)). The efficiency during the partial load operation slowly degrades by 10% when the partial load factor decreases to 0.5. However, the partial load efficiency degradation becomes significant when the partial load factor becomes less than
0.5. Therefore, it is important to size the microturbine output such that the partial load factor is maintained greater than 0.5. Since only limited information on these microturbine characteristics is available from the open literature, the microturbine characteristics illustrated in Fig. 3 are used in this study. To recover the waste heat from the microturbine, the exhaust pipe of the microturbine is extended to the absorption chiller and the exhaust gas passes through the generator of the absorption chiller. Then this additional process causes a back pressure to the exhaust (it was approximately 1 kPa [12]) and can be compensated by either modifying the combustion fan of the generator or adding additional exhaust fan. The waste heat flow from the microturbine to the absorption chiller was assumed based on Cowie et al. [12] as indicated below.
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Fig. 4. COP and COP improvement of refrigeration system with subcooler over base refrigeration system.
38% of the fuel energy from the microturbine exhaust gas is delivered to the absorption chiller. Of this delivered heat, only 69% is utilized by the absorption chiller. It should be noted that this assumption might be different for different systems. 2.3. Performance improvement of refrigeration system with absorption chiller The performance of the RMA is analyzed in two respects: refrigeration cycle efficiency enhancement and overall system energy use reduction. First, the enhancement of the refrigeration cycle efficiency is analyzed for three options. It should be noted that only the results for the aircooled absorption system case are presented here since similar performance enhancement was observed for the water-cooled absorption system case. 2.3.1. Refrigeration system with condenser subcooler In the subcooler cycle using an absorption chiller, the subcooler cools the refrigerant exiting the condenser and provides lower vapor quality refrigerant to the evaporator, which then provides a larger cooling capacity for the same refrigerant mass flow rate of the original vapor compression system. The performance of the RMA with subcooler using R22 as its refrigerant was modeled using EES. The following assumptions were used in the model: † Evaporator: † Refrigeration capacity is 100 kW. Refrigerant mass flow rate is adjusted for each case. † Evaporation temperature is 10 K lower than the refrigerated air temperature. † Superheat of vapor is 5 K exiting the evaporator. † Pressure drop is 50 kPa. † Condenser: † Condensing temperature is 10 K higher than the inlet air temperature.
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† Degree of subcooling at condenser outlet is 5 K. † Pressure drop is 50 kPa. † Compressor: † Isentropic and volumetric efficiencies (hise ; hvol ) change depending upon the pressure ratio (PR) as given in the following equations [14]. † hise ¼ 0:85 2 0:0467PR; hvol ¼ 1:08 2 0:04PR † Subcooler: † H 2 O/LiBr absorption chiller provides the additional subcooling. † Absorption chiller works only for an ambient temperature between 7.2 and 40.6 8C. † The subcooled liquid refrigerant and chilled water exchange heat in counterflow configuration. The approach temperature is 2 K. † The approach temperature limits the degree of subcooling. † The subcooler does not affect the pressures of the evaporator and condenser. † When the air-cooled absorption chiller is used, the evaporating temperature of the absorption chiller is controlled dynamically according to the ambient temperature (Table A2). When the water-cooled absorption chiller is used, the evaporating temperature is maintained at 3 8C independent of the ambient temperature. † Fan motor: † Airflow rate through the evaporator is 0.0537 m3/s for 1 kW refrigeration capacity. † Power input to the evaporator and condenser fan motor is 775 W for 1 m3/s airflow rate. † Evaporator motor is located in the cold air stream and works as thermal load.
The COP and COP improvement of the refrigeration system with subcooler by the air-cooled absorption chiller over the base refrigeration system at various bin temperatures is depicted in Fig. 4. The COP decreases but the COP improvement increases as the ambient temperature increases (Fig. 4). The COP improvement starts at 7.2 8C since the absorption chiller starts from this temperature bin. The maximum COP improvement is 4 and 25% at 7.2 and 40.6 8C, respectively. The subcooling energy use ratio ðRsubcooling Þ of the refrigeration system with condenser subcooler is illustrated in Fig. 5(a). Here, the Rsubcooling is defined as the actual amount of cooling energy used for the subcooling as compared to the maximum cooling capacity available from the absorption chiller. The cooling capacity by the absorption chiller is used only up to 52% (Fig. 5(a)). In other words, the subcooling capacity is great enough to reach the minimum possible subcooler outlet temperature for the entire range of ambient temperatures. This results in the limitation of COP improvement.
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Fig. 5. Use of surplus cooling in refrigeration system with subcooler. (a) Subcooling energy use ratio of cooling from air-cooled absorption chiller. (b) Surplus cooling available after subcooling and cooling capacity used for microturbine intake air precooling.
2.3.2. Refrigeration system with subcooler and microturbine intake air precooler Since the utilization of the subcooling capacity from the subcooler is limited, the surplus cooling capacity by the subcooler can be employed to precool the intake air entering the microturbine. The surplus cooling capacity ðQsurplus Þ and the amount of the cooling capacity used to precool the microturbine intake air ðQmt;precooler Þ are compared to the subcooling capacity available from the subcooler ðQsubcool Þ (Fig. 5(b)). The surplus cooling capacity decreases as the ambient temperature increases (Fig. 5(b)). If this surplus cooling is utilized to precool the intake air entering the microturbine, the microturbine efficiency can be improved. For example, the microturbine efficiency improvement is 13% if all the surplus cooling capacity is used at 40.6 8C. However, heat transfer limits the utilization of the surplus cooling in most cases because the intake air temperature cannot be lower than the subcooler temperature. To account for this heat transfer limitation, the ambient air temperature
was assumed to be higher than that of the subcooler temperature by 2 K. The cooling capacity used to precool the intake air increases as the ambient temperature increases but its utilization is clearly limited especially at low ambient temperature (Fig. 5(b)). This heat transfer limitation results in the limitation of the microturbine efficiency improvement. Moreover, it means that the current waste heat recovery is large enough and more waste heat recovery can not contribute to the further enhancement. 2.3.3. Refrigeration system with condenser air precooler In the precooler cycle using an absorption chiller, the precooler cools the air entering the condenser and supplies the air at reduced temperature. The lower air temperature entering the condenser results in a lower pressure ratio and reduced compressor power consumption. The performance of the RMA with precooler using R22 was modeled using EES. The same assumptions used for the evaporator, condenser and compressor of the cycle with subcooler were applied in the
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Fig. 6. COP improvement and condenser intake air temperature reduction of refrigeration system with precooler over base refrigeration system.
model. The following assumptions for the precooler were used in addition to the previous assumptions: † Precooler: † Airflow rate is the same as that of the conventional cycle to increase the air temperature by 5 K across the condenser. † H2O/LiBr absorption chiller provides the precooling. † Absorption chiller works only for ambient temperature between 7.2 and 40.6 8C. † Precooler does not affect the evaporation pressure. † Refrigerant and air heat exchange occurs in crosscounterflow configuration. The approach temperature is 2 K. † The same assumption for the evaporating temperature of the absorption chiller used for the subcooler cycle is used. The COP improvement and the condenser intake air temperature reduction of the refrigeration system with precooler over the base refrigeration system at various bin temperatures is depicted in Fig. 6. The COP improvement increases as the ambient temperature increases (Fig. 6). The COP improvement begins at 7.2 8C since the absorption chiller can start from this temperature bin. The maximum COP improvement is approximately 3 and 5% at 7.2 and 40.6 8C, respectively. This COP improvement is induced by the condenser inlet air temperature reduction as much as 0.9– 1.5 K (Fig. 6). The COP improvement is not limited by heat transfer in this case. 2.4. Annual energy consumption of RMA To analyze the overall performance potential of the RMA over the entire year, the annual energy consumption of RMA options were investigated as described next.
2.4.1. Comparison of annual energy consumption of RMA options The annual energy consumption of each of the RMA options operating alone with power supplied by the microturbine was investigated to compare the potential energy saving of each. The following assumptions were used in the calculation of the annual energy consumption by the RMA. † Bin temperature distribution and annual bin hours are based on US averages, which are the average of four US regional bin hours (Midwest, Southwest, Northeast, and Southeast) [15]. † The indoor temperature and refrigeration load (100 kW) are constant throughout the year. † Absorption chiller is used only when the ambient bin temperature is between 7.2 and 40.6 8C. Otherwise, only the base refrigeration system is used. † The performance of the microturbine is based on the results described in Section 2.2. † Efficiency of the refrigeration system assisted by the absorption chiller is based on the results described in Section 2.3. † If a surplus cooling capacity after the subcooling is available, it is used to precool the intake air to the microturbine. The calculated annual power consumption of RMA options using 100% microturbine as its power source are compared to each other in Table 1. RMA with subcooler, RMA with subcooler and microturbine intake air precooler, and RMA with condenser air precooler can reduce the annual energy consumption by 12, 19, and 3%, respectively, compared to the refrigeration system operating alone with power supplied by the microturbine without the integration of the absorption chiller. These energy savings are due to the refrigeration cycle efficiency enhancement, which is
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Tamb (8C)
Bin (h)
COP Base system
29.4 88 6.16 23.9 350 5.14 1.7 788 4.34 7.2 964 3.70 12.8 1139 3.16 18.3 1489 2.71 23.9 2102 2.32 29.4 1314 1.99 35.0 350 1.69 40.6 175 1.43 Annual energy consumption (kWh) Annual energy consumption ratio
Energy consumption (kWh) System with absorption chiller
Microturbine
Sub-cooler
Pre-cooler
Base system
System with subcooler
System with subcooler, MT precooler
System with condenser precooler
6.16 5.14 4.34 3.84 3.36 2.96 2.60 2.30 2.03 1.78
6.16 5.14 4.34 3.79 3.25 2.79 2.40 2.06 1.76 1.49
7599 34,668 88,710 122,780 164,755 244,969 396,720 285,095 88,727 54,819 1,488,844 1.00
7025 32,177 82,689 111,922 148,320 217,237 345,345 242,960 73,653 44,010 1,305,337 0.88
6739 30,943 79,738 107,139 140,557 203,600 319,419 222,255 66,524 37,770 1,214,685 0.81
7472 34,117 87,373 118,914 159,464 236,818 382,890 274,564 85,174 52,358 1,439,145 0.97
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Table 1 Comparison of energy consumption for US average bin hours (microturbine efficiency 26% at ISO condition)
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Fig. 7. Annual energy consumption ratio vs. microturbine efficiency.
especially effective at higher ambient temperatures. Increased efficiency of the refrigeration cycle reduces the required microturbine size by 20, 29, and 5%, respectively, for each option. The smaller size of the microturbine further enhances the partial load efficiency, especially for the lower ambient temperature. To further investigate the annual energy savings potential of each option in a typical US location, the energy consumption estimation was repeated for four US cities (Chicago, Los Angeles, Miami, and Seattle) and based on annual bin data for those cities [16]. Table 2 provides a comparison of annual energy consumption ratio between options as compared to the base system for four cities and US average. As shown in Table 2, the energy consumption ratio of RMA with subcooler and RMA with condenser air precooler is similar to that for the US average within a 2% variation, while the energy consumption ratio of the RMA with subcooler and microturbine intake air precooler is varied up to 5% from that for the US average. However, the rank of energy savings potential is maintained. It should be further noted, that the energy consumption ratio of each option at Miami is similar to that of the US average case.
2.4.2. Effects of microturbine efficiency on annual energy consumption of RMA As described earlier, the efficiencies of microturbines referred from the open literature are 26 – 28% at ISO conditions (15 8C at sea level). To investigate the effect of the microturbine efficiency on the annual energy consumption of each RMA option, cases of three microturbine efficiencies (26, 28 and 30%) are evaluated as illustrated in Fig. 7. When the microturbine efficiency increases the annual energy consumption of all cases including the base system decreases. Moreover, the trend of the energy saving of each option is remained the same for three efficiencies examined. Therefore, the annual energy saving of each option is almost similar for three microturbine efficiencies. 2.4.3. Comparison of annual energy consumption of RMA with air-cooled and water-cooled absorption chiller The annual energy consumption of the RMA with subcooler and microturbine intake air precooler by the water-cooled absorption system is only 1% less than or the same as that by the air-cooled absorption system. Therefore, the air-cooled absorption system is recommended since the
Table 2 Comparison of energy consumption for four US cities (microturbine efficiency 26% at ISO condition) City
Base system
System with subcooler
System with subcooler, MT precooler
System with condenser precooler
Chicago Los Angeles Miami Seattle US Average
1.00 1.00 1.00 1.00 1.00
0.90 0.89 0.87 0.90 0.88
0.84 0.84 0.80 0.86 0.81
0.97 0.97 0.96 0.97 0.97
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Table 3 Economic analysis of RMA options System
Base system
System with subcooler
System with subcooler, MT precooler
System with condenser precooler
Annual fuel cost ($) Annual fuel cost ratio MT size ratio MT cost ($) Absorption þ heat recover cost Total initial cost ($) Initial cost ratio Payback period (yr)
24,119 1.00 1.00 66,057 0 66,057 1.00 –
21,146 0.88 0.80 53,062 34,168 87,230 1.32 7
19,678 0.82 0.71 46,898 31,453 78,351 1.19 3
23,314 0.97 0.95 63,058 38,409 101,467 1.54 44
system design is much simpler and lower maintenance, compared to the water-cooled system, yet consumes a similar amount of energy to the water-cooled system. 2.5. Economic analysis An annual energy saving and an initial cost increase due to the additional heat recovery of each RMA options as compared to the base system were investigated to estimate the payback period of each option. The following assumptions were used in the calculation of the payback period. † Annual energy consumption is based on Table 1. † Fuel cost is based $4.5/106 kJ [17]. † Initial cost includes cost of devices (microturbine, absorption chiller, heat recovery) and their installation cost. † Costs of microturbine and heat recovery device except the absorption chiller are linear to their capacity. Whereas, the cost of absorption chiller per unit cooling capacity increases as the capacity decreases based on the profile described by EPRI [18]. † Microturbine cost is based on $800/kW [19]. † Absorption chiller cost is $440/kW to $490/kW depending upon the capacity [18]. † Heat recovery heat exchanger and subcooler or precooler costs are based on $28/kW [19]. Based on the above assumptions, an annual fuel cost saving, total initial cost increase and a simple payback period were calculated as summarized in Table 3. From the results, the annual energy cost savings of three options are 12, 19, and 3%, respectively, compared to the base system. Since the cost of the microturbine decreases as the energy saving increases, it is expected that the costs of the microturbine of three options are reduced by 20, 29, and 5%. However, the initial cost increases by the absorption chiller, which is approximately 60 – 70% of the microturbine cost in the base system. Then total initial costs of three options increase by 32, 19, and 54%, respectively. As results, the payback periods of three options are 7, 3, and 44 years. Therefore, it can be concluded that RMA with
subcooler and microturbine intake air precooler is an economically feasible solution among the options investigated. It should be noted that these results can be varied depending upon geological and economical parameters such as the location, fuel cost variation, and unit cost variations.
3. Conclusions Microturbines have emerged as one of new on-site power generation technologies to provide a reliable power supply against the increasingly susceptible electric grid. The overall energy efficiency of the system using microturbines as its power supply can be improved, if the waste heat from microturbines is recovered by an absorption chiller. However, the implementation of the absorption chiller has been slowed since toxicity management is required for the NH3/H2O absorption chiller and the cooling tower is required for the H2O/LiBr absorption chiller due to the temperature limitations of an absorber. To resolve this issue, the implementation of an air-cooled H2O/LiBr absorption chiller was investigated. Based on the modeling results, by adjusting the evaporating temperature according to the ambient temperature, safe marginal concentration differences from the crystallization line could be ensured, thus Table A1 Comparison of air-cooled and water-cooled absorption systems at 35 8C ambient temperature Condenser/absorber
Air-cooled
Water-cooled
Evaporating temperature (8C) Evaporating pressure (kPa) Condenser/absorber temperature (8C) Condensing pressure (kPa) Concentration of rich solution (%) Marginal concentration from the crystallization line (%) COP
5 0.87 45
10 1.2 45
5 0.87 31.5
10 1.2 31.5
9.6 65
9.6 62
4.6 58
4.6 55
1
4
8
0.74
0.75
0.79
11 0.82
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Fig. A1. Comparison of cycle diagrams for air-cooled and water-cooled H2O/LiBr absorption systems. (a) Air-cooled absorption system. (b) Water-cooled absorption system.
enabling the use of an air-cooled H2O/LiBr absorption chiller. The performance potential of a refrigeration system integrated with a microturbine and an air-cooled absorption chiller was investigated theoretically. Once the absorption chiller is integrated into the refrigeration system with a microturbine, the assistance by the absorption chiller enhances the refrigeration cycle efficiency, especially at higher ambient temperatures, by as much as 25 and 5% for
RMA with subcooler and precooler, respectively. Increased efficiency of the refrigeration cycle reduces the required microturbine size, which enhances the part load efficiency especially for the lower ambient temperature. With increased microturbine efficiency, RMA with subcooler, RMA with subcooler and microturbine intake air precooler, and RMA with condenser air precooler can reduce the annual energy consumption by 12, 19, and 3%, respectively, compared to the refrigeration system
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Table A2 Concentration of rich solution vs. ambient temperature (a) Air-cooled system Ambient temperature (8C) Evaporating temperature (8C) Evaporating pressure (kPa) Condensing/absorber temperature (8C) Condensing pressure (kPa) Concentration of rich solution (%) Marginal concentration from the crystallization line (%) (b) Water-cooled system Ambient temperature (8C) Evaporating temperature (8C) Evaporating pressure (kPa) Condensing/absorber temperature (8C) Condensing pressure (kPa) Concentration of rich solution (%) Marginal concentration from the crystallization line (%)
7.2 2.0 0.71 17.2 1.96 50.8 15.2
12.8 3.6 0.79 22.8 2.77 53.5 12.5
18.3 5.2 0.88 28.3 3.84 55.8 10.2
23.9 6.8 0.99 33.9 5.28 58.0 8.0
29.4 8.4 1.10 39.4 7.14 60.0 6.0
35.0 10.0 1.23 45.0 9.57 61.9 4.1
40.6 11.6 1.37 50.6 12.72 63.8 2.2
7.2 3 0.76 9.9 1.22 41.4 24.6
12.8 3 0.76 14.4 1.64 47.5 18.5
18.3 3 0.76 18.6 2.15 52.1 14.9
23.9 3 0.76 23.0 2.80 54.0 12.0
29.4 3 0.76 27.2 3.61 56.6 9.4
35.0 3 0.76 31.5 4.64 59.0 7.0
40.6 3 0.76 35.9 5.91 61.4 5.6
operating with power supplied by the microturbine but without any waste heat utilization from the microturbine. Therefore, RMA with subcooler and microturbine intake air precooler has the best potential of energy savings. The payback period of RMA with subcooler and microturbine intake air precooler is estimated in 3 years, which facilitates it as an economically feasible solution among the options investigated.
Appendix A
A.1. Comparison of air-cooled and water-cooler absorption systems at 35 8C ambient temperature A significant disadvantage of the conventional H2O/LiBr absorption system compared to the NH3/H2O absorption system is the limitation of the absorber temperature due to crystallization. The crystallization limitation forces the H2O/LiBr absorption system to use a cooling tower instead of an air-cooled absorber in certain applications. As an example, two H2O/LiBr absorption systems with air-cooled and water-cooled condensers/absorbers operating at 35 8C are compared (Table A1). If the absorber releases heat at 45 8C to the ambient air of 35 8C with an evaporation temperature of 10 8C (corresponding saturation pressure ¼ 1.2 kPa), the concentration of poor and rich solution in the absorber becomes 58 and 62%, respectively. Then the cycle becomes 11-2-7-5-4-10 (Fig. A1(a)). Therefore, the concentration of the LiBr rich solution is 4% below the crystallization line. If the evaporation temperature is 5 8C (corresponding saturation pressure ¼ 0.9 kPa), the concentration of poor and rich solution in the absorber becomes 61 and 65%, respectively. Then the cycle becomes 11a-2a-7a-5a-4a-10a (Fig. A1(a)). Then, the concentration of the LiBr rich solution is only 1% below the crystallization
line. Therefore, the conventional H2O/LiBr absorption system cannot use the air-cooled absorber for the moderate evaporation temperature around 5 8C or less at high ambient temperature due to potential crystallization. If the evaporation temperature is higher than 10 8C while the absorber releases heat at 45 8C to the ambient air of 35 8C, the aircooled absorber and condenser can be used for high temperature applications such as the subcooling of the condenser outlet of a refrigeration system or the precooling of ambient air. Since the cooling water temperature from the cooling tower is assumed to be 5.6 K higher than the wet-bulb temperature, concentrations and temperatures of the condenser and absorber of the water-cooled system are lower than those of the air-cooled system. It is assumed that the same type of evaporator used for air-cooled system is also used in the water-cooled system. If the absorber can release heat at 31.4 8C to the 29.4 8C cooling water and the evaporation temperature is 10 8C, the concentrations of poor and rich solutions in the absorber become 51 and 55%, respectively. Then the cycle becomes 11-2-7-5-4-10 (Fig. A1(b)). Therefore, the concentration of LiBr rich solution is 11% below the crystallization line. If the evaporation temperature is 5 8C, the concentrations of poor and rich solutions in the absorber become 54 and 58%, respectively. Then the cycle becomes 11b-2b-7b-5b-4b-10b (Fig. A1(b)). Therefore, the concentration of LiBr rich solution is 8% below the crystallization line. A.2. Comparison of air-cooled and water-cooler absorption systems for wide range of ambient temperatures In this study, the average bin temperature distributions of four regions of the US [15] are used. The average operating hours occur in a temperature range between 29.4 and 40.6 8C. Therefore, the system behavior at ambient
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temperatures for this temperature range should be accounted for. As the ambient temperature decreases, the condenser outlet temperature of the vapor compression system decreases, thus reducing the effective temperature difference between the subcooled refrigerant of the vapor compression system and the evaporating refrigerant of the absorption chiller in the subcooler. Therefore, a lower evaporating temperature of the absorption chiller is desirable when the ambient temperature decreases. Since the conventional H2O/LiBr absorption system can be used only above the freezing temperature of water, the minimum evaporator temperature is assumed to be 2 8C. However, the evaporating temperature should be determined in consideration of the marginal concentration from the crystallization line. When air-cooled heat exchangers were used, the absorber and condenser temperatures were assumed to be 10 K higher than the ambient temperature to release heat to the air. The concentration of the rich solution approaches the crystallization line as the ambient temperature increases (Fig. A1). Thus, the margin between the crystallization line and the concentration of the rich solution is reduced as the evaporating temperature decreases. If the maximum ambient temperature is assumed to be 35 8C, the minimum evaporating temperature is 9 8C with 0.7% marginal concentration from the crystallization line. If 1% marginal concentration is used, the minimum evaporating temperature becomes 10 8C. A concentration difference less than 1% is not considered practical. However, if the evaporating temperature is adjusted according to the ambient temperature, the marginal concentration can be increased (Table A2). The marginal concentration differences of Table A2 are considered manageable. When the water-cooled system is used, it is assumed that the evaporating temperature is maintained at 3 8C with an approach temperature of 2 K to prevent freezing (corresponding saturation pressure ¼ 0.8 kPa) independent of the ambient temperature. The absorption temperature increases as the ambient temperature increases, as does the concentration of rich solution (Table A2). The minimum marginal concentration from the crystallization line is 6% when the ambient temperature is at its maximum, 40.6 8C. As discussed above, the water-cooled system has a larger marginal concentration difference to the crystallization line than the air-cooled system since the condenser/absorber temperature is lower than that of the air-cooled system. For
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the same reason, the pressure ratio of the water-cooled system is only half that of the air-cooled system, resulting in a 7–9% higher capacity and COP (Table A1). However, the use of a cooling tower has many disadvantages such as additional first cost, installation, maintenance, and operating cost.
References [1] UNFCC, United Nations Framework Convention on Climate Change Website, http://www.unfccc.org/; 2002. [2] Energy Information Administration (EIA), Annu Energy Rev (2001). [3] R. Sweetser, Technical, economic, and regulatory challenges to installing eight giga watts of new CHP for building systems by the year 2020, ASHRAE Trans 108 (Pt 2) (2002) HI-02-5-1. [4] Gas Research Institute, Absorption Chiller Products, URL, http://www.gri.org/gas_cooling; 2003. [5] Yazaki Energy Systems, Inc, URL, http://www.yazaki.com; 2003. [6] Robur Corp, URL, http://www.robur.com; 2003. [7] Engineering Equation Solver (EES), F-Chart Software; 2002. [8] J. Judge, Y. Hwang, R. Radermacher, Results of two drop-in replacement refrigerants for HCFC-22, Proceedings of 19th International Congress of Refrigeration, vol. IVb, The Hague, The Netherlands, 1995, pp. 1168–1175. [9] Capstone Turbine Corp, URL, http://www.microturbine.com; 2003. [10] Honeywell, URL, http://www.honeywell.com; 2003. [11] IR PowerWorks, URL, http://www.irpowerwork.com; 2003. [12] M. Cowie, X. Liao, R. Radermacher, Second generation integrated microturbine, absorption chiller, and solid wheel desiccant system, Proceedings of International Congress of Refrigeration, 2003. [13] P. Fairchild, S. Labinov, A. Zaltash, D. Rizy, Experimental and theoretical study of microturbine-based BCHP system, Proceedings of 2001 ASME IMECE, 2001, AES-23622. [14] ASHRAE, ASHRAE handbook-systems and equipment, 2000, chapter 34. [15] USAF, Engineering Weather Data, Departments of the Air Force, Army, and the Navy, AFP 88-29; 1978. [16] ASHRAE, ASHRAE handbook-fundamentals, 1997, chapter 30. [17] J. Cuttica, W. Worek, University combined heat and power systems, ASHRAE Trans 108 (Pt 2) (2002) HI-02-5-2. [18] EPRI Report EPRI TR-100537, CFCs Electric Chillers (1992) p. 7-4. [19] R. Brandon, C. Snoek, Microturbine cogeneration, ASHRAE Trans 106 (Pt 1) (2003).
Potential energy benefits of integrated refrigeration system with microturbine and absorption chiller Yunho Hwang* Department of Mechanical Engineering, Center for Environmental Energy Engineering, University of Maryland, Martin Hall Building, Rm 3135, College Park, MD 20742-3035, USA Received 14 April 2003; received in revised form 8 September 2003; accepted 7 January 2004
Abstract This paper presents and analyzes the performance potential of a refrigeration system that is integrated with a microturbine and an absorption chiller (RMA). The waste heat from the microturbine operates the absorption chiller, which provides additional cooling. This additional cooling capacity can be utilized either to subcool the liquid exiting the condenser of the refrigeration system or to precool the air entering the condenser in the refrigeration system. Moreover, any surplus cooling capacity not utilized in the subcooler can be utilized to precool the microturbine intake air. The additional assistance to the refrigeration system enhances the efficiency of the refrigeration cycle, which in turn reduces the required microturbine size. The smaller size of the microturbine enhances the part load efficiency, especially in lower ambient temperatures. With increased microturbine efficiency, RMA with subcooler, RMA with subcooler and microturbine intake air precooler, and RMA with condenser air precooler can reduce the annual energy consumption by 12, 19, and 3%, respectively, as compared to a refrigeration system operating without any waste heat utilization from the microturbine. Therefore, RMA with subcooler and microturbine intake air precooler has the best potential of energy savings. The payback period of RMA with subcooler and microturbine intake air precooler is estimated in 3 years, which facilitates it as an economically feasible solution among the options investigated. q 2004 Published by Elsevier Ltd and IIR. Keywords: Design; Refrigerating system; Cogeneration; Generation; Electricity; Absorption system; Performance
Syste`me frigorifique muni d’une microturbine et d’un refroidisseur a` absorption: avantages e´nerge´tiques potentielles Mots-cle´s: Conception; Syste`me frigorifique; Coge´ne´ration; Ge´ne´ration; Electricite´; Syste`me a` absorption; Performance
1. Introduction 1.1. Environmentally friendly energy use Greenhouse warming occurs when carbon dioxide, released mostly from the burning of fossil fuels (oil, natural * Corresponding author. Tel.: þ1-301-405-5247; fax: þ 1-301405-2025. E-mail address: [email protected] (Y. Hwang).
gas and coal), and other gases such as methane, nitrous oxide, ozone, CFCs, HCFCs, and water vapor, build up in the atmosphere, increasing the Earth’s natural greenhouse effect. This buildup is a concern that has resulted in the Kyoto Protocol [1], which sets greenhouse gas emission objectives for each developed country. The Kyoto Protocol aims at a 5% reduction in greenhouse gas emissions from 1990 levels over the 2008–2012 period. Policies to reduce global warming require industries to develop technologies that will reduce energy consumption and increase energy efficiency.
0140-7007/$ - see front matter q 2004 Published by Elsevier Ltd and IIR. doi:10.1016/j.ijrefrig.2004.01.007
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Nomenclature CEEE Center for Environmental Energy Engineering COP Coefficients of performance EES Engineering Equation Solver EPRI Electric Power Research Institute Phigh High pressure level in the absorption cycle Plow Low pressure level in the absorption cycle PR Pressure ratio Qmt;precooler Cooling capacity used to precool the microturbine intake air Qsubcool Cooling capacity available from the subcooler Qsurplus Surplus cooling capacity available after the subcooler is satisfied RMA Refrigeration system integrated with microturbine and absorption chiller Rsubcooling Amount of cooling capacity used for the subcooling as compared to the maximum cooling capacity available from the absorption chiller Tamb Ambient temperature hise Isentropic efficiency hvol Volumetric efficiency
1.2. Demands for on-site power generation The electricity consumption in the US during 1990s has increased by approximately 2% annually [2]. However, the electric grid is becoming increasingly susceptible to service interruptions due to peak demands. Grid reliability becomes a critical issue in a deregulated market as clearly illustrated by the recent power outages in California. These power outages have drawn attention to on-site power generation. Increased power demands by the cooling systems and IT systems of the existing old buildings, which may have only limited power supply also requires on-site power generation in addition to the existing power supply. Therefore, the onsite power generation becomes more important, not only to ease Electric Peak load but also as means to satisfy new demands independently from the grid.
1.3. On-site power generation For on-site power generation, engines and combustion turbines have been available for many years in a wide range of capacities. In the late 1990s, microturbines and fuel cells emerged as new on-site power generation technologies, targeted especially for small capacity ranges between 1 and 400 kW [3]. Microturbines are small-scale electric generators that are based on the Brayton cycle whose electrical output ranges from 30 to 500 kW. Recent successful development of microturbines enables commercialization by achieving maintenance intervals longer than 8000 h in advance of fuel cells. Microturbines customarily use natural gas as their energy source. High efficiencies of 27 – 30% are reached by using recuperation.
2. Refrigeration system integrated with microturbine and absorption chiller Similar to the integration of the cooling, heating, and power generation to commercial buildings, a refrigeration system can be integrated with a power generation system (microturbine) and an absorption chiller. Hereafter, such a system is referred to as RMA. Electricity in an RMA is provided by the microturbine. The waste heat from the microturbine is utilized to operate the absorption chiller. Then, the additional cooling capacity by the absorption chiller can be utilized in the refrigeration system in the following ways: † Direct refrigeration (2 20 8C) which requires an NH3/H2O absorption chiller † Subcooling the condenser outlet refrigerant of the refrigeration system † Precooling the air entering the microturbine † Precooling the air entering the condenser of the refrigeration system In this paper the latter three options were investigated in detail in anticipation of integrating an air-cooled H2O/LiBr absorption chiller, as will be discussed later. An energy flow diagram and a schematic diagram of an RMA are found in Figs. 1 and 2. As illustrated, the microturbine produces electricity; its exhaust, ranging from 232 to 260 8C drives the absorption chiller. The electricity produced by the microturbine is used to drive the vapor compression refrigeration system, and the chilled water produced by the absorption chiller is used to: further subcool the refrigerant from the condenser (extended solid line from the absorption chiller in Fig. 2); first subcool the refrigerant from the condenser and precool the microturbine intake air with any surplus capacity (extended dashed line from the
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Fig. 1. Energy flow in RMA.
subcooler in Fig. 2) and; precool the cooling air entering the condenser (extended dashed line from the absorption chiller in Fig. 2). The design aspects and performance potential of these three RMAs were theoretically investigated in detail as described next. 2.1. Absorption chiller An absorption chiller is a piece of heat-operated refrigeration equipment, whose cooling capacity is produced using waste heat instead of electricity as the
primary energy source. This ability provides an energy saving opportunity in cases where the cooling demand coincides with surplus heat. The cooling capacity of the current study ranges from 10 to 70 kW. The Gas Research Institute lists eight major manufacturers of absorption chillers in the US [4]. However, only two manufacturers produce absorption chillers in the capacity range of interest herein. One manufacturer produces 4 – 18 kW capacity absorption chillers with cooling tower [5] and the other manufacture produces 4–25 kW capacity absorption chillers with air-cooled condenser/absorber [6].
Fig. 2. Schematic diagram of RMA.
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2.1.1. Working fluids In the absorption cycle, H2O/LiBr and NH3/H2O are the two preferred refrigerant- and absorbent-pairs. While H2O/LiBr absorption systems operate at vacuum pressures between 0.8 and 20 kPa, NH3/H2O absorption systems operate at much higher pressures, between 400 and 2500 kPa. A disadvantage of H2O/LiBr is its narrow solution field, which is limited by crystallization, resulting in an absorber temperature limitation of approximately 40 8C thus requiring a cooling tower. NH3/H2O can operate at higher temperatures but the toxicity of NH3 requires careful attention to system design, installation, and maintenance. H2O/LiBr is used for small to large capacity applications (residential to industrial installations from 10 to 26,000 kW) while NH3/H2O is used for either very small (small refrigerator and residential application less than 70 kW) or large capacity installations. The cooling capacity of the current study ranges from 10 to 70 kW. However, the results generated in this study can be scaled to any capacity. Since H2O/LiBr is used in the capacity range of interest herein and also non-toxic, it was chosen as the working fluid in this study. 2.1.2. Modeling of single-stage H2O/LiBr absorption chiller A single-stage H 2O/LiBr absorption system was modeled using the computer software Engineering Equation Solver [7]. The following assumptions were used in the model: † Evaporator: † Refrigerant is pure water. † Stream exiting the evaporator is saturated vapor. † Pressure level is Plow : † Condenser: † Refrigerant is pure water. † Saturated liquid leaves the condenser. † Air-cooled condenser is used. The approach temperature of the condenser varies depending upon the performance of the condenser and boundary conditions. To make an economically feasible assumption on the performance of the condenser, the approach temperature of the condenser is based on the approach temperature of the commercially available air-cooled condenser, approximately 10 K [8]. † Pressure level is Phigh as determined by the condenser temperature. † Absorber: † Saturated liquid leaves the absorber at the condensing temperature. † Air-cooled absorber is used. Approach temperature of the absorber is 10 K as described for the condenser. The concentration difference between the rich and poor streams is 4%. † Pressure level is Plow as determined through the evaporator temperature.
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† Solution pump: † Efficiency is 100%. † Saturated liquid enters the pump from the absorber. † Pressure level increases from Plow to Phigh : † Solution heat exchanger: † Exiting rich solution does not enter the two-phase region (maximum enthalpy is for saturated liquid). † Entering poor solution is saturated liquid (at generator conditions). † An approach temperature is used in the heat transfer between the exiting poor solution and the entering rich solution ðDT ¼ 10 KÞ: (Here, ‘poor solution’ refers to a solution with a relatively low mass fraction of sorbent (LiBr) and accordingly a relatively high mass fraction of refrigerant (H2O); ‘rich solution’ refers to a solution with a relatively high mass fraction of sorbent and low mass fraction of refrigerant.) † Pressure level is Phigh : † Generator: † The generator heat is supplied from the microturbine waste heat. † Saturated water vapor leaves the generator at the same temperature as the entering poor solution. † Saturated liquid leaves the generator as the rich solution. † Pressure level is Phigh : † Operating conditions: † For water-cooled system, the cooling water temperature from the cooling tower is assumed to be 5.6 K higher than the wet-bulb temperature that is determined by the ambient temperature and relative humidity. † The relative humidity of the ambient air is constant at 40% RH for all ambient temperatures. 2.2. Microturbine There are only limited numbers of microturbine manufactures in the US. These manufactures produce 30 – 250 kW capacity microturbines [9– 11]. The efficiencies of these microturbines are 26 – 28% at ISO conditions (15 8C at sea level) based on natural gas low heating value. Three important characteristics of the microturbine performance should be considered: the dependence of the microturbine efficiency on the ambient temperature, the dependence of the power factor on the ambient temperature, and the part load efficiency degradation. Here, the power factor is defined as the ratio of the microturbine power output to its rated power output. Characteristics of typical microturbines are illustrated in Fig. 3 [12,13]. As the ambient temperature increases the efficiency and power factor of the microturbine both decrease (Fig. 3(a)). The main contribution to this performance degradation is the reduction in the intake air and fuel flow rate as the air and fuel density decrease. The
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Fig. 3. Characteristics of start-of-the-art microturbine. (a) Performance of microturbine vs. ambient temperature [12]. (b) Part load efficiency of microturbine vs. part load factor [13].
power factor decreases as the ambient temperature increases greater than 26.7 8C although the turbine speed and fuel gas compressor speed are maintained at their maximum. However, the power factor can be maintained by adjusting the turbine speed and fuel gas compressor speed when the ambient temperature decreases below 26.7 8C. It should be noted that this performance dependence on the ambient temperature can be changed through a different design approach. During the partial load operation, the turbine speed and fuel supply are adjusted. With a significant turbine speed reduction, the microturbine efficiency decreases during the partial load operation (Fig. 3(b)). The efficiency during the partial load operation slowly degrades by 10% when the partial load factor decreases to 0.5. However, the partial load efficiency degradation becomes significant when the partial load factor becomes less than
0.5. Therefore, it is important to size the microturbine output such that the partial load factor is maintained greater than 0.5. Since only limited information on these microturbine characteristics is available from the open literature, the microturbine characteristics illustrated in Fig. 3 are used in this study. To recover the waste heat from the microturbine, the exhaust pipe of the microturbine is extended to the absorption chiller and the exhaust gas passes through the generator of the absorption chiller. Then this additional process causes a back pressure to the exhaust (it was approximately 1 kPa [12]) and can be compensated by either modifying the combustion fan of the generator or adding additional exhaust fan. The waste heat flow from the microturbine to the absorption chiller was assumed based on Cowie et al. [12] as indicated below.
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Fig. 4. COP and COP improvement of refrigeration system with subcooler over base refrigeration system.
38% of the fuel energy from the microturbine exhaust gas is delivered to the absorption chiller. Of this delivered heat, only 69% is utilized by the absorption chiller. It should be noted that this assumption might be different for different systems. 2.3. Performance improvement of refrigeration system with absorption chiller The performance of the RMA is analyzed in two respects: refrigeration cycle efficiency enhancement and overall system energy use reduction. First, the enhancement of the refrigeration cycle efficiency is analyzed for three options. It should be noted that only the results for the aircooled absorption system case are presented here since similar performance enhancement was observed for the water-cooled absorption system case. 2.3.1. Refrigeration system with condenser subcooler In the subcooler cycle using an absorption chiller, the subcooler cools the refrigerant exiting the condenser and provides lower vapor quality refrigerant to the evaporator, which then provides a larger cooling capacity for the same refrigerant mass flow rate of the original vapor compression system. The performance of the RMA with subcooler using R22 as its refrigerant was modeled using EES. The following assumptions were used in the model: † Evaporator: † Refrigeration capacity is 100 kW. Refrigerant mass flow rate is adjusted for each case. † Evaporation temperature is 10 K lower than the refrigerated air temperature. † Superheat of vapor is 5 K exiting the evaporator. † Pressure drop is 50 kPa. † Condenser: † Condensing temperature is 10 K higher than the inlet air temperature.
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† Degree of subcooling at condenser outlet is 5 K. † Pressure drop is 50 kPa. † Compressor: † Isentropic and volumetric efficiencies (hise ; hvol ) change depending upon the pressure ratio (PR) as given in the following equations [14]. † hise ¼ 0:85 2 0:0467PR; hvol ¼ 1:08 2 0:04PR † Subcooler: † H 2 O/LiBr absorption chiller provides the additional subcooling. † Absorption chiller works only for an ambient temperature between 7.2 and 40.6 8C. † The subcooled liquid refrigerant and chilled water exchange heat in counterflow configuration. The approach temperature is 2 K. † The approach temperature limits the degree of subcooling. † The subcooler does not affect the pressures of the evaporator and condenser. † When the air-cooled absorption chiller is used, the evaporating temperature of the absorption chiller is controlled dynamically according to the ambient temperature (Table A2). When the water-cooled absorption chiller is used, the evaporating temperature is maintained at 3 8C independent of the ambient temperature. † Fan motor: † Airflow rate through the evaporator is 0.0537 m3/s for 1 kW refrigeration capacity. † Power input to the evaporator and condenser fan motor is 775 W for 1 m3/s airflow rate. † Evaporator motor is located in the cold air stream and works as thermal load.
The COP and COP improvement of the refrigeration system with subcooler by the air-cooled absorption chiller over the base refrigeration system at various bin temperatures is depicted in Fig. 4. The COP decreases but the COP improvement increases as the ambient temperature increases (Fig. 4). The COP improvement starts at 7.2 8C since the absorption chiller starts from this temperature bin. The maximum COP improvement is 4 and 25% at 7.2 and 40.6 8C, respectively. The subcooling energy use ratio ðRsubcooling Þ of the refrigeration system with condenser subcooler is illustrated in Fig. 5(a). Here, the Rsubcooling is defined as the actual amount of cooling energy used for the subcooling as compared to the maximum cooling capacity available from the absorption chiller. The cooling capacity by the absorption chiller is used only up to 52% (Fig. 5(a)). In other words, the subcooling capacity is great enough to reach the minimum possible subcooler outlet temperature for the entire range of ambient temperatures. This results in the limitation of COP improvement.
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Fig. 5. Use of surplus cooling in refrigeration system with subcooler. (a) Subcooling energy use ratio of cooling from air-cooled absorption chiller. (b) Surplus cooling available after subcooling and cooling capacity used for microturbine intake air precooling.
2.3.2. Refrigeration system with subcooler and microturbine intake air precooler Since the utilization of the subcooling capacity from the subcooler is limited, the surplus cooling capacity by the subcooler can be employed to precool the intake air entering the microturbine. The surplus cooling capacity ðQsurplus Þ and the amount of the cooling capacity used to precool the microturbine intake air ðQmt;precooler Þ are compared to the subcooling capacity available from the subcooler ðQsubcool Þ (Fig. 5(b)). The surplus cooling capacity decreases as the ambient temperature increases (Fig. 5(b)). If this surplus cooling is utilized to precool the intake air entering the microturbine, the microturbine efficiency can be improved. For example, the microturbine efficiency improvement is 13% if all the surplus cooling capacity is used at 40.6 8C. However, heat transfer limits the utilization of the surplus cooling in most cases because the intake air temperature cannot be lower than the subcooler temperature. To account for this heat transfer limitation, the ambient air temperature
was assumed to be higher than that of the subcooler temperature by 2 K. The cooling capacity used to precool the intake air increases as the ambient temperature increases but its utilization is clearly limited especially at low ambient temperature (Fig. 5(b)). This heat transfer limitation results in the limitation of the microturbine efficiency improvement. Moreover, it means that the current waste heat recovery is large enough and more waste heat recovery can not contribute to the further enhancement. 2.3.3. Refrigeration system with condenser air precooler In the precooler cycle using an absorption chiller, the precooler cools the air entering the condenser and supplies the air at reduced temperature. The lower air temperature entering the condenser results in a lower pressure ratio and reduced compressor power consumption. The performance of the RMA with precooler using R22 was modeled using EES. The same assumptions used for the evaporator, condenser and compressor of the cycle with subcooler were applied in the
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Fig. 6. COP improvement and condenser intake air temperature reduction of refrigeration system with precooler over base refrigeration system.
model. The following assumptions for the precooler were used in addition to the previous assumptions: † Precooler: † Airflow rate is the same as that of the conventional cycle to increase the air temperature by 5 K across the condenser. † H2O/LiBr absorption chiller provides the precooling. † Absorption chiller works only for ambient temperature between 7.2 and 40.6 8C. † Precooler does not affect the evaporation pressure. † Refrigerant and air heat exchange occurs in crosscounterflow configuration. The approach temperature is 2 K. † The same assumption for the evaporating temperature of the absorption chiller used for the subcooler cycle is used. The COP improvement and the condenser intake air temperature reduction of the refrigeration system with precooler over the base refrigeration system at various bin temperatures is depicted in Fig. 6. The COP improvement increases as the ambient temperature increases (Fig. 6). The COP improvement begins at 7.2 8C since the absorption chiller can start from this temperature bin. The maximum COP improvement is approximately 3 and 5% at 7.2 and 40.6 8C, respectively. This COP improvement is induced by the condenser inlet air temperature reduction as much as 0.9– 1.5 K (Fig. 6). The COP improvement is not limited by heat transfer in this case. 2.4. Annual energy consumption of RMA To analyze the overall performance potential of the RMA over the entire year, the annual energy consumption of RMA options were investigated as described next.
2.4.1. Comparison of annual energy consumption of RMA options The annual energy consumption of each of the RMA options operating alone with power supplied by the microturbine was investigated to compare the potential energy saving of each. The following assumptions were used in the calculation of the annual energy consumption by the RMA. † Bin temperature distribution and annual bin hours are based on US averages, which are the average of four US regional bin hours (Midwest, Southwest, Northeast, and Southeast) [15]. † The indoor temperature and refrigeration load (100 kW) are constant throughout the year. † Absorption chiller is used only when the ambient bin temperature is between 7.2 and 40.6 8C. Otherwise, only the base refrigeration system is used. † The performance of the microturbine is based on the results described in Section 2.2. † Efficiency of the refrigeration system assisted by the absorption chiller is based on the results described in Section 2.3. † If a surplus cooling capacity after the subcooling is available, it is used to precool the intake air to the microturbine. The calculated annual power consumption of RMA options using 100% microturbine as its power source are compared to each other in Table 1. RMA with subcooler, RMA with subcooler and microturbine intake air precooler, and RMA with condenser air precooler can reduce the annual energy consumption by 12, 19, and 3%, respectively, compared to the refrigeration system operating alone with power supplied by the microturbine without the integration of the absorption chiller. These energy savings are due to the refrigeration cycle efficiency enhancement, which is
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Tamb (8C)
Bin (h)
COP Base system
29.4 88 6.16 23.9 350 5.14 1.7 788 4.34 7.2 964 3.70 12.8 1139 3.16 18.3 1489 2.71 23.9 2102 2.32 29.4 1314 1.99 35.0 350 1.69 40.6 175 1.43 Annual energy consumption (kWh) Annual energy consumption ratio
Energy consumption (kWh) System with absorption chiller
Microturbine
Sub-cooler
Pre-cooler
Base system
System with subcooler
System with subcooler, MT precooler
System with condenser precooler
6.16 5.14 4.34 3.84 3.36 2.96 2.60 2.30 2.03 1.78
6.16 5.14 4.34 3.79 3.25 2.79 2.40 2.06 1.76 1.49
7599 34,668 88,710 122,780 164,755 244,969 396,720 285,095 88,727 54,819 1,488,844 1.00
7025 32,177 82,689 111,922 148,320 217,237 345,345 242,960 73,653 44,010 1,305,337 0.88
6739 30,943 79,738 107,139 140,557 203,600 319,419 222,255 66,524 37,770 1,214,685 0.81
7472 34,117 87,373 118,914 159,464 236,818 382,890 274,564 85,174 52,358 1,439,145 0.97
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Table 1 Comparison of energy consumption for US average bin hours (microturbine efficiency 26% at ISO condition)
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825
Fig. 7. Annual energy consumption ratio vs. microturbine efficiency.
especially effective at higher ambient temperatures. Increased efficiency of the refrigeration cycle reduces the required microturbine size by 20, 29, and 5%, respectively, for each option. The smaller size of the microturbine further enhances the partial load efficiency, especially for the lower ambient temperature. To further investigate the annual energy savings potential of each option in a typical US location, the energy consumption estimation was repeated for four US cities (Chicago, Los Angeles, Miami, and Seattle) and based on annual bin data for those cities [16]. Table 2 provides a comparison of annual energy consumption ratio between options as compared to the base system for four cities and US average. As shown in Table 2, the energy consumption ratio of RMA with subcooler and RMA with condenser air precooler is similar to that for the US average within a 2% variation, while the energy consumption ratio of the RMA with subcooler and microturbine intake air precooler is varied up to 5% from that for the US average. However, the rank of energy savings potential is maintained. It should be further noted, that the energy consumption ratio of each option at Miami is similar to that of the US average case.
2.4.2. Effects of microturbine efficiency on annual energy consumption of RMA As described earlier, the efficiencies of microturbines referred from the open literature are 26 – 28% at ISO conditions (15 8C at sea level). To investigate the effect of the microturbine efficiency on the annual energy consumption of each RMA option, cases of three microturbine efficiencies (26, 28 and 30%) are evaluated as illustrated in Fig. 7. When the microturbine efficiency increases the annual energy consumption of all cases including the base system decreases. Moreover, the trend of the energy saving of each option is remained the same for three efficiencies examined. Therefore, the annual energy saving of each option is almost similar for three microturbine efficiencies. 2.4.3. Comparison of annual energy consumption of RMA with air-cooled and water-cooled absorption chiller The annual energy consumption of the RMA with subcooler and microturbine intake air precooler by the water-cooled absorption system is only 1% less than or the same as that by the air-cooled absorption system. Therefore, the air-cooled absorption system is recommended since the
Table 2 Comparison of energy consumption for four US cities (microturbine efficiency 26% at ISO condition) City
Base system
System with subcooler
System with subcooler, MT precooler
System with condenser precooler
Chicago Los Angeles Miami Seattle US Average
1.00 1.00 1.00 1.00 1.00
0.90 0.89 0.87 0.90 0.88
0.84 0.84 0.80 0.86 0.81
0.97 0.97 0.96 0.97 0.97
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Table 3 Economic analysis of RMA options System
Base system
System with subcooler
System with subcooler, MT precooler
System with condenser precooler
Annual fuel cost ($) Annual fuel cost ratio MT size ratio MT cost ($) Absorption þ heat recover cost Total initial cost ($) Initial cost ratio Payback period (yr)
24,119 1.00 1.00 66,057 0 66,057 1.00 –
21,146 0.88 0.80 53,062 34,168 87,230 1.32 7
19,678 0.82 0.71 46,898 31,453 78,351 1.19 3
23,314 0.97 0.95 63,058 38,409 101,467 1.54 44
system design is much simpler and lower maintenance, compared to the water-cooled system, yet consumes a similar amount of energy to the water-cooled system. 2.5. Economic analysis An annual energy saving and an initial cost increase due to the additional heat recovery of each RMA options as compared to the base system were investigated to estimate the payback period of each option. The following assumptions were used in the calculation of the payback period. † Annual energy consumption is based on Table 1. † Fuel cost is based $4.5/106 kJ [17]. † Initial cost includes cost of devices (microturbine, absorption chiller, heat recovery) and their installation cost. † Costs of microturbine and heat recovery device except the absorption chiller are linear to their capacity. Whereas, the cost of absorption chiller per unit cooling capacity increases as the capacity decreases based on the profile described by EPRI [18]. † Microturbine cost is based on $800/kW [19]. † Absorption chiller cost is $440/kW to $490/kW depending upon the capacity [18]. † Heat recovery heat exchanger and subcooler or precooler costs are based on $28/kW [19]. Based on the above assumptions, an annual fuel cost saving, total initial cost increase and a simple payback period were calculated as summarized in Table 3. From the results, the annual energy cost savings of three options are 12, 19, and 3%, respectively, compared to the base system. Since the cost of the microturbine decreases as the energy saving increases, it is expected that the costs of the microturbine of three options are reduced by 20, 29, and 5%. However, the initial cost increases by the absorption chiller, which is approximately 60 – 70% of the microturbine cost in the base system. Then total initial costs of three options increase by 32, 19, and 54%, respectively. As results, the payback periods of three options are 7, 3, and 44 years. Therefore, it can be concluded that RMA with
subcooler and microturbine intake air precooler is an economically feasible solution among the options investigated. It should be noted that these results can be varied depending upon geological and economical parameters such as the location, fuel cost variation, and unit cost variations.
3. Conclusions Microturbines have emerged as one of new on-site power generation technologies to provide a reliable power supply against the increasingly susceptible electric grid. The overall energy efficiency of the system using microturbines as its power supply can be improved, if the waste heat from microturbines is recovered by an absorption chiller. However, the implementation of the absorption chiller has been slowed since toxicity management is required for the NH3/H2O absorption chiller and the cooling tower is required for the H2O/LiBr absorption chiller due to the temperature limitations of an absorber. To resolve this issue, the implementation of an air-cooled H2O/LiBr absorption chiller was investigated. Based on the modeling results, by adjusting the evaporating temperature according to the ambient temperature, safe marginal concentration differences from the crystallization line could be ensured, thus Table A1 Comparison of air-cooled and water-cooled absorption systems at 35 8C ambient temperature Condenser/absorber
Air-cooled
Water-cooled
Evaporating temperature (8C) Evaporating pressure (kPa) Condenser/absorber temperature (8C) Condensing pressure (kPa) Concentration of rich solution (%) Marginal concentration from the crystallization line (%) COP
5 0.87 45
10 1.2 45
5 0.87 31.5
10 1.2 31.5
9.6 65
9.6 62
4.6 58
4.6 55
1
4
8
0.74
0.75
0.79
11 0.82
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Fig. A1. Comparison of cycle diagrams for air-cooled and water-cooled H2O/LiBr absorption systems. (a) Air-cooled absorption system. (b) Water-cooled absorption system.
enabling the use of an air-cooled H2O/LiBr absorption chiller. The performance potential of a refrigeration system integrated with a microturbine and an air-cooled absorption chiller was investigated theoretically. Once the absorption chiller is integrated into the refrigeration system with a microturbine, the assistance by the absorption chiller enhances the refrigeration cycle efficiency, especially at higher ambient temperatures, by as much as 25 and 5% for
RMA with subcooler and precooler, respectively. Increased efficiency of the refrigeration cycle reduces the required microturbine size, which enhances the part load efficiency especially for the lower ambient temperature. With increased microturbine efficiency, RMA with subcooler, RMA with subcooler and microturbine intake air precooler, and RMA with condenser air precooler can reduce the annual energy consumption by 12, 19, and 3%, respectively, compared to the refrigeration system
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Table A2 Concentration of rich solution vs. ambient temperature (a) Air-cooled system Ambient temperature (8C) Evaporating temperature (8C) Evaporating pressure (kPa) Condensing/absorber temperature (8C) Condensing pressure (kPa) Concentration of rich solution (%) Marginal concentration from the crystallization line (%) (b) Water-cooled system Ambient temperature (8C) Evaporating temperature (8C) Evaporating pressure (kPa) Condensing/absorber temperature (8C) Condensing pressure (kPa) Concentration of rich solution (%) Marginal concentration from the crystallization line (%)
7.2 2.0 0.71 17.2 1.96 50.8 15.2
12.8 3.6 0.79 22.8 2.77 53.5 12.5
18.3 5.2 0.88 28.3 3.84 55.8 10.2
23.9 6.8 0.99 33.9 5.28 58.0 8.0
29.4 8.4 1.10 39.4 7.14 60.0 6.0
35.0 10.0 1.23 45.0 9.57 61.9 4.1
40.6 11.6 1.37 50.6 12.72 63.8 2.2
7.2 3 0.76 9.9 1.22 41.4 24.6
12.8 3 0.76 14.4 1.64 47.5 18.5
18.3 3 0.76 18.6 2.15 52.1 14.9
23.9 3 0.76 23.0 2.80 54.0 12.0
29.4 3 0.76 27.2 3.61 56.6 9.4
35.0 3 0.76 31.5 4.64 59.0 7.0
40.6 3 0.76 35.9 5.91 61.4 5.6
operating with power supplied by the microturbine but without any waste heat utilization from the microturbine. Therefore, RMA with subcooler and microturbine intake air precooler has the best potential of energy savings. The payback period of RMA with subcooler and microturbine intake air precooler is estimated in 3 years, which facilitates it as an economically feasible solution among the options investigated.
Appendix A
A.1. Comparison of air-cooled and water-cooler absorption systems at 35 8C ambient temperature A significant disadvantage of the conventional H2O/LiBr absorption system compared to the NH3/H2O absorption system is the limitation of the absorber temperature due to crystallization. The crystallization limitation forces the H2O/LiBr absorption system to use a cooling tower instead of an air-cooled absorber in certain applications. As an example, two H2O/LiBr absorption systems with air-cooled and water-cooled condensers/absorbers operating at 35 8C are compared (Table A1). If the absorber releases heat at 45 8C to the ambient air of 35 8C with an evaporation temperature of 10 8C (corresponding saturation pressure ¼ 1.2 kPa), the concentration of poor and rich solution in the absorber becomes 58 and 62%, respectively. Then the cycle becomes 11-2-7-5-4-10 (Fig. A1(a)). Therefore, the concentration of the LiBr rich solution is 4% below the crystallization line. If the evaporation temperature is 5 8C (corresponding saturation pressure ¼ 0.9 kPa), the concentration of poor and rich solution in the absorber becomes 61 and 65%, respectively. Then the cycle becomes 11a-2a-7a-5a-4a-10a (Fig. A1(a)). Then, the concentration of the LiBr rich solution is only 1% below the crystallization
line. Therefore, the conventional H2O/LiBr absorption system cannot use the air-cooled absorber for the moderate evaporation temperature around 5 8C or less at high ambient temperature due to potential crystallization. If the evaporation temperature is higher than 10 8C while the absorber releases heat at 45 8C to the ambient air of 35 8C, the aircooled absorber and condenser can be used for high temperature applications such as the subcooling of the condenser outlet of a refrigeration system or the precooling of ambient air. Since the cooling water temperature from the cooling tower is assumed to be 5.6 K higher than the wet-bulb temperature, concentrations and temperatures of the condenser and absorber of the water-cooled system are lower than those of the air-cooled system. It is assumed that the same type of evaporator used for air-cooled system is also used in the water-cooled system. If the absorber can release heat at 31.4 8C to the 29.4 8C cooling water and the evaporation temperature is 10 8C, the concentrations of poor and rich solutions in the absorber become 51 and 55%, respectively. Then the cycle becomes 11-2-7-5-4-10 (Fig. A1(b)). Therefore, the concentration of LiBr rich solution is 11% below the crystallization line. If the evaporation temperature is 5 8C, the concentrations of poor and rich solutions in the absorber become 54 and 58%, respectively. Then the cycle becomes 11b-2b-7b-5b-4b-10b (Fig. A1(b)). Therefore, the concentration of LiBr rich solution is 8% below the crystallization line. A.2. Comparison of air-cooled and water-cooler absorption systems for wide range of ambient temperatures In this study, the average bin temperature distributions of four regions of the US [15] are used. The average operating hours occur in a temperature range between 29.4 and 40.6 8C. Therefore, the system behavior at ambient
Y. Hwang / International Journal of Refrigeration 27 (2004) 816–829
temperatures for this temperature range should be accounted for. As the ambient temperature decreases, the condenser outlet temperature of the vapor compression system decreases, thus reducing the effective temperature difference between the subcooled refrigerant of the vapor compression system and the evaporating refrigerant of the absorption chiller in the subcooler. Therefore, a lower evaporating temperature of the absorption chiller is desirable when the ambient temperature decreases. Since the conventional H2O/LiBr absorption system can be used only above the freezing temperature of water, the minimum evaporator temperature is assumed to be 2 8C. However, the evaporating temperature should be determined in consideration of the marginal concentration from the crystallization line. When air-cooled heat exchangers were used, the absorber and condenser temperatures were assumed to be 10 K higher than the ambient temperature to release heat to the air. The concentration of the rich solution approaches the crystallization line as the ambient temperature increases (Fig. A1). Thus, the margin between the crystallization line and the concentration of the rich solution is reduced as the evaporating temperature decreases. If the maximum ambient temperature is assumed to be 35 8C, the minimum evaporating temperature is 9 8C with 0.7% marginal concentration from the crystallization line. If 1% marginal concentration is used, the minimum evaporating temperature becomes 10 8C. A concentration difference less than 1% is not considered practical. However, if the evaporating temperature is adjusted according to the ambient temperature, the marginal concentration can be increased (Table A2). The marginal concentration differences of Table A2 are considered manageable. When the water-cooled system is used, it is assumed that the evaporating temperature is maintained at 3 8C with an approach temperature of 2 K to prevent freezing (corresponding saturation pressure ¼ 0.8 kPa) independent of the ambient temperature. The absorption temperature increases as the ambient temperature increases, as does the concentration of rich solution (Table A2). The minimum marginal concentration from the crystallization line is 6% when the ambient temperature is at its maximum, 40.6 8C. As discussed above, the water-cooled system has a larger marginal concentration difference to the crystallization line than the air-cooled system since the condenser/absorber temperature is lower than that of the air-cooled system. For
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the same reason, the pressure ratio of the water-cooled system is only half that of the air-cooled system, resulting in a 7–9% higher capacity and COP (Table A1). However, the use of a cooling tower has many disadvantages such as additional first cost, installation, maintenance, and operating cost.
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