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Proposal and analysis of a novel ammonia–water cycle for power and refrigeration cogeneration
ARTICLE IN PRESS
Energy 32 (2007) 961–970 www.elsevier.com/locate/energy
Proposal and analysis of a novel ammonia–water cycle for power and refrigeration cogeneration Meng Liua,b, Na Zhanga, a
Institute of Engineering Thermophysics, Chinese Academy of Sciences Graduate School of the Chinese Academy of Sciences, PO Box 2706, Beijing 100080, P.R. China
b
Abstract Cogeneration has improved sustainability as it can improve the energy utilization efficiency significantly. In this paper, a novel ammonia-water cycle is proposed for the cogeneration of power and refrigeration. In order to meet the different concentration requirements in the cycle heat addition process and the condensation process, a splitting /absorption unit is introduced and integrated with an ammonia–water Rankine cycle and an ammonia refrigeration cycle. This system can be driven by industrial waste heat or a gas turbine flue gas. The cycle performance was evaluated by the exergy efficiency, which is 58% for the base case system (with the turbine inlet parameters of 450 1C/11.1 MPa and the refrigeration temperature below 15 1C). It is found that there are certain split fractions which maximize the exergy efficiency for given basic working fluid concentration. Compared with the conventional separate generation system of power and refrigeration, the cogeneration system has an 18.2% reduction in energy consumption. r 2006 Elsevier Ltd. All rights reserved. Keywords: Power/refrigeration cogeneration; Ammonia–water; Exergy efficiency
1. Introduction Binary component mixtures exhibit a boiling temperature that varies during the boiling process, and their employment as working fluids thus allows maintenance of a more constant temperature difference between them and variable temperature heat sources, and consequently reduced exergy losses in the heat addition process. To take advantage of the feature, Maloney and Robertson [1] first introduced the use of an ammonia/water mixture as the working fluid in an absorption power cycle in the early 1950s. However, the cycle condensation process takes place at a variable temperature too, resulting in a higher turbine backpressure (or lower turbine pressure ratio) than that of the conventional steam Rankine cycle. Higher backpressure is good to prevent air inleakage into the system, but unfavourable to the power generation and efficiency. Kalina proposed a combined power cycle which employs an ammonia/water mixture as the bottoming cycle working Corresponding author. Tel.: +86 10 82543030; fax: +86 10 62575913.
E-mail address: [email protected] (N. Zhang). 0360-5442/$ - see front matter r 2006 Elsevier Ltd. All rights reserved. doi:10.1016/j.energy.2006.09.012
fluid, and solved this problem by replacing the condensation process with an absorption process [2]. A combined cycle for power/refrigeration cogeneration was proposed by Goswami [3], and some further researches on the cycle performance were carried out [4–6]. This cycle employs the ammonia-rich vapor separated by a rectifier unit as the turbine working fluid to generate power, and then the turbine exhaust provides cooling by transferring sensible heat to the chilled water. Only about 16% of the total working fluid is used to generate power and refrigeration, and the turbine backpressure has to be maintained at a rather high level (0.2 MPa) to condense the working fluid in the absorption process. Zheng et al. [7] also proposed a combined cycle based on the Kalina cycle, which replaces the flash tank with a rectification column, the outflow from the top of the rectifier is throttled by a valve and then produces refrigeration before mixing the main stream. An energy efficiency of 24.2% and an exergy efficiency of 37.3% were reported with the turbine inlet parameters of 350 1C/5 MPa. In integrated refrigeration/power combined cycles, generally a basic refrigeration subcycle and a basic power
ARTICLE IN PRESS M. Liu, N. Zhang / Energy 32 (2007) 961–970
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Nomenclature
Subscripts
E e h m p pb Q RR SF s T t Wnet X DTp y
a ambient state Bs basic solution cw cooling water EVA evaporator HEX heat exchanger hs heat source fluid i inlet in input o outlet REB reboiler REC rectifier T turbine 0 base case 1,2,y26 states on the cycle flow sheet
exergy (kW) specific exergy (kJ/kg) specific enthalpy (kJ/kg) mass flow rate (kg/s) pressure (MPa) turbine backpressure (MPa) heat duty (kW) rectifier molar reflux ratio split fraction specific entropy (kJ/kg K) temperature (K) temperature (1C) net power output (kW) ammonia mass fraction (kg ammonia/kg mixture) pinch point temperature difference (K) exergy efficiency (%)
subcycle can be identified, and their integration can be accomplished by different configurations. The authors and colleagues have proposed and analyzed two basic cycle configurations: the parallel- and the series-connected ones [8–10]. It is found that both have some advantages and disadvantages [11]. In the parallel system, the working fluid in the power subcycle has a fairly low concentration, hence the desired temperature glide in the power subcycle boiling process does not occur, but the turbine may have a low backpressure. In the series-connected cycle, the working fluid has a higher ammonia concentration in the heat addition process and thus a better temperature glide and heat transfer match, but the turbine backpressure is high, thus reducing power output and increasing the exergy loss in the condensation process. The ideal system should have low exergy losses in both the heat addition and condensation processes. Obviously, different ammonia concentrations should be maintained in these processes. In this paper, a more advanced cycle configuration is proposed and analyzed, as an improvement upon the series-connected configuration [9]. A splitting/absorption unit is employed to adjust some stream mass flow rates and thus maintain the desired ammonia concentrations in different processes. Some different cycle configurations are available too by varying the split fractions.
2. The base case cycle configuration description Fig. 1 shows the base case cycle configuration. The refrigeration subcycle can be identified as (5-6-7-8-9-9A/ 9B-9C), and the power subcycle as (10–18). A splitting/ absorption unit (including SPL1, SPL2, ABS1, ABS2 and ABS3) is inserted in between, to accomplish the absorption and condensation processes, and at the same time to adjust some stream mass flow rates for meeting the different
concentration requirements in the cycle heat addition process and the condensation process. An ammonia–water mixture, referred as the basic concentration solution (1) is split into two streams (1A, 1B) by the splitter SPL1. Stream (1A) is heated and sent to the rectifier REC, where it is separated into a high concentration vapor (almost pure ammonia) and a weak concentration solution (19). The high concentration vapor from the top of the rectifier REC is condensed and partially sent to the refrigeration subcycle (5–9) to provide refrigeration. It is then split into two streams (9A, 9B) by the splitter SPL2, and sent to the absorbers ABS1 and ABS2, respectively. Stream (1B) is the main working fluid in the power subcycle. Before being heated in the recuperator REP, it is enriched by mixing with the strong solution (9C) in the absorber ABS2. This helps to improve the heat transfer match in the subsequent boiling process in the boiler B. The splitters SPL1 and SPL2 thus function to adjust both the concentration and the mass flow rate of stream (10), and thereby affect the power output. Moreover, the mass flow rate of the refrigerant (5) and the related refrigeration output would be varied too along with the variation of SPL1 split fraction. In the power subcycle, stream (10) is pumped to the system high pressure level (11) and evaporated and superheated to the highest cycle temperature (14). It expands to generate power in the turbine T. The turbine exhaust (15) is cooled, and then diluted by merging with the weak concentration solution (22) in the absorber ABS3. Consequently, the power subcycle condensation process is replaced by an absorption process, and the related exergy destruction could be reduced with a possibly lower condensation pressure. The working fluid has five ammonia mass concentration levels as shown in Fig. 1, and the combined cycle also has four pressure levels.
ARTICLE IN PRESS M. Liu, N. Zhang / Energy 32 (2007) 961–970
963
8 C
5
6
V1
EVA
7
SPL2
9 CON
9B
9A HEX2
4
REC
25 19 REB
24
13
3
HEX1
26 20
2
1A SPL1
1B
22 14 T
23
16
15 REP
V2
18
V3
B
1 ABS1
21
12
PRE
P1
P3 ABS3
17
P2 11
10
ABS2
9C
High concentration solution (refrigeration subcycle) Intermediate concentration solution (power subcycle) Basic concentration solution Intermediate concentration solution (power subcycle) Low concentration solution Heat source fluid Cooling water
ABS - Absorber HEX - Heat exchanger REP - Recuperator
B - Boiler P - Pump SPL- Splitter
C - Cooler PRE - Preheater T-Turbine
CON - Condenser REB - Reboiler V - Valve
EVA - Evaporator REC - Rectifier
Fig. 1. The flow sheet of the combined power/refrigeration cycle.
In this cycle, two improvements can be identified comparing with the series-connected configuration in [9]. Firstly, the splitting/absorption unit brings a stronger solution into the power subcycle heat addition process, thus making a better temperature match with the heat source fluid. Meanwhile, the mass flow rate of the working fluid in the refrigeration subcycle can be adjusted too. Secondly, the turbine exhaust is diluted in the absorption-condensation process, allowing a lower turbine backpressure and thus less exergy destruction during the condensation process. The cycle can be heated by the flue gas of a gas turbine or any other industrial waste heat with a suitable temperature. The heat source fluid (23) flows through the hot sides of the boiler B (24), the reboiler REB (25) and the heat exchanger HEX2 (26) in turn, and is eventually exhausted to the environment. This novel cycle has two useful outputs, power and refrigeration. Obviously the power and refrigeration outputs and the heat input are different in energy quality, hence the exergy efficiency is chosen to be the criterion for the cycle performance evaluation.
The exergy efficiency is defined as the exergy outputs divided by the exergy input to the cycle: y¼
W net þ E EVA , E in
(1)
where the net power output Wnet is the power output from the turbine, reduced by the power input to the pumps, and the exergy of the heat source fluid, Ein, is given as E in ¼ mhs ½ðhhs;i hhs;a Þ T a ðshs;i shs;a Þ.
(1a)
Since the heat source fluid is finally exhausted to the environment, the calculation of the exergy input is based on the difference between its initial state and the environment state. EEVA is the exergy associated with the refrigeration output, which is calculated as the working fluid exergy difference across the evaporator EVA. E EVA ¼ mEVA;i ½ðhEVA;i hEVA;o Þ T a ðsEVA;i sEVA;o Þ. (1b)
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Table 1 Main assumptions for the base-case calculation Items
Parameters
Values
Cycle parameter
Ammonia mass fraction of the basic solution XBs Cooling water temperature tcw (1C) SF1 SF2 Pressure loss (%) Refrigeration temperature upper limit (1C) Pressure loss (%) Temperature ta (1C) Pressure pa (MPa) Inlet temperature tT,i (1C) Isentropic efficiency (%) Minimum exhaust vapor fraction Theoretical stage number Molar reflux ratio RR Operation pressure pREC (MPa) Pressure loss (%) Outlet temperature tREB,o (1C) Pinch point temperature difference DTp (K) Pressure loss (%) Efficiency (%) Air Inlet temperature ths,i (1C) The minimal exhaust temperature (1C)
0.23 30 0.6 0.96 3 15 3 25 0.101 450 87 0.88 6 0.3 1.4 3 165 5 or 15 (if one side is air) 1.0–3.0 75 79% N2 and 21% O2 (by volume) 465 90
Split fraction Evaporator EVA Absorber ABS Ambient state Turbine T
Rectifier REC
Reboiler REB Heat exchangers (ABS, B, C, CON, HEX, PRE, REB, REP) Pumps P Heat source fluid
The split fractions of the splitters SPL1, SPL2 are defined
3. The base case cycle performance as It is assumed that the system operates at steady state. The simulation was carried out using the commercial Aspen Plus [12] code, in which the component models are based on energy and mass balances, with the default relative convergence error tolerance of 0.01%; the thermal property were calculated with the Electrolyte NRTL model or the SR-Polar model for high temperature (4246 1C) and pressure (410 MPa) application. To validate the properties calculations, the property results from Aspen+ and the data published by the International Institute of Refrigeration [13] were compared, and the result shows good agreement between them. The distillation column has been simulated by the RadFrac model, which is designed for simulating multistage vapour–liquid fractionation operation and is suitable for two-phase system. The main assumptions are shown in Table 1. The basic concentration solution (1) has the ammonia mass fraction of 0.23 and the mass flow rate of 2.0 kg/s. The gas turbine flue gas is one of the possible heat sources of this cycle. Since the small/middle sized gas turbine exhaust temperature values at 500–550 1C typically, the heating fluid inlet (23) and minimal outlet (26) temperatures of the proposed system are chosen to be 450 and 90 1C, respectively. The working fluid temperature at the absorber ABS2 outlet is set to be 35 1C in order to keep the minimal heat transfer temperature difference (5 K) with the external cooling water (30 1C).
SF 1 ¼ m1A =m1 ,
(2)
SF 2 ¼ m9A =m9 .
(3)
Table 2 summarizes the parameters of some main streams, including temperature t, pressure p, vapor fraction and ammonia mass fraction X, of the base case cycle configuration. The base case cycle performance is reported in Table 3. Keeping the pinch point temperature of 15 1C in the heat addition process, the heat source fluid mass flow rate should be 7.5 kg/s, and it exits the system at 90 1C (the prescribed minimal exhaust temperature). The system exergy efficiency is found to be 57.6%. 4. The effect of the split fractions 4.1. The effect of the split fractions on the cycle configuration By varying the split fractions of SPL1 and SPL2, some other cycle configurations are available as follows: (A) configuration A (0oSF1o1, SF2 ¼ 0): a power/ refrigeration cogeneration system as shown in Fig. 2, in which the strong solution from the refrigeration subcycle is all sent to ABS2 to enrich the power subcycle working fluid. As a result, the irreversibility
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Table 2 The base case cycle stream states
C
5
State No.
t (1C)
p (MPa)
965
Vapor fraction
m (kg/s)
V1
6
EVA
7
X
9
CON
1 1A 4 7 8 9 9A 10 14 15 17 19 22
37 37 134.4 34.4 15 24.2 24.2 35 450 68 35 165 47.5
0.077 0.077 1.442 0.084 0.082 0.079 0.079 0.075 11.10 0.039 0.037 1.400 0.038
0 0 0 0.1 0.79 0.91 0.91 0 1 0.88 0 0 0
2 1.2 1.2 0.195 0.195 0.195 0.187 0.808 0.808 0.808 1.813 1.005 1.005
0.23 0.23 0.23 0.92 0.92 0.92 0.92 0.237 0.237 0.237 0.159 0.096 0.096
HEX2
4
REC
25 19 REB 24
13
HEX1
3 26 20
21
12
V3
PRE
B
P1
2
1A SPL1
1
1B
22 14 T
23
P3
ABS3
16
15 REP
17
P2
11
ABS2
10
Fig. 2. Configuration A of the novel cycle (0oSF1o1, SF2 ¼ 0). Table 3 The base case cycle performance summary Parameters Turbine T work (kW) Pump work (kW) P1 P2 P3 Recuperator REP heat load (kW) Absorber heat load (kW) ABS1 ABS2 ABS3 Heat exchangers heat input (kW) HEX1 HEX2 PRE Condenser CON load (kW) Cooler C load (kW) Boiler B heat input (kW) Reboiler REB heat input (kW) Total heat input Qin (kW) Refrigeration capacity QEVA (kW) Refrigeration exergy EEVA (kW) Net power output Wnet (kW) Exergy input Ein (kW) Exergy efficiency y (%)
Values
8
754.3 CON
2.6 14.3 0.1 89.6 315.7 21.5 1717.1
HEX2
25 19 REB
24
6
V1
13
3
HEX1
26 20
21
12
V3
PRE
B
would decrease in the boiling process but would increase in the condensation process; (B) configuration B (0oSF1o1, SF2 ¼ 1): a power/refrigeration cogeneration system as shown in Fig. 3, in which the strong solution from EVA is all sent to ABS1. This configuration has lower irreversibility in the condensation process but not in the boiling process; (C) configuration C (SF1 ¼ 0): a regenerative Rankine type power cycle; (D) configuration D (SF1 ¼ 1, SF2 ¼ 1): a conventional ammonia absorption refrigeration cycle.
23
2
P1
1A SPL1
1 ABS1
1B
18
22
14
141 641.5 457.4 362.4 61.5 1926.7 381.6 3444.4 203.8 46.7 737.3 1361.5 57.6
EVA
7
9 4
REC
C
5
T
16
15 REP
11
P3 ABS3
17
P2
Fig. 3. Configuration B of the novel cycle (0oSF1o1, SF2 ¼ 1).
4.2. The effect of the split fractions on the cycle performance The split fractions SF1, SF2 are the key parameters to the cycle performance, their effects are investigated and the results are shown in Figs. 4a and b. To accomplish the condensation process in the absorbers ABS1 and ABS2, their operation pressures need to be maintained above certain values for given cooling water temperature. This imposes a low limit on the operation pressure of the evaporator EVA. When SF1 increases or SF2 decreases, the ammonia concentration of stream 10 would increase, requiring higher absorption pressure in ABS2 for its condensation. Accordingly, the EVA operation pressure need to be elevated too, and so does the produced refrigeration temperature level. When the EVA working fluid inlet temperature t7 increases to 15 1C, no refrigeration would be produced. In this calculation, t7 is set to be under 21 1C to ensure certain
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58
1.2 (%)
EEVA/EEVA,0
1.4 1.0 0.8
54
0.6 SF1=0.50 SF1=0.60 SF1=0.70
0.4 1.4 1.2
52
1.0 0.8 0.6
1.2 1.0 0.8
0.0
0.2
0.4
0.6
0.8
0.0
1.0
SF2
(a)
SF1=0.50 SF1=0.60 SF1=0.70
1.4 mhs/mhs,0
Wnet/Wnet,0
56
0.2
0.4
0.6
0.8
1.0
SF2
(b)
Fig. 4. (a) The effects of the split fractions on the normalized net power output and refrigeration exergy output. (b) The effects of split fractions on the normalized heating fluid mass flow rate and the exergy efficiency.
Table 4 The effect of X
Bs
on the optimum split fractions and the optimized exergy efficiency
Ammonia mass fraction of the basic solution, XBs
XBs ¼ 0.20 XBs ¼ 0.23 (base case) XBs ¼ 0.25 XBs ¼ 0.27 XBs ¼ 0.30
Optimum split fractions
The optimized exergy efficiency (%)
SF1
SF2
0.66 0.6 0.57 0.58 0.59
0.71 0.96 1 1 1
amount refrigeration output be produced, which requires SF2 be above certain value for higher SF1 application. Fig. 4a shows the effect of the split fractions on the net power output Wnet and the refrigeration exergy output EEVA. When SF1 increases, more working fluid is sent to the refrigeration subcycle, therefore the refrigeration exergy output increases; when SF2 increases, less working fluid is sent to the power subcycle, therefore and the power output Wnet drops. At the same time, the needed external heat amount drops too, as represented by the heating fluid mass flow rate mhs in Fig. 4b. It is found that for given SF1, there is a SF2 which maximizes the efficiency. For the base case cycle, the highest exergy efficiency is found to be 57.6%, when SF1 is 0.60 and SF2 is 0.96. It should be pointed out that the values of SF1 and SF2 which maximize the exergy efficiency are different for different ammonia mass fractions of the basic solution (1). Calculations are conducted to find out such optimum SF1 and SF2 pairs, and the results are reported in Table 4. It is noteworthy that when the ammonia mass fraction of the basic working fluid XBS is above 0.25, the optimum SF2 is found to be 1, the corresponding cycle configuration is the one shown in Fig. 3.
56.8 57.6 57.9 57.0 55.4
5. Performance comparison and discussion 5.1. Comparison with a system for separate generation of power and refrigeration In this section, the base case cycle (SF1 ¼ 0.60, SF2 ¼ 0.96) as a cogeneration system is compared with a separate generation system. The latter consists of a conventional steam Rankine cycle and an ammonia/water absorption refrigeration cycle. The main parameters and performance of the two systems are summarized in Table 5. As shown in Table 5, compared with the separate generation system with the same power and refrigeration outputs, the base case cogeneration cycle has a much better performance in terms of the exergy efficiency, which is enhanced by 22.3% relatively, and the energy consumption (which is proportional to mhs) is reduced by 18.2% as a result. The base case cycle performance improvement mainly attributes to the effective utilization of the heat source energy. Figs. 5 and 6 are the t–Q diagrams of the heat addition processes (the heat transfer process from the heat source fluid to the working fluid) in the separate generation system and the base case cycle, respectively. For the ammonia–water mixture in the base case cycle, the varied
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Table 5 Comparison between the cogeneration and the separate generation systems Cogeneration system (the base case cycle)
Working fluid Ammonia mass fraction of the basic working fluid Working fluid mass flow rate (kg/s) Turbine inlet temperature tT,i (1C) Turbine back pressure pb (MPa) Turbine power output (kW) Cycle output (kW) Refrigeration capacity QEVA Refrigeration exergy EEVA Net power output Wnet Heat source fluid Mass fluid rate mhs (kg/s) Inlet temperature t23 (1C) Exhaust temperature t26 (1C) Total heat source fluid mass flow rate mhs (kg/s) Total exergy input Ein (kW) Exergy efficiency y (%)
500 Tp
Water — 0.711 450 0.006 744.2
Ammonia/water 0.23 1.18 — — —
203.8 46.7 737.3
— — 737.3
203.8 46.7
7.496 465 90 7.496 1361.5 57.6
7.532 465 174.8 9.166 1664.8 47.1
(Refrigeration cycle)
Heating air(7.53kg/s)
Heating air (1.63kg/s)
500
REB HEX
Tp
NH3/H2O (1.18kg/s)
100
Tp
HEX2
Heating air(7.5kg/s)
Tp
300 NH3/H2O(0.81kg/s)
NH3/H2O(1kg/s)
100
Ambient state(25°C)
Ambient state(25°C)
0.0 0.4
NH3/H2O (1.2kg/s)
0
0 0.2
REB
200
H2O(0.71kg/s)
0.0
1.634 465 136.8
B
Tp
400
t (°C)
t (°C)
Refrigeration cycle
Ammonia/water 0.23 2 450 0.039 754.3
400
200
Rankine cycle
(Rankine cycle) B
300
Separate generation system
0.6
0.8
1.0
0.2
0.4
0.6
0.8
1.0
Q/Qhs
Q/Qhs
Fig. 5. The t–Q diagram of the heat addition process in the separate generation system.
boiling temperature leads to a better match with the heat source fluid in the boiling process. Moreover, the positions of the boiler B, the reboiler REB, and the heat exchanger HEX2 are arranged according to their temperature levels, which allows the external input heat to be utilized in a cascade way. Consequently, the heat source fluid energy in the base case cycle can be recovered more thoroughly and thus its exhaust temperature (90 1C) is much lower than that (136.8 1C) in the separate generation system. 5.2. Comparison with configuration A In this section, a comparison is conducted between the base case cycle (SF1 ¼ 0.60, SF2 ¼ 0.96) and configuration
Fig. 6. The t–Q diagram of the heat addition process in the base case cycle.
A (SF1 ¼ 0.72, SF2 ¼ 0.0) in Fig. 2. The comparison is made based on the same outputs as obtained for the aboveanalyzed base case cogeneration cycle. Configuration A is identical to the combined cycle proposed by Zheng et al. [7]. It is a modification from the Kalina cycle by replacing the flash tank with a rectifier REC. The strong solution from the top of the rectifier provides refrigeration, and the weak solution from the bottom helps to dilute the power cycle working fluid before condensation. Apparently, configuration A develops Kalina cycle into a power and refrigeration cogeneration system. A higher concentration of the ammonia/water mixture in boiling process is preferred to produce a better temperature
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profile. In both systems, stream 10 is the working fluid which extracts heat from the heating fluid in the boiler. In order to heighten the concentration of stream 10, the high concentration solution (9) is all sent into the absorber ABS2 in configuration A, while only a fraction of it is sent to ABS2 in the base case cycle. As a result, the concentration of stream 10 in the configuration A is higher than that in the base case cycle. To the configuration A, this will on one side improve the thermal match in the heat addition process; but on the other side will lead to the increase of the operation pressures in both ABS2 and EVA, and the backpressure of the turbine T. Consequently, not only the refrigeration output quantity but also its quality (exergy) would decrease, and the turbine specific power output (per unit working fluid mass flow rate) drops too. Higher working fluid mass flow rate thus is needed to produce the prescriptive refrigeration and power outputs. Configuration A is simulated with same main calculation assumptions in Table 1. In this calculation, the split fraction SF1 (0.72) and the turbine inlet pressure (16.8 MPa) chosen were the values that gave the highest exergy efficiency. The mass flow rate of the basic solution (stream 1) is 2.15 kg/s, which is higher than that in the base case cycle because of its low specific outputs. Table 6 shows the main parameters and performance of these two systems. The two systems have same power output Wnet (737.7 kW) and refrigeration capacity QEVA (203.8 kW). However, in the base case cogeneration
cycle, because of the lower concentration of stream 10, the evaporator EVA operates at a lower pressure and therefore produces refrigeration at a lower temperature level than that in configuration A. As shown in Table 6, the evaporator working fluid inlet/outlet temperatures are 6.5/1.6 1C in configuration A, and 34.4/15 1C in the base case cycle. Consequently, the refrigeration contributions to the exergy efficiency are very different in the two systems. From Table 6, the base case cycle produces 79.6% (20.7 kW) more refrigeration exergy, which accounts for 2.7% increase in the overall exergy outputs. In addition, the base case cycle needs less external heat input to produce the same power/refrigeration outputs, its energy consumption is lower by 2.4% (33.2 kW) than that of configuration A. An exergy analysis is conducted to disassemble the exergy destruction inside the two systems. The results reported in Table 7 only further verify the qualitatively analysis conclusions that: the base case cycle is superior to configuration A by consuming less exergy input while producing more exergy outputs. It is found that the base case cycle has an 8.5% (53.9 kW) reduction in the total exergy loss compared with that of configuration A, leading to a reduction of energy consumption (represented by the heating fluid mass flow rate) by 2.4%, this mainly attributes to the reduction of heat transfer related exergy destruction the heat addition process.
Table 6 Comparison between the cogeneration systems The configuration A (SF1 ¼ 0.72, SF2 ¼ 0.0) Basic working fluid concentration Working fluid mass flow rate (kg/s) Power subcycle Turbine working fluid mass flow rate mT (kg/s) Turbine inlet temperature tT,i (1C) Turbine inlet pressure pT,i (MPa) Turbine back pressure pb (MPa) Turbine exhaust vapor fraction Turbine power output (kW) Refrigeration subcycle Pressure of the evaporator (MPa) Temperature of evaporator inlet/outlet (1C) Mass concentration of refrigrant Mass flow rate of refrigerant (kg/s) Cycle output Refrigeration output QEVA (kW) Refrigeration exergy EEVA (kW) Net power output Wnet (kW) Heat source fluid Mass fluid rate mhs (kg/s) Inlet temperature t23 (1C) Exhaust temperature t26 (1C) Total exergy input Ein (kW) Exergy efficiency y (%)
0.23 2.148
The base case cycle (SF1 ¼ 0.6, SF2 ¼ 0.96) 0.23 2
0.852 450 16.8 0.075 0.88 766.1
0.808 450 11.1 0.039 0.88 754.3
0.289 6.5/1.6 0.92 0.251
0.082 34.4/-15 0.92 0.195
203.8 26 737.3
203.8 46.7 737.3
7.69 465 90 1394.7 54.7
7.496 465 90 1361.5 57.6
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Table 7 Exergy balance of the two cogeneration systems
Exergy input Exergy loss distribution Heat addition process Boiler B Rectifier unit (CON, REC, REB) Heat exchanger HEX2 Total
The configuration A (SF1 ¼ 0.72, SF2 ¼ 0.0)
The base case cycle (SF1 ¼ 0.6, SF2 ¼ 0.96)
(kW)
(%)
(kW)
(%)
1394.7
100
1361.5
100
109.3 127.8 27.3 264.4
7.84 9.16 1.96 18.96
113.5 88.3 14.7 216.5
8.34 6.49 1.08 15.91
— 27.1 157.7 184.8
— 1.94 11.31 13.25
18.6 1.1 157.3 177
1.37 0.08 11.55 13
Heat exchange process Recuperator REP Preheater PRE Heat exchanger HEX1 Cooler C Total
3.4 1.5 5.7 0.3 10.9
0.24 0.11 0.41 0.02 0.78
4.9 8.6 3.3 1.8 18.6
0.36 0.63 0.24 0.13 1.37
Pressure change process Pump P1,P2,P3 Valve V1,V2,V3 Total
22.2 2.5 24.7
1.59 0.18 1.77
9.3 10 19.3
0.68 0.73 1.42
Evaporator EVA Turbine T Splitter SPL1,SPL2 Flue exhaust Total system exergy loss
0 98.2 0 48.4 631.4
0 7.04 0 3.47 45.27
0 99 0 47.1 577.5
0 7.27 0 3.46 42.42
737.3 26 763.3
52.86 1.86 54.73
737.3 46.7 784
54.15 3.43 57.58
Absorption/condensation process Absorber ABS1 Absorber ABS2 Absorber ABS3 Total
Exergy output Power Refrigeration Total
B
500 Tp
REB
HEX2
reboiler REB and the heat exchanger HEX2. Furthermore, the exergy loss in each component is proportional to its working fluid mass flow rate, and configuration A has higher working fluid mass fluid rate because of its lower specific outputs. The calculation results in Table 7 demonstrates that the base case cycle has 18.1% (47.9 kW) less exergy loss in the heat addition process compared to that in the configuration A.
Heating air(7.69kg/s)
t (°C)
400
300 NH3/H2O(0.85kg/s) 200
6. Conclusion NH3/H2O(1.3kg/s)
100
Ambient state(25°C)
NH3/H2O (1.55kg/s)
0 0.0
0.2
0.4
0.6
0.8
1.0
Q/Qhs Fig. 7. The t–Q diagram of the heat addition process in configuration A.
Fig. 7 is the t–Q diagram of the heat addition process in configuration A. It is inferior to the base case cycle (Fig. 6) mainly in respect of the thermal match in the
Cogeneration has improved sustainability as it can improve the energy utilization efficiency significantly. In this paper, a novel ammonia–water combined cycle is proposed for power/refrigeration cogeneration. A splitting/ absorption unit is inserted into the combination of an ammonia-water Rankine subcycle and an ammonia refrigeration subcycle, and the condenser in the Rankine subcycle is replaced by an absorber. These ameliorations increase the working fluid ammonia concentration in the heat addition process and lower it in the
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M. Liu, N. Zhang / Energy 32 (2007) 961–970
absorption–condensation process. The irreversibility of both processes could thus be reduced and the cycle performance could be improved remarkably. The exergy efficiency is used as the evaluation criterion of the cycle performance, and it is found to be 58% for the base case study. The hardware used is conventional and commercially available; no hardware additional to that needed in conventional power and absorption cycles is needed. The proposed cogeneration system is merely a conceptual design in this paper. A detailed engineering analysis is needed for a more comprehensive evaluation. A study is conducted to investigate the effect of the split fractions on the cycle performance, and it is found that there is an optimum pair of the two split fractions, SF1 and SF2, at which the maximum exergy efficiency is obtained, and their values are varied for different basic working fluid concentrations. Compared with the separate generation system and the configuration A with the same outputs, this novel cogeneration system has a 22.3% and 5.3% increase in exergy efficiency respectively, and an 18.2% and 2.4% reduction in the energy consumption respectively.
Acknowledgment The authors gratefully acknowledge the support of the Chinese Natural Science Foundation Project (No. 50576096).
References [1] Maloney JD, Robertson RC. Thermodynamic study of ammoniawater heat power cycles. Oak Ridge National Laboratory report 1953 CF-53-8-43. [2] Kalina AI. Combined cycle system with novel bottoming cycle. J Eng Gas Turbines Power 1984;106:737–42. [3] Goswami DY. Solar thermal technology: present status and ideas for the future. Energy Sources 1998;20:137–45. [4] Goswami DY, Xu F. Analysis of a new thermodynamic cycle for combined power and cooling using low and mid temperature solar collectors. J Solar Energy Eng 1999;121:91–7. [5] Lu S, Goswami DY. Optimization of a novel combined power/ refrigeration thermodynamic cycle. J Solar Energy Eng 2003;125:212–7. [6] Vijayaraghavang S, Goswami DY. On evaluation efficiency of a combined power and cooling cycle. J Energy Resour Technol 2003;125:221–7. [7] Zheng DX, Chen B, Qi Y. Thermodynamic analysis of a novel absorption power/cooling combined cycle. J Eng Thermophys 2002;23:201–4. [8] Zhang N, Cai RX, Lior N. A novel ammonia–water cycle for power and refrigeration cogeneration. ASME international mechanical engineering congress & RD&D Expo, Anaheim, November 13–19, 2004. [9] Liu M, Zhang N, Cai RX. A series connected ammonia absorption power/cooling combined cycle. J Eng Thermophys 2006;27(1):9–12. [10] Liu M, Zhang N, Cai RX. Series connected ammonia–water refrigeration/power combined cycle and its sensitivity analysis. Proc CSEE 2006;26(1):1–8. [11] Zhang N, Lior N. Configuration selection methodology for combined power/refrigeration generation ammonia–water cycles. Proc ECOS, Trondheim 2005:953–60. [12] Aspen Pluss. Aspen Technology, Inc., version 11.1. See also:http:// www.aspentech.com/. [13] International Institute of Refrigeration, Thermodynamic and Physical Properties NH3–H2O, 1994, Paris, France.
Energy 32 (2007) 961–970 www.elsevier.com/locate/energy
Proposal and analysis of a novel ammonia–water cycle for power and refrigeration cogeneration Meng Liua,b, Na Zhanga, a
Institute of Engineering Thermophysics, Chinese Academy of Sciences Graduate School of the Chinese Academy of Sciences, PO Box 2706, Beijing 100080, P.R. China
b
Abstract Cogeneration has improved sustainability as it can improve the energy utilization efficiency significantly. In this paper, a novel ammonia-water cycle is proposed for the cogeneration of power and refrigeration. In order to meet the different concentration requirements in the cycle heat addition process and the condensation process, a splitting /absorption unit is introduced and integrated with an ammonia–water Rankine cycle and an ammonia refrigeration cycle. This system can be driven by industrial waste heat or a gas turbine flue gas. The cycle performance was evaluated by the exergy efficiency, which is 58% for the base case system (with the turbine inlet parameters of 450 1C/11.1 MPa and the refrigeration temperature below 15 1C). It is found that there are certain split fractions which maximize the exergy efficiency for given basic working fluid concentration. Compared with the conventional separate generation system of power and refrigeration, the cogeneration system has an 18.2% reduction in energy consumption. r 2006 Elsevier Ltd. All rights reserved. Keywords: Power/refrigeration cogeneration; Ammonia–water; Exergy efficiency
1. Introduction Binary component mixtures exhibit a boiling temperature that varies during the boiling process, and their employment as working fluids thus allows maintenance of a more constant temperature difference between them and variable temperature heat sources, and consequently reduced exergy losses in the heat addition process. To take advantage of the feature, Maloney and Robertson [1] first introduced the use of an ammonia/water mixture as the working fluid in an absorption power cycle in the early 1950s. However, the cycle condensation process takes place at a variable temperature too, resulting in a higher turbine backpressure (or lower turbine pressure ratio) than that of the conventional steam Rankine cycle. Higher backpressure is good to prevent air inleakage into the system, but unfavourable to the power generation and efficiency. Kalina proposed a combined power cycle which employs an ammonia/water mixture as the bottoming cycle working Corresponding author. Tel.: +86 10 82543030; fax: +86 10 62575913.
E-mail address: [email protected] (N. Zhang). 0360-5442/$ - see front matter r 2006 Elsevier Ltd. All rights reserved. doi:10.1016/j.energy.2006.09.012
fluid, and solved this problem by replacing the condensation process with an absorption process [2]. A combined cycle for power/refrigeration cogeneration was proposed by Goswami [3], and some further researches on the cycle performance were carried out [4–6]. This cycle employs the ammonia-rich vapor separated by a rectifier unit as the turbine working fluid to generate power, and then the turbine exhaust provides cooling by transferring sensible heat to the chilled water. Only about 16% of the total working fluid is used to generate power and refrigeration, and the turbine backpressure has to be maintained at a rather high level (0.2 MPa) to condense the working fluid in the absorption process. Zheng et al. [7] also proposed a combined cycle based on the Kalina cycle, which replaces the flash tank with a rectification column, the outflow from the top of the rectifier is throttled by a valve and then produces refrigeration before mixing the main stream. An energy efficiency of 24.2% and an exergy efficiency of 37.3% were reported with the turbine inlet parameters of 350 1C/5 MPa. In integrated refrigeration/power combined cycles, generally a basic refrigeration subcycle and a basic power
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Nomenclature
Subscripts
E e h m p pb Q RR SF s T t Wnet X DTp y
a ambient state Bs basic solution cw cooling water EVA evaporator HEX heat exchanger hs heat source fluid i inlet in input o outlet REB reboiler REC rectifier T turbine 0 base case 1,2,y26 states on the cycle flow sheet
exergy (kW) specific exergy (kJ/kg) specific enthalpy (kJ/kg) mass flow rate (kg/s) pressure (MPa) turbine backpressure (MPa) heat duty (kW) rectifier molar reflux ratio split fraction specific entropy (kJ/kg K) temperature (K) temperature (1C) net power output (kW) ammonia mass fraction (kg ammonia/kg mixture) pinch point temperature difference (K) exergy efficiency (%)
subcycle can be identified, and their integration can be accomplished by different configurations. The authors and colleagues have proposed and analyzed two basic cycle configurations: the parallel- and the series-connected ones [8–10]. It is found that both have some advantages and disadvantages [11]. In the parallel system, the working fluid in the power subcycle has a fairly low concentration, hence the desired temperature glide in the power subcycle boiling process does not occur, but the turbine may have a low backpressure. In the series-connected cycle, the working fluid has a higher ammonia concentration in the heat addition process and thus a better temperature glide and heat transfer match, but the turbine backpressure is high, thus reducing power output and increasing the exergy loss in the condensation process. The ideal system should have low exergy losses in both the heat addition and condensation processes. Obviously, different ammonia concentrations should be maintained in these processes. In this paper, a more advanced cycle configuration is proposed and analyzed, as an improvement upon the series-connected configuration [9]. A splitting/absorption unit is employed to adjust some stream mass flow rates and thus maintain the desired ammonia concentrations in different processes. Some different cycle configurations are available too by varying the split fractions.
2. The base case cycle configuration description Fig. 1 shows the base case cycle configuration. The refrigeration subcycle can be identified as (5-6-7-8-9-9A/ 9B-9C), and the power subcycle as (10–18). A splitting/ absorption unit (including SPL1, SPL2, ABS1, ABS2 and ABS3) is inserted in between, to accomplish the absorption and condensation processes, and at the same time to adjust some stream mass flow rates for meeting the different
concentration requirements in the cycle heat addition process and the condensation process. An ammonia–water mixture, referred as the basic concentration solution (1) is split into two streams (1A, 1B) by the splitter SPL1. Stream (1A) is heated and sent to the rectifier REC, where it is separated into a high concentration vapor (almost pure ammonia) and a weak concentration solution (19). The high concentration vapor from the top of the rectifier REC is condensed and partially sent to the refrigeration subcycle (5–9) to provide refrigeration. It is then split into two streams (9A, 9B) by the splitter SPL2, and sent to the absorbers ABS1 and ABS2, respectively. Stream (1B) is the main working fluid in the power subcycle. Before being heated in the recuperator REP, it is enriched by mixing with the strong solution (9C) in the absorber ABS2. This helps to improve the heat transfer match in the subsequent boiling process in the boiler B. The splitters SPL1 and SPL2 thus function to adjust both the concentration and the mass flow rate of stream (10), and thereby affect the power output. Moreover, the mass flow rate of the refrigerant (5) and the related refrigeration output would be varied too along with the variation of SPL1 split fraction. In the power subcycle, stream (10) is pumped to the system high pressure level (11) and evaporated and superheated to the highest cycle temperature (14). It expands to generate power in the turbine T. The turbine exhaust (15) is cooled, and then diluted by merging with the weak concentration solution (22) in the absorber ABS3. Consequently, the power subcycle condensation process is replaced by an absorption process, and the related exergy destruction could be reduced with a possibly lower condensation pressure. The working fluid has five ammonia mass concentration levels as shown in Fig. 1, and the combined cycle also has four pressure levels.
ARTICLE IN PRESS M. Liu, N. Zhang / Energy 32 (2007) 961–970
963
8 C
5
6
V1
EVA
7
SPL2
9 CON
9B
9A HEX2
4
REC
25 19 REB
24
13
3
HEX1
26 20
2
1A SPL1
1B
22 14 T
23
16
15 REP
V2
18
V3
B
1 ABS1
21
12
PRE
P1
P3 ABS3
17
P2 11
10
ABS2
9C
High concentration solution (refrigeration subcycle) Intermediate concentration solution (power subcycle) Basic concentration solution Intermediate concentration solution (power subcycle) Low concentration solution Heat source fluid Cooling water
ABS - Absorber HEX - Heat exchanger REP - Recuperator
B - Boiler P - Pump SPL- Splitter
C - Cooler PRE - Preheater T-Turbine
CON - Condenser REB - Reboiler V - Valve
EVA - Evaporator REC - Rectifier
Fig. 1. The flow sheet of the combined power/refrigeration cycle.
In this cycle, two improvements can be identified comparing with the series-connected configuration in [9]. Firstly, the splitting/absorption unit brings a stronger solution into the power subcycle heat addition process, thus making a better temperature match with the heat source fluid. Meanwhile, the mass flow rate of the working fluid in the refrigeration subcycle can be adjusted too. Secondly, the turbine exhaust is diluted in the absorption-condensation process, allowing a lower turbine backpressure and thus less exergy destruction during the condensation process. The cycle can be heated by the flue gas of a gas turbine or any other industrial waste heat with a suitable temperature. The heat source fluid (23) flows through the hot sides of the boiler B (24), the reboiler REB (25) and the heat exchanger HEX2 (26) in turn, and is eventually exhausted to the environment. This novel cycle has two useful outputs, power and refrigeration. Obviously the power and refrigeration outputs and the heat input are different in energy quality, hence the exergy efficiency is chosen to be the criterion for the cycle performance evaluation.
The exergy efficiency is defined as the exergy outputs divided by the exergy input to the cycle: y¼
W net þ E EVA , E in
(1)
where the net power output Wnet is the power output from the turbine, reduced by the power input to the pumps, and the exergy of the heat source fluid, Ein, is given as E in ¼ mhs ½ðhhs;i hhs;a Þ T a ðshs;i shs;a Þ.
(1a)
Since the heat source fluid is finally exhausted to the environment, the calculation of the exergy input is based on the difference between its initial state and the environment state. EEVA is the exergy associated with the refrigeration output, which is calculated as the working fluid exergy difference across the evaporator EVA. E EVA ¼ mEVA;i ½ðhEVA;i hEVA;o Þ T a ðsEVA;i sEVA;o Þ. (1b)
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Table 1 Main assumptions for the base-case calculation Items
Parameters
Values
Cycle parameter
Ammonia mass fraction of the basic solution XBs Cooling water temperature tcw (1C) SF1 SF2 Pressure loss (%) Refrigeration temperature upper limit (1C) Pressure loss (%) Temperature ta (1C) Pressure pa (MPa) Inlet temperature tT,i (1C) Isentropic efficiency (%) Minimum exhaust vapor fraction Theoretical stage number Molar reflux ratio RR Operation pressure pREC (MPa) Pressure loss (%) Outlet temperature tREB,o (1C) Pinch point temperature difference DTp (K) Pressure loss (%) Efficiency (%) Air Inlet temperature ths,i (1C) The minimal exhaust temperature (1C)
0.23 30 0.6 0.96 3 15 3 25 0.101 450 87 0.88 6 0.3 1.4 3 165 5 or 15 (if one side is air) 1.0–3.0 75 79% N2 and 21% O2 (by volume) 465 90
Split fraction Evaporator EVA Absorber ABS Ambient state Turbine T
Rectifier REC
Reboiler REB Heat exchangers (ABS, B, C, CON, HEX, PRE, REB, REP) Pumps P Heat source fluid
The split fractions of the splitters SPL1, SPL2 are defined
3. The base case cycle performance as It is assumed that the system operates at steady state. The simulation was carried out using the commercial Aspen Plus [12] code, in which the component models are based on energy and mass balances, with the default relative convergence error tolerance of 0.01%; the thermal property were calculated with the Electrolyte NRTL model or the SR-Polar model for high temperature (4246 1C) and pressure (410 MPa) application. To validate the properties calculations, the property results from Aspen+ and the data published by the International Institute of Refrigeration [13] were compared, and the result shows good agreement between them. The distillation column has been simulated by the RadFrac model, which is designed for simulating multistage vapour–liquid fractionation operation and is suitable for two-phase system. The main assumptions are shown in Table 1. The basic concentration solution (1) has the ammonia mass fraction of 0.23 and the mass flow rate of 2.0 kg/s. The gas turbine flue gas is one of the possible heat sources of this cycle. Since the small/middle sized gas turbine exhaust temperature values at 500–550 1C typically, the heating fluid inlet (23) and minimal outlet (26) temperatures of the proposed system are chosen to be 450 and 90 1C, respectively. The working fluid temperature at the absorber ABS2 outlet is set to be 35 1C in order to keep the minimal heat transfer temperature difference (5 K) with the external cooling water (30 1C).
SF 1 ¼ m1A =m1 ,
(2)
SF 2 ¼ m9A =m9 .
(3)
Table 2 summarizes the parameters of some main streams, including temperature t, pressure p, vapor fraction and ammonia mass fraction X, of the base case cycle configuration. The base case cycle performance is reported in Table 3. Keeping the pinch point temperature of 15 1C in the heat addition process, the heat source fluid mass flow rate should be 7.5 kg/s, and it exits the system at 90 1C (the prescribed minimal exhaust temperature). The system exergy efficiency is found to be 57.6%. 4. The effect of the split fractions 4.1. The effect of the split fractions on the cycle configuration By varying the split fractions of SPL1 and SPL2, some other cycle configurations are available as follows: (A) configuration A (0oSF1o1, SF2 ¼ 0): a power/ refrigeration cogeneration system as shown in Fig. 2, in which the strong solution from the refrigeration subcycle is all sent to ABS2 to enrich the power subcycle working fluid. As a result, the irreversibility
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Table 2 The base case cycle stream states
C
5
State No.
t (1C)
p (MPa)
965
Vapor fraction
m (kg/s)
V1
6
EVA
7
X
9
CON
1 1A 4 7 8 9 9A 10 14 15 17 19 22
37 37 134.4 34.4 15 24.2 24.2 35 450 68 35 165 47.5
0.077 0.077 1.442 0.084 0.082 0.079 0.079 0.075 11.10 0.039 0.037 1.400 0.038
0 0 0 0.1 0.79 0.91 0.91 0 1 0.88 0 0 0
2 1.2 1.2 0.195 0.195 0.195 0.187 0.808 0.808 0.808 1.813 1.005 1.005
0.23 0.23 0.23 0.92 0.92 0.92 0.92 0.237 0.237 0.237 0.159 0.096 0.096
HEX2
4
REC
25 19 REB 24
13
HEX1
3 26 20
21
12
V3
PRE
B
P1
2
1A SPL1
1
1B
22 14 T
23
P3
ABS3
16
15 REP
17
P2
11
ABS2
10
Fig. 2. Configuration A of the novel cycle (0oSF1o1, SF2 ¼ 0). Table 3 The base case cycle performance summary Parameters Turbine T work (kW) Pump work (kW) P1 P2 P3 Recuperator REP heat load (kW) Absorber heat load (kW) ABS1 ABS2 ABS3 Heat exchangers heat input (kW) HEX1 HEX2 PRE Condenser CON load (kW) Cooler C load (kW) Boiler B heat input (kW) Reboiler REB heat input (kW) Total heat input Qin (kW) Refrigeration capacity QEVA (kW) Refrigeration exergy EEVA (kW) Net power output Wnet (kW) Exergy input Ein (kW) Exergy efficiency y (%)
Values
8
754.3 CON
2.6 14.3 0.1 89.6 315.7 21.5 1717.1
HEX2
25 19 REB
24
6
V1
13
3
HEX1
26 20
21
12
V3
PRE
B
would decrease in the boiling process but would increase in the condensation process; (B) configuration B (0oSF1o1, SF2 ¼ 1): a power/refrigeration cogeneration system as shown in Fig. 3, in which the strong solution from EVA is all sent to ABS1. This configuration has lower irreversibility in the condensation process but not in the boiling process; (C) configuration C (SF1 ¼ 0): a regenerative Rankine type power cycle; (D) configuration D (SF1 ¼ 1, SF2 ¼ 1): a conventional ammonia absorption refrigeration cycle.
23
2
P1
1A SPL1
1 ABS1
1B
18
22
14
141 641.5 457.4 362.4 61.5 1926.7 381.6 3444.4 203.8 46.7 737.3 1361.5 57.6
EVA
7
9 4
REC
C
5
T
16
15 REP
11
P3 ABS3
17
P2
Fig. 3. Configuration B of the novel cycle (0oSF1o1, SF2 ¼ 1).
4.2. The effect of the split fractions on the cycle performance The split fractions SF1, SF2 are the key parameters to the cycle performance, their effects are investigated and the results are shown in Figs. 4a and b. To accomplish the condensation process in the absorbers ABS1 and ABS2, their operation pressures need to be maintained above certain values for given cooling water temperature. This imposes a low limit on the operation pressure of the evaporator EVA. When SF1 increases or SF2 decreases, the ammonia concentration of stream 10 would increase, requiring higher absorption pressure in ABS2 for its condensation. Accordingly, the EVA operation pressure need to be elevated too, and so does the produced refrigeration temperature level. When the EVA working fluid inlet temperature t7 increases to 15 1C, no refrigeration would be produced. In this calculation, t7 is set to be under 21 1C to ensure certain
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58
1.2 (%)
EEVA/EEVA,0
1.4 1.0 0.8
54
0.6 SF1=0.50 SF1=0.60 SF1=0.70
0.4 1.4 1.2
52
1.0 0.8 0.6
1.2 1.0 0.8
0.0
0.2
0.4
0.6
0.8
0.0
1.0
SF2
(a)
SF1=0.50 SF1=0.60 SF1=0.70
1.4 mhs/mhs,0
Wnet/Wnet,0
56
0.2
0.4
0.6
0.8
1.0
SF2
(b)
Fig. 4. (a) The effects of the split fractions on the normalized net power output and refrigeration exergy output. (b) The effects of split fractions on the normalized heating fluid mass flow rate and the exergy efficiency.
Table 4 The effect of X
Bs
on the optimum split fractions and the optimized exergy efficiency
Ammonia mass fraction of the basic solution, XBs
XBs ¼ 0.20 XBs ¼ 0.23 (base case) XBs ¼ 0.25 XBs ¼ 0.27 XBs ¼ 0.30
Optimum split fractions
The optimized exergy efficiency (%)
SF1
SF2
0.66 0.6 0.57 0.58 0.59
0.71 0.96 1 1 1
amount refrigeration output be produced, which requires SF2 be above certain value for higher SF1 application. Fig. 4a shows the effect of the split fractions on the net power output Wnet and the refrigeration exergy output EEVA. When SF1 increases, more working fluid is sent to the refrigeration subcycle, therefore the refrigeration exergy output increases; when SF2 increases, less working fluid is sent to the power subcycle, therefore and the power output Wnet drops. At the same time, the needed external heat amount drops too, as represented by the heating fluid mass flow rate mhs in Fig. 4b. It is found that for given SF1, there is a SF2 which maximizes the efficiency. For the base case cycle, the highest exergy efficiency is found to be 57.6%, when SF1 is 0.60 and SF2 is 0.96. It should be pointed out that the values of SF1 and SF2 which maximize the exergy efficiency are different for different ammonia mass fractions of the basic solution (1). Calculations are conducted to find out such optimum SF1 and SF2 pairs, and the results are reported in Table 4. It is noteworthy that when the ammonia mass fraction of the basic working fluid XBS is above 0.25, the optimum SF2 is found to be 1, the corresponding cycle configuration is the one shown in Fig. 3.
56.8 57.6 57.9 57.0 55.4
5. Performance comparison and discussion 5.1. Comparison with a system for separate generation of power and refrigeration In this section, the base case cycle (SF1 ¼ 0.60, SF2 ¼ 0.96) as a cogeneration system is compared with a separate generation system. The latter consists of a conventional steam Rankine cycle and an ammonia/water absorption refrigeration cycle. The main parameters and performance of the two systems are summarized in Table 5. As shown in Table 5, compared with the separate generation system with the same power and refrigeration outputs, the base case cogeneration cycle has a much better performance in terms of the exergy efficiency, which is enhanced by 22.3% relatively, and the energy consumption (which is proportional to mhs) is reduced by 18.2% as a result. The base case cycle performance improvement mainly attributes to the effective utilization of the heat source energy. Figs. 5 and 6 are the t–Q diagrams of the heat addition processes (the heat transfer process from the heat source fluid to the working fluid) in the separate generation system and the base case cycle, respectively. For the ammonia–water mixture in the base case cycle, the varied
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Table 5 Comparison between the cogeneration and the separate generation systems Cogeneration system (the base case cycle)
Working fluid Ammonia mass fraction of the basic working fluid Working fluid mass flow rate (kg/s) Turbine inlet temperature tT,i (1C) Turbine back pressure pb (MPa) Turbine power output (kW) Cycle output (kW) Refrigeration capacity QEVA Refrigeration exergy EEVA Net power output Wnet Heat source fluid Mass fluid rate mhs (kg/s) Inlet temperature t23 (1C) Exhaust temperature t26 (1C) Total heat source fluid mass flow rate mhs (kg/s) Total exergy input Ein (kW) Exergy efficiency y (%)
500 Tp
Water — 0.711 450 0.006 744.2
Ammonia/water 0.23 1.18 — — —
203.8 46.7 737.3
— — 737.3
203.8 46.7
7.496 465 90 7.496 1361.5 57.6
7.532 465 174.8 9.166 1664.8 47.1
(Refrigeration cycle)
Heating air(7.53kg/s)
Heating air (1.63kg/s)
500
REB HEX
Tp
NH3/H2O (1.18kg/s)
100
Tp
HEX2
Heating air(7.5kg/s)
Tp
300 NH3/H2O(0.81kg/s)
NH3/H2O(1kg/s)
100
Ambient state(25°C)
Ambient state(25°C)
0.0 0.4
NH3/H2O (1.2kg/s)
0
0 0.2
REB
200
H2O(0.71kg/s)
0.0
1.634 465 136.8
B
Tp
400
t (°C)
t (°C)
Refrigeration cycle
Ammonia/water 0.23 2 450 0.039 754.3
400
200
Rankine cycle
(Rankine cycle) B
300
Separate generation system
0.6
0.8
1.0
0.2
0.4
0.6
0.8
1.0
Q/Qhs
Q/Qhs
Fig. 5. The t–Q diagram of the heat addition process in the separate generation system.
boiling temperature leads to a better match with the heat source fluid in the boiling process. Moreover, the positions of the boiler B, the reboiler REB, and the heat exchanger HEX2 are arranged according to their temperature levels, which allows the external input heat to be utilized in a cascade way. Consequently, the heat source fluid energy in the base case cycle can be recovered more thoroughly and thus its exhaust temperature (90 1C) is much lower than that (136.8 1C) in the separate generation system. 5.2. Comparison with configuration A In this section, a comparison is conducted between the base case cycle (SF1 ¼ 0.60, SF2 ¼ 0.96) and configuration
Fig. 6. The t–Q diagram of the heat addition process in the base case cycle.
A (SF1 ¼ 0.72, SF2 ¼ 0.0) in Fig. 2. The comparison is made based on the same outputs as obtained for the aboveanalyzed base case cogeneration cycle. Configuration A is identical to the combined cycle proposed by Zheng et al. [7]. It is a modification from the Kalina cycle by replacing the flash tank with a rectifier REC. The strong solution from the top of the rectifier provides refrigeration, and the weak solution from the bottom helps to dilute the power cycle working fluid before condensation. Apparently, configuration A develops Kalina cycle into a power and refrigeration cogeneration system. A higher concentration of the ammonia/water mixture in boiling process is preferred to produce a better temperature
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profile. In both systems, stream 10 is the working fluid which extracts heat from the heating fluid in the boiler. In order to heighten the concentration of stream 10, the high concentration solution (9) is all sent into the absorber ABS2 in configuration A, while only a fraction of it is sent to ABS2 in the base case cycle. As a result, the concentration of stream 10 in the configuration A is higher than that in the base case cycle. To the configuration A, this will on one side improve the thermal match in the heat addition process; but on the other side will lead to the increase of the operation pressures in both ABS2 and EVA, and the backpressure of the turbine T. Consequently, not only the refrigeration output quantity but also its quality (exergy) would decrease, and the turbine specific power output (per unit working fluid mass flow rate) drops too. Higher working fluid mass flow rate thus is needed to produce the prescriptive refrigeration and power outputs. Configuration A is simulated with same main calculation assumptions in Table 1. In this calculation, the split fraction SF1 (0.72) and the turbine inlet pressure (16.8 MPa) chosen were the values that gave the highest exergy efficiency. The mass flow rate of the basic solution (stream 1) is 2.15 kg/s, which is higher than that in the base case cycle because of its low specific outputs. Table 6 shows the main parameters and performance of these two systems. The two systems have same power output Wnet (737.7 kW) and refrigeration capacity QEVA (203.8 kW). However, in the base case cogeneration
cycle, because of the lower concentration of stream 10, the evaporator EVA operates at a lower pressure and therefore produces refrigeration at a lower temperature level than that in configuration A. As shown in Table 6, the evaporator working fluid inlet/outlet temperatures are 6.5/1.6 1C in configuration A, and 34.4/15 1C in the base case cycle. Consequently, the refrigeration contributions to the exergy efficiency are very different in the two systems. From Table 6, the base case cycle produces 79.6% (20.7 kW) more refrigeration exergy, which accounts for 2.7% increase in the overall exergy outputs. In addition, the base case cycle needs less external heat input to produce the same power/refrigeration outputs, its energy consumption is lower by 2.4% (33.2 kW) than that of configuration A. An exergy analysis is conducted to disassemble the exergy destruction inside the two systems. The results reported in Table 7 only further verify the qualitatively analysis conclusions that: the base case cycle is superior to configuration A by consuming less exergy input while producing more exergy outputs. It is found that the base case cycle has an 8.5% (53.9 kW) reduction in the total exergy loss compared with that of configuration A, leading to a reduction of energy consumption (represented by the heating fluid mass flow rate) by 2.4%, this mainly attributes to the reduction of heat transfer related exergy destruction the heat addition process.
Table 6 Comparison between the cogeneration systems The configuration A (SF1 ¼ 0.72, SF2 ¼ 0.0) Basic working fluid concentration Working fluid mass flow rate (kg/s) Power subcycle Turbine working fluid mass flow rate mT (kg/s) Turbine inlet temperature tT,i (1C) Turbine inlet pressure pT,i (MPa) Turbine back pressure pb (MPa) Turbine exhaust vapor fraction Turbine power output (kW) Refrigeration subcycle Pressure of the evaporator (MPa) Temperature of evaporator inlet/outlet (1C) Mass concentration of refrigrant Mass flow rate of refrigerant (kg/s) Cycle output Refrigeration output QEVA (kW) Refrigeration exergy EEVA (kW) Net power output Wnet (kW) Heat source fluid Mass fluid rate mhs (kg/s) Inlet temperature t23 (1C) Exhaust temperature t26 (1C) Total exergy input Ein (kW) Exergy efficiency y (%)
0.23 2.148
The base case cycle (SF1 ¼ 0.6, SF2 ¼ 0.96) 0.23 2
0.852 450 16.8 0.075 0.88 766.1
0.808 450 11.1 0.039 0.88 754.3
0.289 6.5/1.6 0.92 0.251
0.082 34.4/-15 0.92 0.195
203.8 26 737.3
203.8 46.7 737.3
7.69 465 90 1394.7 54.7
7.496 465 90 1361.5 57.6
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Table 7 Exergy balance of the two cogeneration systems
Exergy input Exergy loss distribution Heat addition process Boiler B Rectifier unit (CON, REC, REB) Heat exchanger HEX2 Total
The configuration A (SF1 ¼ 0.72, SF2 ¼ 0.0)
The base case cycle (SF1 ¼ 0.6, SF2 ¼ 0.96)
(kW)
(%)
(kW)
(%)
1394.7
100
1361.5
100
109.3 127.8 27.3 264.4
7.84 9.16 1.96 18.96
113.5 88.3 14.7 216.5
8.34 6.49 1.08 15.91
— 27.1 157.7 184.8
— 1.94 11.31 13.25
18.6 1.1 157.3 177
1.37 0.08 11.55 13
Heat exchange process Recuperator REP Preheater PRE Heat exchanger HEX1 Cooler C Total
3.4 1.5 5.7 0.3 10.9
0.24 0.11 0.41 0.02 0.78
4.9 8.6 3.3 1.8 18.6
0.36 0.63 0.24 0.13 1.37
Pressure change process Pump P1,P2,P3 Valve V1,V2,V3 Total
22.2 2.5 24.7
1.59 0.18 1.77
9.3 10 19.3
0.68 0.73 1.42
Evaporator EVA Turbine T Splitter SPL1,SPL2 Flue exhaust Total system exergy loss
0 98.2 0 48.4 631.4
0 7.04 0 3.47 45.27
0 99 0 47.1 577.5
0 7.27 0 3.46 42.42
737.3 26 763.3
52.86 1.86 54.73
737.3 46.7 784
54.15 3.43 57.58
Absorption/condensation process Absorber ABS1 Absorber ABS2 Absorber ABS3 Total
Exergy output Power Refrigeration Total
B
500 Tp
REB
HEX2
reboiler REB and the heat exchanger HEX2. Furthermore, the exergy loss in each component is proportional to its working fluid mass flow rate, and configuration A has higher working fluid mass fluid rate because of its lower specific outputs. The calculation results in Table 7 demonstrates that the base case cycle has 18.1% (47.9 kW) less exergy loss in the heat addition process compared to that in the configuration A.
Heating air(7.69kg/s)
t (°C)
400
300 NH3/H2O(0.85kg/s) 200
6. Conclusion NH3/H2O(1.3kg/s)
100
Ambient state(25°C)
NH3/H2O (1.55kg/s)
0 0.0
0.2
0.4
0.6
0.8
1.0
Q/Qhs Fig. 7. The t–Q diagram of the heat addition process in configuration A.
Fig. 7 is the t–Q diagram of the heat addition process in configuration A. It is inferior to the base case cycle (Fig. 6) mainly in respect of the thermal match in the
Cogeneration has improved sustainability as it can improve the energy utilization efficiency significantly. In this paper, a novel ammonia–water combined cycle is proposed for power/refrigeration cogeneration. A splitting/ absorption unit is inserted into the combination of an ammonia-water Rankine subcycle and an ammonia refrigeration subcycle, and the condenser in the Rankine subcycle is replaced by an absorber. These ameliorations increase the working fluid ammonia concentration in the heat addition process and lower it in the
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absorption–condensation process. The irreversibility of both processes could thus be reduced and the cycle performance could be improved remarkably. The exergy efficiency is used as the evaluation criterion of the cycle performance, and it is found to be 58% for the base case study. The hardware used is conventional and commercially available; no hardware additional to that needed in conventional power and absorption cycles is needed. The proposed cogeneration system is merely a conceptual design in this paper. A detailed engineering analysis is needed for a more comprehensive evaluation. A study is conducted to investigate the effect of the split fractions on the cycle performance, and it is found that there is an optimum pair of the two split fractions, SF1 and SF2, at which the maximum exergy efficiency is obtained, and their values are varied for different basic working fluid concentrations. Compared with the separate generation system and the configuration A with the same outputs, this novel cogeneration system has a 22.3% and 5.3% increase in exergy efficiency respectively, and an 18.2% and 2.4% reduction in the energy consumption respectively.
Acknowledgment The authors gratefully acknowledge the support of the Chinese Natural Science Foundation Project (No. 50576096).
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