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Open air–vapor compression refrigeration system for air conditioning and hot water cooled by cool water
Energy Conversion and Management 48 (2007) 2255–2260 www.elsevier.com/locate/enconman
Open air–vapor compression refrigeration system for air conditioning and hot water cooled by cool water Shaobo Hou b
a,b,*
, Huacong Li a, Hefei Zhang
a
a School of Power and Energy, Northwestern Polytechnical University, Xi’an, Shaanxi 710072, PR China College of Engineering, Guangdong Ocean University, East Jiefang Road, No. 40, Xiashan, Zhanjiang, Guangdong 524006, PR China
Received 25 January 2005; received in revised form 22 September 2006; accepted 9 May 2007 Available online 18 June 2007
Abstract This paper presents an open air–vapor compression refrigeration system for air conditioning and hot water cooled by cool water and proves its feasibility through performance simulation. Pinch technology is used in analysis of heat exchange in the surface heat exchanger, and the temperature difference at the pinch point is selected as 6 °C. Its refrigeration depends mainly on both air and vapor, more efficient than a conventional air cycle, and the use of turbo-machinery makes this possible. This system could use the cool in the cool water, which could not be used to cool air directly. Also, the heat rejected from this system could be used to heat cool water to 33– 40 °C. The sensitivity analysis of COP to gc and gt and the simulated results T4, T7, T8, q1, q2 and Wm of the cycle are given. The simulations show that the COP of this system depends mainly on T7, gc and gt and varies with T3 or Twet and that this cycle is feasible in some regions, although the COP is sensitive to the efficiencies of the axial compressor and turbine. The optimum pressure ratio in this system could be lower, and this results in a fewer number of stages of the axial compressor. Adjusting the rotation speed of the axial compressor can easily control the pressure ratio, mass flow rate and the refrigerating capacity. The adoption of this cycle will make the air conditioned room more comfortable and reduce the initial investment cost because of the obtained very low temperature air. Humid air is a perfect working fluid for central air conditioning and no cost to the user. The system is more efficient because of using cool water to cool the air before the turbine. In addition, pinch technology is a good method to analyze the wet air heat exchange with water. Ó 2007 Elsevier Ltd. All rights reserved. Keywords: Turbo-machinery; Air cycle; Air conditioning units; Natural working fluid; Refrigeration; Pinch technology
1. Introduction The outdoor air temperature can reach 38–40 °C in middle China in summer, while the temperature of the water from the water supply system and underground is about 17 °C and that of the refilled underground water could be lower. These waters are too cool for people to bath directly at the hotel, and the bath water is usually heated by a boiler. If these waters were used to cool air for air conditioning * Corresponding author. Address: School of Power and Energy, Northwestern Polytechnical University, Xi’an, Shaanxi 710072, PR China. Tel.: +86 759 2886705. E-mail addresses: [email protected], [email protected] (S. Hou).
0196-8904/$ - see front matter Ó 2007 Elsevier Ltd. All rights reserved. doi:10.1016/j.enconman.2007.05.005
directly, it is not efficient because of the small temperature difference and large amount of consumed water. The air compression refrigeration cycle was studied long ago. Several disadvantages prevented air from being used as a working fluid in refrigeration. These included low volumetric refrigerating effect, which may result in a large compressor, and low COP due to low efficiencies of the compressor and expander. After CFC’s invention in the 1930s, people paid little attention to air compression refrigeration. Recently, as a result of the destruction of the ozonosphere by chlorofluorocarbons (CFC) and the pressure of environmental protection, research upon air refrigeration cycles has received more attention [1–3]. Optimizations of
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S. Hou et al. / Energy Conversion and Management 48 (2007) 2255–2260
Nomenclature B D H P T Twet Wc Wt Wm gc gt ratio
wet air pressure, Pa humidity ratio of wet air, g/kg(d.a.) enthalpy of wet air, kJ/kg(d.a.) pressure, Pa temperature, K or °C wet bulb temperature, K or °C ideal input work of compressor, kJ/kg(d.a.) ideal output work of turbine, kJ/kg(d.a.) practical work consumed by system, kJ/kg(d.a.) efficiency of compressor efficiency of turbine mass flow rate ratio of cooling water to dry air
air cycles are also conducted using finite time thermodynamics (FTT) or entropy generation minimization (EGM) [4–7]. Chen et al. investigated the cooling load versus COP characteristics of a simple [8] and a regenerated [9,10] air refrigeration cycle with heat transfer loss and/or other irreversibilities. Luo et al. [11] optimized the cooling load and the COP of a simple irreversible air refrigeration cycle by searching for the optimum pressure ratio of the compressor and the optimum distribution of heat conductance of the hot and cold side heat exchangers for fixed total heat exchanger inventory. With the development of the aeronautical industry, highly efficient axial compressors and turbines have become a reality. At present, the stagnation isentropic efficiencies of a single stage axial compressor and a turbine can reach 0.88–0.91 [12]. High speed fans have been used in ordinary air conditioning systems nowadays. However, the water vapor in the working fluid was not considered in all the above researches [1–11] on the air compression refrigeration cycle, and the used equipments were a centrifugal compressor and a centripetal turbine, which have lower efficiencies than axial compressors and turbines. The amount of water extracted from high pressure wet air can reach 18–30 g/kg(d.a.), and the amount of latent heat discharged from the vapor condensed is about 45–75 kJ/kg(d.a.), exceeding the sensible heat from the air of 30–50 kJ/kg(d.a.). Hou and Li presented an axial flow air–vapor compression refrigerating system for air conditioning in 1992 [13] in which wet air is the working fluid and an axial compressor and turbine were used, but these have not yet attracted people’s attention so far. Hou and Zhang presented an axial flow air–vapor compression refrigerating system for air conditioning cooled by circulating water in Ref. [14] (2004) in which wet air is the working fluid, an axial compressor and turbine were used and the wet air is cooled by circulating water. The paper proves its feasibility through performance simulation and also indicates its advantages. These include the possibility
R n
gas constant, kJ/kg K exponent
Subscripts air dry air vapor water vapor in moist air s saturated last last time hot hot stream cold cold stream w water
to simplify air conditioning systems, to reduce the amount of initial investment of an air conditioning system and to make air conditioned rooms more comfortable. The aim of this paper is to present an open system, which is an open air–vapor compression refrigeration system for air conditioning and hot water cooled by cool water and its performance from simulation. In this open air–vapor refrigeration cycle, water from the water supply system or underground is used. Thus, we could get a lower wet air temperature before the turbine. In addition, the cool water is heated for the bath. 2. System Representation on enthalpy–entropy coordinates and a circuit diagram of an open air-compression refrigeration system for air conditioning and hot water cooled by cool water are shown in Fig. 1. The outdoor air at 2 is drawn into the atomizing chamber, cooled to saturated air at 3 with some fine water droplets and then compressed by an axial compressor. A flow of compressed air at 4 with higher temperature, T4, and high pressure, P4, is obtained. Then, the compressed air at 4 is cooled to saturated air at 7 with a temperature T7 by cool water/underground water in a surface heat exchanger after
Fig. 1. Representation on enthalpy–entropy coordinates and circuit diagram of an open air-compression refrigeration system for air conditioning and hot water cooled by cool water.
S. Hou et al. / Energy Conversion and Management 48 (2007) 2255–2260
the axial compressor outlet. Some vapor is condensed, and the latent heat of the vapor is discharged from 4 to 7. Then, the saturated air at 7 is expanded and cooled to the cool air at 8 in the turbine. The cool air at 8 is then ducted to the air conditioned rooms. The cool water is heated in the surface heat exchanger. Water injection before the axial compressor aims to decrease both the temperature of the working fluid and the polytropic exponent in the compression process. Thus, we can save some compression work. This method has been used in a jet engine when a fighter plane increases its speed. However, the difference is that what is injected in a jet engine is water, alcohol, etc. [12]. The water vapor in the compressed air can easily be extracted by a surface heat exchanger. With the same temperature, the humidity ratio of the saturated wet air at high pressure P4 is only about P3/P4 of that at pressure P3. The method of using compressed air to acquire dry air has been used in some workshops in southern China. The system above differs from a conventional air cycle system. There are many characteristics in this air–vapor refrigeration circle. Firstly, an axial compressor and a turbine are used in the above system. The characteristics of turbo-machines are large mass flow rate and high efficiency. The other types of compressor and expander have none of the above advantages. Secondly, this refrigeration system intakes precooled wet air with fine water droplets, and some vapor is condensed during the air cooling from 4 to 7. The amount of water extracted from the high pressure wet air can reach 18–30 g/kg(d.a.), and the amount of latent heat discharged from the vapor condensed, about 45–75 kJ/kg(d.a.), exceeds the sensible heat from the air, 30–50 kJ/kg(d.a.). For this reason, the refrigeration load in this air–vapor refrigeration system depends on a combination of the sensible heat of air and the latent heat of vapor. Lastly, the cool from the cooling water was used. Usually, it cannot be used. 3. Performance simulation 3.1. Wet air The humidity ratio of wet air, D, is obtained from D ¼ 621:98
P vap B P vap
ð1Þ
The enthalpy of wet air, H, is calculated from H ¼ 1:006t þ 0:001Dð2501 þ 1:805tÞ
ð2Þ
The adopted relation for water vapor between saturation pressure and saturation temperature, Ps = f(ts), is selected from Ref. [15]. To calculate the saturated temperature of the wet air from the saturated enthalpy, Eqs. (1) and (2) and Ps = f(ts) are used.
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3.2. Axial compressor During the compression process of the wet air, the fine water droplets in the air may evaporate. Because the evaporation of water takes in heat, we can regard the ideal compression process of the wet air in the compressor as a polytropic process. Therefore, we can obtain the ideal work of the compressor per kilogram dry air, Wc, from " # n1 n p4 n ðRair þ 0:001DRvapor ÞT 3 1 Wc ¼ ð3Þ n1 p3 in which n is the polytropic exponent for the compression process. The practical work consumed by the axial compressor is Wc/gc in which gc is the thermal efficiency of the compressor.
3.3. Turbine The saturated air with a pressure of P7 and a temperature of T7 before the turbine has been dehumidifed in the surface heat exchanger by cooling water. At point 7, the amount of vapor included in the saturated air is very small, about P3/P7 of the amount included in saturated air at P3. Thus, the water condensed in the air is in fog. Nevertheless, expansion of the saturated air in the turbine cannot be regarded as an adiabatic expansion of an ideal gas. With the decrease of the wet air pressure in the turbine, the temperature of the wet air decreases, and some heat is discharged during the condensation of some water vapor. The heat discharged may cause increases in both the temperature of the turbine outlet and the work done in the expansion. For this problem, we can imagine that no phase change exists and that there is some heat added to the wet air during the expansion process when we calculate the work done by the expansion process. According to the above assumption, this problem can be simplified to a problem of the polytropic expansion of an ideal gas. Consequently, we can obtain the ideal work done by the expansion, Wt, through iteration and then obtain the real work generated by the turbine and the temperature of the turbine outlet. The following are the steps to calculate Wt and T8. 1. Determine Ps7 from T7, D7 from Ps7 and P7, and H7 from D7 and T7. 2. Get gas constant R for the saturated air at 7 by using the formula: R ¼ 0:001ð287 þ 0:461D7 Þ:
ð4Þ
3. Calculate the initial Wt to iterate according to an adiabatic expansion of an ideal gas. 4. Calculate the enthalpy of the saturated air at 8 by using H 8 = H 7 W t. 5. Determine T8, Ps8 D8 using H8 and P8.
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6. Obtain
n n1
logðP 7 =P 8 Þ ¼ logðT : 7 =T 8 Þ
n 7. W t;new ¼ n1 RðT 7 T 8 Þ:
ð5Þ
Wt ¼ H 8 = H 7 W t. T8last = T8. Determine T8, Ps8, D8 using H8 and P8. If abs (T8last T8) < 0.1 go to step 13. otherwise, go to step 5. 13. Obtain the enthalpy using the formula, H8 = H7 W t Æ g t. 14. Determine T8, Ps8, D8 using H8 and P8. 3.4. Surface heat exchanger Pinch technology is used in the analysis of heat exchange in the surface heat exchanger, and the temperature difference at the pinch point is 6 °C. Pinch technology is a graphical method of identifying technically and economically interesting energy efficiency measures. The minimum cooling and heating demands in the system can thereby be determined, together with the net heat for each temperature level. The concepts and methodology of pinch technology are well explained in the works of Linhoff et al. [16], Eastop and Croft [17], Linnhoff [18] and Mubarak Ebrahim [19]. The cold stream is cooling water, and the enthalpy is determined by Hcold = cpwt. The hot stream is hot wet air with a pressure of P4, and the enthalpy is determined by H hot ¼ 1:006t þ 0:001Ds ð2501 þ 1:805tÞ
ð7Þ
in which Ds ¼ 621:98
P s ðtÞ P 4 P s ðtÞ
q2 Wm
ð11Þ
ð6Þ
W t þW t;new . 2
8. 9. 10. 11. 12.
COP ¼
ð8Þ
The optimum mass flow rate ratio of cool water to dry air can be obtained from the hot and cold curves according to pinch technology.
5. Results There are many factors that may influence the COP of this refrigerating system for air conditioning and hot water cooled by cool water. These include the pressure ratio of the axial compressor, P4/P3, the efficiencies of the axial compressor and turbine, the wet bulb temperature of the atmosphere Twet and the cool water temperature. During simulation, the pressure ratio of the axial compressor was varied from 1.6 to 2.5, the wet bulb temperature of the outdoor air from 20 to 30 °C and the cooling water temperature from 15 to 25 °C. There is 300 Pa pressure loss before the axial compressor, 300 Pa between the axial compressor and turbine and 600 Pa after the turbine. Some encouraging results are illustrated in Figs. 2–6. The results are calculated from a lower pressure ratio of the axial compressor than that in Ref. [14]. The sensitivity of COP to the efficiencies of the axial compressor and turbine is illustrated in Fig. 2. The lines in Fig. 2 are the COP lines of an open air compression refrigeration system for air conditioning and hot water cooled by cool water when T7 = 296 K and the efficiencies of the axial compressor and turbine are 90%, 88%, 86% and 80%, respectively. From Fig. 2, the efficiencies of the axial compressor and the turbine influence the COP heavily. The COP is higher because of a lower T7 than we can get than in Ref. [14]. Also, the optimum pressure ratio of the axial compressor declines. That is, the number of stages could be decreased to two. Although the COP is sensitive to the efficiencies of the axial compressor and turbine, these cycles are feasible. Firstly, this new turbo-machinery works near the design point, and the efficiencies of axial compressors and turbines
5.5
4. Performance
5.0
The refrigerating capacity per kilogram dry air, q2, can be determined by the enthalpy difference between the inlet of the compressor and the outlet of the turbine by using the following formula
4.5
nt=nc=0.9 nt=nc=0.88
COP
4.0 3.5
ð9Þ
3.0
The work consumed by the refrigeration cycle is calculated by
2.5
q2 ¼ H 3 H 8
Wm ¼
Wc W t gt gc
ð10Þ
nt=nc=0.86 nt=nc=0.80
2.0 1.5 20
22
24
26
28
o
The COP of this refrigeration system is calculated by the following formula. (The work consumed by the cool water system is not included in Wm.)
T3 / Twet, C Fig. 2. The sensitivity of COP to efficiencies of the axial compressor and turbine when T7 = 23 °C.
S. Hou et al. / Energy Conversion and Management 48 (2007) 2255–2260
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60 50
T4
40
T8
30
4.0
COP
20
3.5
10 0
28 27 26 25 24 23 22 21
-10 20
22
24
26
20
28
t ,o C
3.0 21 22
o
T3 / Twet, C Fig. 3. The simulated temperatures of an open air-compression refrigeration system for air conditioning and hot water cooled by cool water when gc = 0.90 and gt = 0.90 and T7 = 23 °C.
T ,o 7 C
23 24 25
20
T 3 / T we
T, oC
4.5
T7
Fig. 5. The COP of an open air-compression refrigeration system for air conditioning and hot water cooled by cool water when gc = 0.90 and gt = 0.90.
100
80 90 60
q , kJ/kg(d.a) 1
80
2
40
wm, kJ/kg(d.a)
20
q2 , kJ/(d.a)
q , kJ/kg(d.a)
70
60
0 24
26
28
T3 / Twet,oC Fig. 4. The simulated q1, q2, Wm of an open air-compression refrigeration system for air conditioning and hot water cooled by cool water when gc = 0.90 and gt = 0.90 and T7 = 23 °C.
are very high at the design point, about 0.89–0.91. Secondly, the intake air is clean and without dust, and therefore, the efficiencies of the axial compressor and turbine will not drop greatly while working. Thirdly, there is no great complexity of the combustion chamber and high temperature turbine in the turbo-machinery. Consequently, it is much easier to accomplish than many people imagine. Lastly, the efficiencies of the axial compressor and turbine have room for improvement with additional design measures. The simulations of an open air–vapor compression refrigeration system for air conditioning and hot water cooled by cool water when gc = 0.90 and gt = 0.90 and T7 = 296 K are illustrated in Figs. 3 and 4. Fig. 3 gives the relations of the temperature after the compressor, T4, the temperature before the turbine, T7, and the temperature after the turbine, T8 to the temperature before the
28 27 26 25 24 23 22 21
50 20
t ,o C
22
21 22
T ,o 7 C
23 24 25
20
T 3 / T we
20
Fig. 6. The q2 of an open air-compression refrigeration system for air conditioning and hot water cooled by cool water when gc = 0.90 and gt = 0.90.
compressor, T3 (Twet). Fig. 4 gives the relations of the refrigerating capacity per kilogram dry air, q2, the discharging heat per kilogram dry air, q1, and the work consumed by the refrigeration system, Wm, to the temperature before the compressor, T3 (Twet). The relation of COP to the temperature before the compressor, T3 (Twet) can be located in Fig. 2. The simulations of an open air–vapor compression refrigeration system for air conditioning and hot water cooled by cool water when gc = 0.90and gt = 0.90 and T7 = 20–25 °C are illustrated in Figs. 5 and 6. Fig. 5 gives the variation of COP with T3 (Twet) and T7. Fig. 6 gives the variation of refrigerating capacity per kilogram dry air, q2, with T3(Twet) and T7.
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6. The COP of this refrigeration air conditioning system rests mainly on gc and gt. The temperature of cool water will also affect it heavily. Although the COP is sensitive to gc and gt, an open air–vapor compression refrigeration cycle for air conditioning and hot water cooled by cool water is still feasible. 7. The optimum pressure ratio of the axial compressor could be lower in this system. This results in a fewer number of stages of the axial compressor. 8. Pinch technology is a good method to analyze the heat exchange of wet air cooling.
70
4 60
4' o
4': T3=20 C
T, oC
50
o
4: T3=27 C
40
Outlet 30
7
Cool Water
20
Inlet 10 40
60
80
100
120
140
16 0
H, kJ/kg Fig. 7. The hot and cold curves of dehumidification process in surface heat exchanger when ratio = 1.
Fig. 7 gives the hot and cold curves of the cooling and dehumidification process in the surface heat exchanger when the mass flow rate ratio of cool water to dry air is one. In Fig. 7, the temperature difference at the pinch point is 6 °C, and there are two hot lines, which is when T3 = 20 °C and T3 = 27 °C, respectively. From Fig. 7, we can get hot water with a temperature of about 33–40 °C and the same mass flow rate as the dry air mass flow rate. 6. Conclusions This study shows the feasibility of an open air–vapor compression refrigeration system for air conditioning and hot water cooled by cool water. The calculation results show: 1. The open air–vapor compression refrigeration system for air conditioning and hot water cooled by cool water given in this paper could use the cool in cool water, which could not be used to cool air directly. Also, the heat rejected by this system could be used to heat water. 2. Humid air is a perfect working fluid for refrigeration in central air conditioning. 3. To compare with air–vapor compression refrigeration system for air conditioning cooled by circulating water [14], this system is suitable to be employed in the place where we could get cold water easily. 4. According to the axial compressor performance characteristics among rotation speed, mass flow rate and pressure ratio, we can easily control its mass flow rate and pressure ratio and the refrigerating capacity by adjusting the rotation speed of the axial compressor. 5. The COP of this refrigeration air conditioning circle varies with the wet bulb temperature of the atmosphere. The higher the wet bulb temperature of the atmosphere, the higher is the COP of this refrigerating air conditioning cycle.
References [1] Hamlin S, Hunt R, Tassou SA. Enhancing the performance of evaporative spray cooling in air cycle refrigeration and air condition technology. Appl Therm Eng 1998;18(10):1139–48. [2] Mc Cormick JA. Progress on the development of miniature turbomachines for low capacity reverse Brayton cryocooler. Proceedings of nineth international cryocooler conference. 1996:1;255–67. [3] Saygin H, Sisman A. Brayton refrigeration cycles working under quantum degeneracy conditions. Appl Energy 2001;69(2):77–85. [4] Sieniutycz S, Salamon P. Advances in thermodynamics. Finite-time thermodynamics and thermoeconomics, vol. 4. New York: Taylor & Francis; 1990. [5] Bejan A. Entropy generation minimization. Boca Raton, FL: CRC Press; 1996. [6] Berry RS, Kazakov VA, Sieniutycz S, Szwast Z, Tsirlin AM. Thermodynamic optimization of finite-time processes. Chichester: Wiley; 1999. [7] Chen L, Wu C, Sun F. Finite-time thermodynamics optimization or entropy generation minimization of energy systems. J Non-Equilib Thermodyn 1999;24(3):327–59. [8] Chen L, Wu C, Sun F. Cooling-load versus COP characteristics for an irreversible air refrigeration cycle. Energy Convers Manage 1998;39(1/2):117–25. [9] Chen L, Wu C, Sun F. Finite-time thermodynamic performance of an isentropic closed regenerated Brayton refrigeration cycle. Int J Energy Environ Econ 1997;4(4):261–74. [10] Chen L, Zhou S, Sun F, et al. Performance of heat-transfer irreversible regenerated Brayton refrigerators. Phys D: Appl Phys 2001;34(5):830–7. [11] Luo J, Chen L, Sun F, Wu C. Optimum allocation of heat-exchanger inventory of irreversible air refrigeration cycles. Phys Scripta 2002;65(5):410–4. [12] Northwestern Polytechnic University etc, Principle of gas turbine aero-engine, National Defense Industry Publisher, Beijing, 1981. [13] Hou SB, Li HC 1992, Chinese patent, 92109202.4. [14] Hou SB, Zhang HF. An axial-flow air-vapor compression refrigerating system for air conditioning cooled by circulating water. ASHRAE Transactions 2004;110(Pt 2):125–9; Hou SB, Zhang HF. Annual Meeting – Technical and Symposium Papers. American Society of Heating, Refrigerating and Air-Conditioning Engineers; 2004. p. 78–82. [15] ASHRAE Handbook – fundamentals volume, 2001, ASHRAE 1791 Tulie Circle, Atlanta, GA 30329, United States of America. [16] Linnhoff B et al. User guide on process integration for efficient use of energy. UK: AIChE; 1982. [17] Eastop TD, Croft DR. Energy efficiency for engineers and technologists. UK: Longman Scientific and Technical/John Wiley & Sons; 1990. [18] Linnhoff B. Pinch analysis – a state-of-the-art overview. Trans Icheme Part A 1993;71(September):503–22. [19] Mubarak Ebrahim A. Pinch technology: an efficient tool for chemical plant energy and capital cost saving. Appl Energy 2000;65(1–4):45–9.
Open air–vapor compression refrigeration system for air conditioning and hot water cooled by cool water Shaobo Hou b
a,b,*
, Huacong Li a, Hefei Zhang
a
a School of Power and Energy, Northwestern Polytechnical University, Xi’an, Shaanxi 710072, PR China College of Engineering, Guangdong Ocean University, East Jiefang Road, No. 40, Xiashan, Zhanjiang, Guangdong 524006, PR China
Received 25 January 2005; received in revised form 22 September 2006; accepted 9 May 2007 Available online 18 June 2007
Abstract This paper presents an open air–vapor compression refrigeration system for air conditioning and hot water cooled by cool water and proves its feasibility through performance simulation. Pinch technology is used in analysis of heat exchange in the surface heat exchanger, and the temperature difference at the pinch point is selected as 6 °C. Its refrigeration depends mainly on both air and vapor, more efficient than a conventional air cycle, and the use of turbo-machinery makes this possible. This system could use the cool in the cool water, which could not be used to cool air directly. Also, the heat rejected from this system could be used to heat cool water to 33– 40 °C. The sensitivity analysis of COP to gc and gt and the simulated results T4, T7, T8, q1, q2 and Wm of the cycle are given. The simulations show that the COP of this system depends mainly on T7, gc and gt and varies with T3 or Twet and that this cycle is feasible in some regions, although the COP is sensitive to the efficiencies of the axial compressor and turbine. The optimum pressure ratio in this system could be lower, and this results in a fewer number of stages of the axial compressor. Adjusting the rotation speed of the axial compressor can easily control the pressure ratio, mass flow rate and the refrigerating capacity. The adoption of this cycle will make the air conditioned room more comfortable and reduce the initial investment cost because of the obtained very low temperature air. Humid air is a perfect working fluid for central air conditioning and no cost to the user. The system is more efficient because of using cool water to cool the air before the turbine. In addition, pinch technology is a good method to analyze the wet air heat exchange with water. Ó 2007 Elsevier Ltd. All rights reserved. Keywords: Turbo-machinery; Air cycle; Air conditioning units; Natural working fluid; Refrigeration; Pinch technology
1. Introduction The outdoor air temperature can reach 38–40 °C in middle China in summer, while the temperature of the water from the water supply system and underground is about 17 °C and that of the refilled underground water could be lower. These waters are too cool for people to bath directly at the hotel, and the bath water is usually heated by a boiler. If these waters were used to cool air for air conditioning * Corresponding author. Address: School of Power and Energy, Northwestern Polytechnical University, Xi’an, Shaanxi 710072, PR China. Tel.: +86 759 2886705. E-mail addresses: [email protected], [email protected] (S. Hou).
0196-8904/$ - see front matter Ó 2007 Elsevier Ltd. All rights reserved. doi:10.1016/j.enconman.2007.05.005
directly, it is not efficient because of the small temperature difference and large amount of consumed water. The air compression refrigeration cycle was studied long ago. Several disadvantages prevented air from being used as a working fluid in refrigeration. These included low volumetric refrigerating effect, which may result in a large compressor, and low COP due to low efficiencies of the compressor and expander. After CFC’s invention in the 1930s, people paid little attention to air compression refrigeration. Recently, as a result of the destruction of the ozonosphere by chlorofluorocarbons (CFC) and the pressure of environmental protection, research upon air refrigeration cycles has received more attention [1–3]. Optimizations of
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S. Hou et al. / Energy Conversion and Management 48 (2007) 2255–2260
Nomenclature B D H P T Twet Wc Wt Wm gc gt ratio
wet air pressure, Pa humidity ratio of wet air, g/kg(d.a.) enthalpy of wet air, kJ/kg(d.a.) pressure, Pa temperature, K or °C wet bulb temperature, K or °C ideal input work of compressor, kJ/kg(d.a.) ideal output work of turbine, kJ/kg(d.a.) practical work consumed by system, kJ/kg(d.a.) efficiency of compressor efficiency of turbine mass flow rate ratio of cooling water to dry air
air cycles are also conducted using finite time thermodynamics (FTT) or entropy generation minimization (EGM) [4–7]. Chen et al. investigated the cooling load versus COP characteristics of a simple [8] and a regenerated [9,10] air refrigeration cycle with heat transfer loss and/or other irreversibilities. Luo et al. [11] optimized the cooling load and the COP of a simple irreversible air refrigeration cycle by searching for the optimum pressure ratio of the compressor and the optimum distribution of heat conductance of the hot and cold side heat exchangers for fixed total heat exchanger inventory. With the development of the aeronautical industry, highly efficient axial compressors and turbines have become a reality. At present, the stagnation isentropic efficiencies of a single stage axial compressor and a turbine can reach 0.88–0.91 [12]. High speed fans have been used in ordinary air conditioning systems nowadays. However, the water vapor in the working fluid was not considered in all the above researches [1–11] on the air compression refrigeration cycle, and the used equipments were a centrifugal compressor and a centripetal turbine, which have lower efficiencies than axial compressors and turbines. The amount of water extracted from high pressure wet air can reach 18–30 g/kg(d.a.), and the amount of latent heat discharged from the vapor condensed is about 45–75 kJ/kg(d.a.), exceeding the sensible heat from the air of 30–50 kJ/kg(d.a.). Hou and Li presented an axial flow air–vapor compression refrigerating system for air conditioning in 1992 [13] in which wet air is the working fluid and an axial compressor and turbine were used, but these have not yet attracted people’s attention so far. Hou and Zhang presented an axial flow air–vapor compression refrigerating system for air conditioning cooled by circulating water in Ref. [14] (2004) in which wet air is the working fluid, an axial compressor and turbine were used and the wet air is cooled by circulating water. The paper proves its feasibility through performance simulation and also indicates its advantages. These include the possibility
R n
gas constant, kJ/kg K exponent
Subscripts air dry air vapor water vapor in moist air s saturated last last time hot hot stream cold cold stream w water
to simplify air conditioning systems, to reduce the amount of initial investment of an air conditioning system and to make air conditioned rooms more comfortable. The aim of this paper is to present an open system, which is an open air–vapor compression refrigeration system for air conditioning and hot water cooled by cool water and its performance from simulation. In this open air–vapor refrigeration cycle, water from the water supply system or underground is used. Thus, we could get a lower wet air temperature before the turbine. In addition, the cool water is heated for the bath. 2. System Representation on enthalpy–entropy coordinates and a circuit diagram of an open air-compression refrigeration system for air conditioning and hot water cooled by cool water are shown in Fig. 1. The outdoor air at 2 is drawn into the atomizing chamber, cooled to saturated air at 3 with some fine water droplets and then compressed by an axial compressor. A flow of compressed air at 4 with higher temperature, T4, and high pressure, P4, is obtained. Then, the compressed air at 4 is cooled to saturated air at 7 with a temperature T7 by cool water/underground water in a surface heat exchanger after
Fig. 1. Representation on enthalpy–entropy coordinates and circuit diagram of an open air-compression refrigeration system for air conditioning and hot water cooled by cool water.
S. Hou et al. / Energy Conversion and Management 48 (2007) 2255–2260
the axial compressor outlet. Some vapor is condensed, and the latent heat of the vapor is discharged from 4 to 7. Then, the saturated air at 7 is expanded and cooled to the cool air at 8 in the turbine. The cool air at 8 is then ducted to the air conditioned rooms. The cool water is heated in the surface heat exchanger. Water injection before the axial compressor aims to decrease both the temperature of the working fluid and the polytropic exponent in the compression process. Thus, we can save some compression work. This method has been used in a jet engine when a fighter plane increases its speed. However, the difference is that what is injected in a jet engine is water, alcohol, etc. [12]. The water vapor in the compressed air can easily be extracted by a surface heat exchanger. With the same temperature, the humidity ratio of the saturated wet air at high pressure P4 is only about P3/P4 of that at pressure P3. The method of using compressed air to acquire dry air has been used in some workshops in southern China. The system above differs from a conventional air cycle system. There are many characteristics in this air–vapor refrigeration circle. Firstly, an axial compressor and a turbine are used in the above system. The characteristics of turbo-machines are large mass flow rate and high efficiency. The other types of compressor and expander have none of the above advantages. Secondly, this refrigeration system intakes precooled wet air with fine water droplets, and some vapor is condensed during the air cooling from 4 to 7. The amount of water extracted from the high pressure wet air can reach 18–30 g/kg(d.a.), and the amount of latent heat discharged from the vapor condensed, about 45–75 kJ/kg(d.a.), exceeds the sensible heat from the air, 30–50 kJ/kg(d.a.). For this reason, the refrigeration load in this air–vapor refrigeration system depends on a combination of the sensible heat of air and the latent heat of vapor. Lastly, the cool from the cooling water was used. Usually, it cannot be used. 3. Performance simulation 3.1. Wet air The humidity ratio of wet air, D, is obtained from D ¼ 621:98
P vap B P vap
ð1Þ
The enthalpy of wet air, H, is calculated from H ¼ 1:006t þ 0:001Dð2501 þ 1:805tÞ
ð2Þ
The adopted relation for water vapor between saturation pressure and saturation temperature, Ps = f(ts), is selected from Ref. [15]. To calculate the saturated temperature of the wet air from the saturated enthalpy, Eqs. (1) and (2) and Ps = f(ts) are used.
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3.2. Axial compressor During the compression process of the wet air, the fine water droplets in the air may evaporate. Because the evaporation of water takes in heat, we can regard the ideal compression process of the wet air in the compressor as a polytropic process. Therefore, we can obtain the ideal work of the compressor per kilogram dry air, Wc, from " # n1 n p4 n ðRair þ 0:001DRvapor ÞT 3 1 Wc ¼ ð3Þ n1 p3 in which n is the polytropic exponent for the compression process. The practical work consumed by the axial compressor is Wc/gc in which gc is the thermal efficiency of the compressor.
3.3. Turbine The saturated air with a pressure of P7 and a temperature of T7 before the turbine has been dehumidifed in the surface heat exchanger by cooling water. At point 7, the amount of vapor included in the saturated air is very small, about P3/P7 of the amount included in saturated air at P3. Thus, the water condensed in the air is in fog. Nevertheless, expansion of the saturated air in the turbine cannot be regarded as an adiabatic expansion of an ideal gas. With the decrease of the wet air pressure in the turbine, the temperature of the wet air decreases, and some heat is discharged during the condensation of some water vapor. The heat discharged may cause increases in both the temperature of the turbine outlet and the work done in the expansion. For this problem, we can imagine that no phase change exists and that there is some heat added to the wet air during the expansion process when we calculate the work done by the expansion process. According to the above assumption, this problem can be simplified to a problem of the polytropic expansion of an ideal gas. Consequently, we can obtain the ideal work done by the expansion, Wt, through iteration and then obtain the real work generated by the turbine and the temperature of the turbine outlet. The following are the steps to calculate Wt and T8. 1. Determine Ps7 from T7, D7 from Ps7 and P7, and H7 from D7 and T7. 2. Get gas constant R for the saturated air at 7 by using the formula: R ¼ 0:001ð287 þ 0:461D7 Þ:
ð4Þ
3. Calculate the initial Wt to iterate according to an adiabatic expansion of an ideal gas. 4. Calculate the enthalpy of the saturated air at 8 by using H 8 = H 7 W t. 5. Determine T8, Ps8 D8 using H8 and P8.
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6. Obtain
n n1
logðP 7 =P 8 Þ ¼ logðT : 7 =T 8 Þ
n 7. W t;new ¼ n1 RðT 7 T 8 Þ:
ð5Þ
Wt ¼ H 8 = H 7 W t. T8last = T8. Determine T8, Ps8, D8 using H8 and P8. If abs (T8last T8) < 0.1 go to step 13. otherwise, go to step 5. 13. Obtain the enthalpy using the formula, H8 = H7 W t Æ g t. 14. Determine T8, Ps8, D8 using H8 and P8. 3.4. Surface heat exchanger Pinch technology is used in the analysis of heat exchange in the surface heat exchanger, and the temperature difference at the pinch point is 6 °C. Pinch technology is a graphical method of identifying technically and economically interesting energy efficiency measures. The minimum cooling and heating demands in the system can thereby be determined, together with the net heat for each temperature level. The concepts and methodology of pinch technology are well explained in the works of Linhoff et al. [16], Eastop and Croft [17], Linnhoff [18] and Mubarak Ebrahim [19]. The cold stream is cooling water, and the enthalpy is determined by Hcold = cpwt. The hot stream is hot wet air with a pressure of P4, and the enthalpy is determined by H hot ¼ 1:006t þ 0:001Ds ð2501 þ 1:805tÞ
ð7Þ
in which Ds ¼ 621:98
P s ðtÞ P 4 P s ðtÞ
q2 Wm
ð11Þ
ð6Þ
W t þW t;new . 2
8. 9. 10. 11. 12.
COP ¼
ð8Þ
The optimum mass flow rate ratio of cool water to dry air can be obtained from the hot and cold curves according to pinch technology.
5. Results There are many factors that may influence the COP of this refrigerating system for air conditioning and hot water cooled by cool water. These include the pressure ratio of the axial compressor, P4/P3, the efficiencies of the axial compressor and turbine, the wet bulb temperature of the atmosphere Twet and the cool water temperature. During simulation, the pressure ratio of the axial compressor was varied from 1.6 to 2.5, the wet bulb temperature of the outdoor air from 20 to 30 °C and the cooling water temperature from 15 to 25 °C. There is 300 Pa pressure loss before the axial compressor, 300 Pa between the axial compressor and turbine and 600 Pa after the turbine. Some encouraging results are illustrated in Figs. 2–6. The results are calculated from a lower pressure ratio of the axial compressor than that in Ref. [14]. The sensitivity of COP to the efficiencies of the axial compressor and turbine is illustrated in Fig. 2. The lines in Fig. 2 are the COP lines of an open air compression refrigeration system for air conditioning and hot water cooled by cool water when T7 = 296 K and the efficiencies of the axial compressor and turbine are 90%, 88%, 86% and 80%, respectively. From Fig. 2, the efficiencies of the axial compressor and the turbine influence the COP heavily. The COP is higher because of a lower T7 than we can get than in Ref. [14]. Also, the optimum pressure ratio of the axial compressor declines. That is, the number of stages could be decreased to two. Although the COP is sensitive to the efficiencies of the axial compressor and turbine, these cycles are feasible. Firstly, this new turbo-machinery works near the design point, and the efficiencies of axial compressors and turbines
5.5
4. Performance
5.0
The refrigerating capacity per kilogram dry air, q2, can be determined by the enthalpy difference between the inlet of the compressor and the outlet of the turbine by using the following formula
4.5
nt=nc=0.9 nt=nc=0.88
COP
4.0 3.5
ð9Þ
3.0
The work consumed by the refrigeration cycle is calculated by
2.5
q2 ¼ H 3 H 8
Wm ¼
Wc W t gt gc
ð10Þ
nt=nc=0.86 nt=nc=0.80
2.0 1.5 20
22
24
26
28
o
The COP of this refrigeration system is calculated by the following formula. (The work consumed by the cool water system is not included in Wm.)
T3 / Twet, C Fig. 2. The sensitivity of COP to efficiencies of the axial compressor and turbine when T7 = 23 °C.
S. Hou et al. / Energy Conversion and Management 48 (2007) 2255–2260
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60 50
T4
40
T8
30
4.0
COP
20
3.5
10 0
28 27 26 25 24 23 22 21
-10 20
22
24
26
20
28
t ,o C
3.0 21 22
o
T3 / Twet, C Fig. 3. The simulated temperatures of an open air-compression refrigeration system for air conditioning and hot water cooled by cool water when gc = 0.90 and gt = 0.90 and T7 = 23 °C.
T ,o 7 C
23 24 25
20
T 3 / T we
T, oC
4.5
T7
Fig. 5. The COP of an open air-compression refrigeration system for air conditioning and hot water cooled by cool water when gc = 0.90 and gt = 0.90.
100
80 90 60
q , kJ/kg(d.a) 1
80
2
40
wm, kJ/kg(d.a)
20
q2 , kJ/(d.a)
q , kJ/kg(d.a)
70
60
0 24
26
28
T3 / Twet,oC Fig. 4. The simulated q1, q2, Wm of an open air-compression refrigeration system for air conditioning and hot water cooled by cool water when gc = 0.90 and gt = 0.90 and T7 = 23 °C.
are very high at the design point, about 0.89–0.91. Secondly, the intake air is clean and without dust, and therefore, the efficiencies of the axial compressor and turbine will not drop greatly while working. Thirdly, there is no great complexity of the combustion chamber and high temperature turbine in the turbo-machinery. Consequently, it is much easier to accomplish than many people imagine. Lastly, the efficiencies of the axial compressor and turbine have room for improvement with additional design measures. The simulations of an open air–vapor compression refrigeration system for air conditioning and hot water cooled by cool water when gc = 0.90 and gt = 0.90 and T7 = 296 K are illustrated in Figs. 3 and 4. Fig. 3 gives the relations of the temperature after the compressor, T4, the temperature before the turbine, T7, and the temperature after the turbine, T8 to the temperature before the
28 27 26 25 24 23 22 21
50 20
t ,o C
22
21 22
T ,o 7 C
23 24 25
20
T 3 / T we
20
Fig. 6. The q2 of an open air-compression refrigeration system for air conditioning and hot water cooled by cool water when gc = 0.90 and gt = 0.90.
compressor, T3 (Twet). Fig. 4 gives the relations of the refrigerating capacity per kilogram dry air, q2, the discharging heat per kilogram dry air, q1, and the work consumed by the refrigeration system, Wm, to the temperature before the compressor, T3 (Twet). The relation of COP to the temperature before the compressor, T3 (Twet) can be located in Fig. 2. The simulations of an open air–vapor compression refrigeration system for air conditioning and hot water cooled by cool water when gc = 0.90and gt = 0.90 and T7 = 20–25 °C are illustrated in Figs. 5 and 6. Fig. 5 gives the variation of COP with T3 (Twet) and T7. Fig. 6 gives the variation of refrigerating capacity per kilogram dry air, q2, with T3(Twet) and T7.
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6. The COP of this refrigeration air conditioning system rests mainly on gc and gt. The temperature of cool water will also affect it heavily. Although the COP is sensitive to gc and gt, an open air–vapor compression refrigeration cycle for air conditioning and hot water cooled by cool water is still feasible. 7. The optimum pressure ratio of the axial compressor could be lower in this system. This results in a fewer number of stages of the axial compressor. 8. Pinch technology is a good method to analyze the heat exchange of wet air cooling.
70
4 60
4' o
4': T3=20 C
T, oC
50
o
4: T3=27 C
40
Outlet 30
7
Cool Water
20
Inlet 10 40
60
80
100
120
140
16 0
H, kJ/kg Fig. 7. The hot and cold curves of dehumidification process in surface heat exchanger when ratio = 1.
Fig. 7 gives the hot and cold curves of the cooling and dehumidification process in the surface heat exchanger when the mass flow rate ratio of cool water to dry air is one. In Fig. 7, the temperature difference at the pinch point is 6 °C, and there are two hot lines, which is when T3 = 20 °C and T3 = 27 °C, respectively. From Fig. 7, we can get hot water with a temperature of about 33–40 °C and the same mass flow rate as the dry air mass flow rate. 6. Conclusions This study shows the feasibility of an open air–vapor compression refrigeration system for air conditioning and hot water cooled by cool water. The calculation results show: 1. The open air–vapor compression refrigeration system for air conditioning and hot water cooled by cool water given in this paper could use the cool in cool water, which could not be used to cool air directly. Also, the heat rejected by this system could be used to heat water. 2. Humid air is a perfect working fluid for refrigeration in central air conditioning. 3. To compare with air–vapor compression refrigeration system for air conditioning cooled by circulating water [14], this system is suitable to be employed in the place where we could get cold water easily. 4. According to the axial compressor performance characteristics among rotation speed, mass flow rate and pressure ratio, we can easily control its mass flow rate and pressure ratio and the refrigerating capacity by adjusting the rotation speed of the axial compressor. 5. The COP of this refrigeration air conditioning circle varies with the wet bulb temperature of the atmosphere. The higher the wet bulb temperature of the atmosphere, the higher is the COP of this refrigerating air conditioning cycle.
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