- Email: [email protected]
Design and modeling of low temperature solar thermal power station
Applied Energy 91 (2012) 180–186
Contents lists available at SciVerse ScienceDirect
Applied Energy journal homepage: www.elsevier.com/locate/apenergy
Design and modeling of low temperature solar thermal power station N. Shankar Ganesh, T. Srinivas ⇑ Energy Division, School of Mechanical and Building Sciences, Vellore Institute of Technology (VIT) University, Vellore 632 014, Tamil Nadu, India
a r t i c l e
i n f o
Article history: Received 16 March 2011 Received in revised form 19 August 2011 Accepted 12 September 2011 Available online 13 October 2011 Keywords: Ammonia–water mixture Heat recovery Kalina cycle Solar energy Thermodynamic analysis
a b s t r a c t During the heat recovery in a Kalina cycle, a binary aqua–ammonia mixture changes its state from liquid to vapor, the more volatile ammonia vaporizes first and then the water starts vaporization to match temperature profile of the hot fluid. In the present work, a low temperature Kalina cycle has been investigated to optimize the heat recovery from solar thermal collectors. Hot fluid coming from solar parabolic trough collector with vacuum tubes is used to generate ammonia rich vapor in a boiler for power generation. The turbine inlet conditions are optimized to match the variable hot fluid temperature with the intermittent nature of the solar radiation. The key parameters discussed in this study are strong solution concentration, separator temperature which affects the hot fluid inlet temperature and turbine ammonia concentration. Solar parabolic collector system with vacuum tubes has been designed at the optimized power plant conditions. This work can be used in the selection of boiler, separator and turbine conditions to maximize the power output as well as efficiency of power generation system. The current model results a maximum limit temperature for separator as 150 °C at the Indian climatic conditions. A maximum specific power of 105 kW per kg/s of working fluid can be obtained at 80% of strong solution concentration with 140 °C separator temperature. The corresponding plant and cycle efficiencies are 5.25% and 13% respectively. But the maximum efficiencies of 6% and 15% can be obtained respectively for plant and Kalina cycle at 150 °C of separator temperature. Ó 2011 Elsevier Ltd. All rights reserved.
1. Introduction The Kalina cycle is a thermodynamic cycle for converting thermal energy to mechanical power which utilizes working fluid comprised of at least two different components and a ratio between those components is varied in different parts of the system to increase thermodynamic reversibility and therefore increase overall thermody namic efficiency. The Kalina cycle was first invented by the Russian Engineer Alexander Kalina [1]. When a binary aqua–ammonia mixture changes its state from liquid to vapor, the ammonia having low boiling point vaporizes first and then the water. Hence, a better match with the temperature profile of the flue gas ensues. Basically this concept is suitable for low and medium temperature heat recovery systems. Waste heat recovery plants using the Kalina cycle technology are in operation at Sumitomo metal steel works and Fuji oil’s refinery in Tokyo Bay. Geothermal plants exist in Husavik, Iceland, and Unterhaching, Germany, recently built by Siemens. The Kalina cycle trademark and patents are owned by Global Geothermal Ltd., the parent of Recurrent Engineering Inc. Sayed and Tribus [2] compared the performance of Kalina cycle with the Rankine cycle. But the configurations developed by them were very much complicated because several heat exchangers had
⇑ Corresponding author. Tel.: +91 924 4557787; fax: +91 416 2243092. E-mail address: [email protected] (T. Srinivas). 0306-2619/$ - see front matter Ó 2011 Elsevier Ltd. All rights reserved. doi:10.1016/j.apenergy.2011.09.021
more than two streams. Later, Marston [3] modified the Sayed and Tribus configuration with a simple two stream heat exchangers. Rogdakis and Antonopolos [4] showed that for fixed upper (i.e. superheating) and lower (i.e. condensation) temperatures, the Kalina cycle shows 20% higher efficiency than that of the Rankine cycle. This cycle was also optimized by Rogdakis [5] who developed correlations describing the optimum operation of the cycle. Thermodynamic properties of ammonia–water system developed by Ziegler and Trepp [6] are used in the current work to develop the thermodynamic model. In 1992, a Kalina demonstration plant started operation at the US Department of Energy’s Energy Technology Engineering Center in California. At first the plant used waste heat at a temperature of approximately 540 °C to generate 3 MWe of power and accumulated 5200 operation hours. In this plant the maximum pressure and temperature of the Kalina cycle were 110 bar and 516 °C [7]. The Kalina cycle turbine was less expensive than the steam turbine, since the volumetric flow rate in the low pressure part of the ammonia–water turbine was much smaller than in the steam turbine, while the Kalina cycle heat exchangers were more expensive. However, the additional power output of the Kalina cycle gave an economic benefit compared with the steam cycle. Power plants using conventional processes and unconventional fluids have a significant potential for the valorization of low and medium temperature renewable energy sources as well as waste heat from industrial, commercial or institutional installations.
N. Shankar Ganesh, T. Srinivas / Applied Energy 91 (2012) 180–186
181
Nomenclature A AP f F h L m N PP T w W x
g /
area, m2 approach point focal length, m vapor mass fraction in separator specific enthalpy, kJ/kg length, m mass, kg number pinch point, K temperature, K specific work output, kJ/kg width of collector, m mass fraction of ammonia, kg/kg mixture efficiency rim angle
bp c g ge fi h KC m p ps rs tro t tot w
bubble point collector global generator fluid inlet hot Kalina cycle mechanical pump parallel segments reflective sheet tracking, receiver and other turbine total water
Subscripts a ambient b beam
Recently researchers are paying interest towards Kalina power generation due to the suitability of this system in conventional and non-conventional areas of energy. Tamm and Goswami [8,9] and Hasan and Goswami [10] proposed a combined power/refrigeration cycle using ammonia–water as the mixed working fluids, and investigated its performance. Tamm et al. [11] showed the feasibility of the vapor generation and absorption condensation processes experimentally. Zheng et al. [12] proposed an absorption power/cooling combined cycle and a thermodynamic analysis of the cycle is performed using log p–T, log p–h and T–s diagrams. Valdimarsson and Eliasson [13] showed the cost benefit of Kalina cycle over organic Rankine cycle (ORC) and found from their contour diagrams, that the best power and best-cost points are different. The potential use of non-conventional fluids in Rankine cycles and the Kalina cycle, to improve the performance with respect to conventional single and dual flash steam power plants has been proved by Desideri and Bidini [14]. Dejfors and Svedberg [15] showed that the vapor absorption cycle has a higher net power output and consequently a lower total exergy loss compared to other Rankine cycles. Hettiarachchi et al. [16] analyzed the performance of the Kalina cycle system for low-temperature geothermal heat sources and is compared with an ORC performance. They investigated the effect of the ammonia fraction and turbine inlet pressure on the cycle performance. Murugan and Subbarao [17] performed the exergy and energy analysis for the proposed Rankine–Kalina combined cycle. Srinivas et al. [18] studied the heat recovery from gas turbine exhaust with Kalina bottoming cycle and also highlighted the benefit over the steam bottoming cycle. Galanis et al. [19] reviewed some prototypes of power plants with unconventional fluids (refrigerants, CO2, binary mixtures) and summarized some of the relevant scientific and technical work. Solar energy is a clean, environmental friendly energy source for power generation, and it results higher efficiency and low cost compared to solar photovoltaic systems at certain conditions [20]. Recently more efficient way of heat recovery technologies are developed from hot gases or solar thermic fluid. Chacartegui et al. [21] developed ORC for combined bottoming cycles. These developments can be shared with other power generators, like thermal solar facilities. Cayer et al. [22] developed the carbon dioxide transcritical cycles for the lower temperature heat recovery. Baik et al. [23] compared carbon dioxide and R125 transcritical cycles for a low grate heat source and proved that the R125
transcritical cycle produced 14% more power than the carbon dioxide transcritical cycle. Wang et al. [24] proposed a combined power and refrigeration cycle, which combines the Rankine cycle and the absorption refrigeration cycle with binary ammonia–water mixture as the working fluid. The main objective of the current work is to examine the effect of the separator temperature which influences the source inlet temperature at fixed degree of superheat and terminal temperature difference, strong solution concentration, turbine concentration and solar direct beam radiation on performance of the solar power plant. A parametric investigation has been carried out to maximize the efficiency and power output. This work also focused on cost minimization of solar parabolic collectors with the heat source conditions. 2. Thermodynamic modeling and analysis of solar thermal power plant Assumptions used in the thermodynamic analysis: Atmospheric condition is taken as 1.01325 bar and 25 °C. Terminal temperature difference (TTD) at heat recovery vapor generator (boiler) inlet with respect to the collector’s hot fluid is taken at 5 °C. Pinch point (PP) in boiler is 5 °C. Approach point (AP) in the boiler is 10 °C. The isentropic efficiency of solution pump and mixture turbine is considered as 75%. The mechanical efficiency of the solution pump (gm,p) and mixture turbine (gm,t) is taken at 96%. Electrical generator efficiency (gge) is taken as 98%. The condensate leaving the condenser is assumed as saturated liquid. Pressure drop and heat loss in pipe lines are neglected. The schematic flow diagram of the Kalina cycle with hot fluid (glycol based water) as heat source coming from a solar concentrating collectors is shown in Fig. 1. The heat from the hot fluid (14) coming from solar collectors is recovered in the superheater and boiler. In separator, the working fluid is separated into rich ammonia water vapor (10) and weak liquid mixture (11). The rich ammonia water vapor’s temperature is increased in the superheater
182
N. Shankar Ganesh, T. Srinivas / Applied Energy 91 (2012) 180–186
is the vapor portion and (1 F) kg/s is the liquid portion to be separated. After applying lever rule for separation process,
Array of solar collectors 16
F¼
T 8 ¼ T bp AP
10 9
1
11
MXT
HT RGN
12 THR 13
8 Circulating pump 6 CFP
2 MXR
T 16 ¼ T bp þ PP
LT RGN CND
5
CW in
G
3
7
ð4Þ
The unknown properties i.e. temperature, concentration and mass flow rates are determined by mass, concentration and energy balance equations. The turbine exit temperature can be determined by entropy equalization for isentropic expansion and the actual temperature by the isentropic efficiency relation. The same can be applied for solution pump also. The energy interactions in the plant components are: Mixture turbine output,
4
CW out
Fig. 1. Schematic flow diagram of the Kalina cycle with hot fluid coming from solar concentrating collectors; CFP: condensate feed pump; CND: condenser; CW: cooling water; HT: high temperature; LT: low temperature; MXR: mixture; MXT: mixture turbine; RGN: regenerator; SEP: separator; SH: superheater; THR: throttling.
W t ¼ m1 ðh1 h2 Þgm;t gge kW
before entering the inlet of the turbine (1). The vapor (1) is expanded in mixture turbine to generate power and it is diluted with a weak solution (13). The liquid weak solution coming from separator has been throttled (12–13) after rejecting heat (11–12) at high temperature regenerator and mixed with turbine exit fluid (2). The mixture (3) again rejects heat (3–4) at low temperature regenerator and condenses to a saturated liquid state (5). The condensate is pumped to separator pressure (6) and heated in a low temperature regenerator (6–7) and high temperature regenerator (7–8). The preheated liquid mixture (8) is converted into liquid vapor mixture (9) in the evaporator of the boiler. The saturated vapor (10) is heated in a superheater before entering into the mixture turbine. This cycle repeats for continuous power generation. Hot fluid exit temperature from the collectors,
ð1Þ
where the temperature, T1 is maintained about the separator temperature which is equal to the degree of superheat. The working fluid at the separator exit (10) is in a saturated vapor condition. The high pressure (P10) can be determined from the separator temperature and vapor concentration (T10, x10), since it is the function of temperature and concentration at the saturated vapor state. Other wise if there is no superheater, the separator pressure (high pressure at turbine inlet) is determined from the inlet conditions of turbine i.e. ammonia concentration (x1) and temperature (T1). The separator temperature, T11 at the liquid outlet is the bubble point temperature at the high pressure and concentration, x11. From this relation only, the liquid portion concentration, x11 can be determined through the iteration. The Kalina cycle is solved with 1 kg/s of strong solution at the separator inlet (9). In separator, out of one kg/s of mixture, F kg/s
ð5Þ
Work input to pump,
Wp ¼
T 14 ¼ T 1 þ TTDSH
ð3Þ
where Tbp is the bubble point temperature at boiler pressure and concentration. The low pressure is determined from mixture ammonia concentration (x5) and temperature (T5) at condenser outlet. The temperature of working fluid between low temperature and high temperature regenerator, T7 is maintained below the temperature, T3. The hot fluid temperature at the evaporator section of boiler,
SEP
Boiler
ð2Þ
To avoid the evaporation in the high temperature regenerator, the boiler liquid inlet temperature is maintained below the bubble point temperature of the solution. The temperature of strong solution at the boiler inlet,
Hot fluid from collectors 14
SH
15
x9 x11 x10 x11
m5 ðh6 h5 Þ
gm;p
kW
ð6Þ
Net output from Kalina cycle,
W net ¼ ðW t W p Þ kW
ð7Þ
Heat supply in boiler,
Q boiler ¼ m8 ðh9 h8 Þ
ð8Þ
Heat supply in superheater,
Q SH ¼ m1 ðh1 h10 Þ
ð9Þ
Design features of parabolic trough collector with vacuum tube at the focal line (specifications are taken from the manufacturer’s data):
Width of the collector;W ¼ 3 m Parabolic rim angle;/rim ¼ 80 2
The focal length is calculated from parabola equation, y ¼ x4f at 80° rim angle.
Focal length;f ¼
W 4 tanð0:5/rim Þ
ð10Þ
Number of parallel sections, nps = 2. The arc length for parabolic reflector sheet,
/ / / / tan þ ln sec þ tan Larc ¼ 2f sec 2 2 2 2
ð11Þ
The outlet temperature of hot fluid in the boiler, Amount of hot fluid to be circulated in the solar concentrating collectors for 1 kg/s of strong solution,
N. Shankar Ganesh, T. Srinivas / Applied Energy 91 (2012) 180–186
mh ¼ m14 ¼
m8 ðh9 h8 Þ þ m10 ðh1 h10 Þ kg=s cph ðT 14 T 16 Þ
ð12Þ
Mass flow rate of hot fluid in each row of parallel collectors,
mh;ps ¼
mh nps
ð13Þ
where nps is number of parallel sections. The collector length in row can be determined from the collector’s efficiency and outlet temperature of the fluid. Parabolic trough collector efficiency [25],
T fi T a Ib
gc ¼ 0:642 0:441
ð14Þ
Length of each parallel line,
L¼
m14 cph ðT 14 T 16 Þ gc Ib Wnps
ð15Þ
Total area of collection,
Ac
tot
¼ nps WL
ð16Þ
Kalina cycle efficiency,
gKC ¼
W net 100 Q boiler þ Q SH
ð17Þ
Solar plant energy efficiency,
g1 ¼
W net 100 Ig Ac tot
ð18Þ
Cost of solar parabolic concentrating collector: Cost of reflective sheet, Crs = Rs. 1000/m2 surface area. Cost of tracking, receiver tube and other equipments, Ctro = Rs. 25,000/m length. Total cost of the collectors,
C tot ¼ C rso nps Larc L þ C tro nps L
ð19Þ
3. Results and discussion The performance of solar thermal power plant is thermodynamically investigated under the operation conditions. The influences of the turbine inlet concentration, strong solution concentration, separator temperature and solar beam radiation have been examined on the performance, solar concentrator area and cost.
183
Fig. 2 shows the effect of strong solution concentration (0.65– 0.8) and separator temperature (110–150 °C) on (a) solar plant efficiency Fig. 2 shows the effect of strong solution concentration (0.65–0.8) and separator temperature (110–150 °C) on (a) solar plant efficiency – specific power and (b) Kalina cycle efficiency – specific power. Under the specified limits of operating conditions, the plant results 2.5–6% of solar power plant efficiency, 6–15% cycle efficiency and 30–110 kW of power output. The maximum possible temperature for separator is found as 150 °C from the parametric simulation at the Indian climatic conditions. Considering the degree of super heat and terminal temperature difference in the boiler, the required hot fluid supply temperature is 170 °C at this conditions. The plant efficiency increases with an increment in the separator temperature up to the maximum value. At 0.65 strong solution concentration, the efficiency maximizes at 140 °C of the separator temperature. The optimum separator temperature increases with an increase in solution concentration. The power output maximizes with 125 °C, 130 °C, 135 °C and 140 °C respectively at 0.65, 0.7, 0.75 and 0.8 solution concentration. They are 60 kW, 75 kW, 85 kW and 105 kW respectively for one kg/s of strong solution flow rate. With the increase in the separator temperature, the high pressure i.e. turbine inlet pressure increases which increases the turbine work. Simultaneously the pump input also increases due to a rise in high pressure at a fixed low pressure. The vapor fraction at the separator decreases with an increase in the separator temperature. On overall basis, the net output increases with the turbine effect and then decreases due to pump work and vapor fraction changes. The maximum power conditions are different from the maximum efficiency conditions. The maximum power can be obtained relatively at lower separator temperature compared to the maximum efficiency conditions. The low pressure i.e. turbine exit pressure increases with the increase in strong solution concentration at a fixed high pressure. It decreases the turbine expansion and also decreases the pump supply. The vapor fraction at the separator increases with increase in the strong solution concentration. These factors lead to increase the net output at a variable phase. Increase in strong solution results higher efficiency only at the high separator temperature i.e. 120 °C. The efficiency decreases with an increase in the strong solution concentration below this temperature (120 °C). On an average basis the specific power output increases from 0.5 to 1.5 kW per 1 °C rise in the separator temperature. Similarly, there is a 0.1% rise in plant efficiency and 0.25% rise in the cycle efficiency per 1 °C rise in the separator temperature. Hettiarachchi et al. [16] reported power output of 60–70 kW per kg/s of strong solution at a heat source
Fig. 2. Variation of plant and cycle performance with strong solution concentration and separator temperature at the turbine inlet concentration of 0.95.
184
N. Shankar Ganesh, T. Srinivas / Applied Energy 91 (2012) 180–186
Fig. 3. Influence of strong solution and turbine inlet concentrations on performance of plant at the separator temperature of 125 °C.
temperature of 90 °C. These are given at a PP of 2 °C. The current model results 50–70 kW per kg/s at 5 °C PP and 110 °C of separator temperature. Pouraghaie et al. [26] resulted the power outputs from 60 to 110 kW per kg/s at the superheater temperature of 125–225 °C. Lolos and Rogdakis [27] gave a turbine output of 216.7 kJ/kg vapor for maximum work condition at the superheated temperature of 130 °C. The current model results 198 kJ/kg vapor at the same superheated temperature. This is lower than the reported result due to the higher sink temperature in the current model compared to the reported sink temperature of 20 °C. They resulted the cycle efficiency from 8.1 to 8.3% and the same for the current model is 8.6%. Fig. 3a and b depicts the effect of strong solution concentration and turbine inlet concentration on the performance of the solar thermal power plant. The efficiencies (plant and cycle) are increasing with an increase in turbine concentration. As per Bombarda et al. [28] results, the optimum strong solution concentration increases with an increase in separator temperature to get a maximum net power output. The current work results the maximum power outputs at 70 kW, 76 kW, 83 kW and 93 kW per kg/s of strong solution respectively at the strong solution concentrations of 0.65, 0.7, 0.75 and 0.8. Turbine inlet concentration influences the high pressure but not the low pressure. Similarly, the strong solution concentration effects the low pressure only. A rise in
turbine concentration increases the high pressure and decreases both the vapor fraction and heat supply in the boiler. The combined influence of these parameters, increases the net power output and then decreases with a rise in turbine inlet concentration. It also causes a rise in plant efficiency. A rise in strong solution concentration increases the low pressure, vapor fraction and heat supply. All these parameters increases the power output with a rise in strong solution concentration but the rise in efficiency takes place above the 0.95 concentration only. The specific power output increases from 1.5 to 3 kW per 1% rise in the strong solution concentration. There is a 0.2% rise in the plant efficiency and 1% rise in the cycle efficiency with a 1% rise in the turbine inlet concentration. Fig. 4 has been plotted for (a) plant performance and (b) cost of solar collectors at various beam radiations and separator temperatures. The solar direct beam radiation is varied from 400 W/m2 to 700 W/m2 as it varies from place to place and also with the time. The separator temperature is varied from 110 °C to 150 °C. The results are plotted based on the fixation of cycle conditions with a suitable collector area as per the available beam radiation. The plant results a maximum power at 135 °C of separator temperature and 0.75 strong solution concentration. The efficiency increases with an increment in sun beam radiation. The solar collector efficiency increases with an increase in beam radiation even through the cycle efficiency is kept constant. The specific power
Fig. 4. Influence of solar beam radiation with separator temperature on plant performance and collectors cost at the strong solution concentration of 0.75.
185
N. Shankar Ganesh, T. Srinivas / Applied Energy 91 (2012) 180–186 Table 1 Solar thermal power plant material flow details with respect to Fig. 1 at Indian climatic conditions and Tsep = 125 °C. State
Pressure (bar)
Temperature (°C)
Ammonia concentration
Flow rate (kg/s)
Specific enthalpy (kJ/kg)
1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16
35.6 7.5 7.5 7.5 7.5 35.6 35.6 35.6 35.6 35.6 35.6 35.6 7.5 – – –
135.0 66.4 57.6 53.4 30.0 31.8 44.6 82.5 125.0 125.0 125.0 49.6 50.1 145.0 143.9 92.5
0.95 0.95 0.70 0.70 0.70 0.70 0.70 0.70 0.70 0.95 0.47 0.47 0.47 – – –
0.48 0.48 1.00 1.00 1.00 1.00 1.00 1.00 1.00 0.48 0.52 0.52 0.52 3.47 3.47 3.47
1545.2 1363.6 651.9 590.1 55.5 45.2 14.6 198.1 943.8 1513.3 342.2 14.3 14.4 501.6 497.2 282.1
Table 2 Specifications of the solar thermal power plant at the hot fluid inlet temperature of 145 °C. Description
Result
Vapor fraction in separator, % Heat load in boiler, kW Heat load in superheater, kW Heat load in low temperature regenerator, kW Heat load in high temperature regenerator, kW Heat load in condenser, kW Work output of mixture turbine, kW Work input to solution pump, kW Net electricity output, kW Kalina cycle energy efficiency, % Solar plant efficiency, % Total length of receiver tubes, m Total area of collectors, m2 Cost of collectors, Rs. in lakhs
48 746 15 62 184 646 86.6 10.6 76 10 4 2722 2042 193
Table 3 Comparison of the present work with the existing plant readings at 75% turbine efficiency (Tsep = 120 °C, x1 = 0.95, T5 = 12 °C, Welectric = 1.82 MW). Sl. no. 1 2 3 4 5 6 7 8 9 10 11 12 12 13 14
Description
Plant result
Predicted result
Hot water requirement, kg/s Hot water inlet temperature, °C Hot water outlet temperature, °C Separator pressure, bar Low pressure, bar Temperature of working fluid at boiler inlet, °C Strong solution concentration Separated liquid concentration Vapor in separator, kg/s Ammonia–water mixture before separation, kg/s Temperature after expansion, °C Temperature after pumping, °C Turbine output, kW Pump input, kW Cycle energy efficiency, %
90 124 80 31 5.5 67
74.2 127 77 30.2 5.3 67
0.81 0.5 11.2 16.3
0.8 0.45 10.6 15.1
60 13 1950 130 11.8
55 14.8 1968 148 11.7
output is constant with the solar radiation as it depends on the cycle conditions in the current study. The variations in collector area, collector cost with separator temperature and direct beam radiation have been plotted in Fig. 4b. The economic conditions are considered with the government subsidy of Rs. 5500/m2 of collection area. At the strong solution of 0.75, the plant efficiency increases with the separator temperature. Therefore the minimum cost can be obtained at higher solar radiation and separator temperature.
The collector cost obtained is matched with the manufacturers cost quotation with the government subsidy. At the predefined conditions and a change in the parameters, 2.5–6% plant efficiency, 65–95 kW of specific electric power, 1500–5000 m2 collector area and Rs. 150 lakhs to Rs. 450 lakhs of collectors cost are resulted with the parametric variations. Table 1 shows the properties of the working fluid and hot fluid at the state points, defined in Fig. 1. The results are plotted at the separator temperature of 125 °C, 0.7 strong solution concentration, 1 kg/s of working fluid and the turbine concentration of 0.95. It results 145° hot fluid inlet temperature and 135 °C separator temperature. A unit mass of working fluid in the power circuit demands 3.47 units of hot fluid at the above said conditions. It also shows 48% of vapor fraction in the separator. The unknown properties at the state points are determined from the equation outlined from Eqs. (1)–(13). Table 2 gives the results of the solar thermal power plant at the operating conditions stated in earlier sections. The rating of heat exchangers, power, efficiency and cost details are developed at the unit mass of the working fluid. These specifications are developed at the same conditions defined for Table 1. The heat load in the high temperature regenerator is nearly three times more than the low temperature regenerator. The present thermodynamic evaluation has been validated by comparing the existing Kalina cycle power plant located in Husavik, Iceland running from hot water (geo thermal resource). Table 3 compares the present simulated results with the plant readings [29]. The plant and current model are solved with the working conditions i.e. at strong solution concentration = 0.81, T15 = 121 °C, x1 = 0.95 and t5 = 12 °C. The considered real power plant is operating without a superheater; therefore in the present simulated model the superheater is bypassed for this study. Most of the calculated results in this work are closely matched with the Husavik power plant conditions. The strong solution concentration is considered at 0.8 in place of the Husavik value of 0.81. The resulted low pressure is 5 bar where as 5.5 bar is the plant reading. The simulated separator liquid concentration is comparatively weak (0.45) over the plant reading (0.5). These results also matched with the literature results reported by Nasruddin et al. [30] who reported 11.4% of energy net efficiency. 4. Conclusions A solar thermal power plant based on Kalina power system has been modeled thermodynamically and analyzed parametrically. The performance of the plant has been studied at the separator temperature of 110–150 °C, strong solution concentration of
186
N. Shankar Ganesh, T. Srinivas / Applied Energy 91 (2012) 180–186
0.65–0.8, turbine inlet concentration of 0.92–0.955 and solar beam radiation of 400–700 W/m2. The optimum separator temperature is different for maximum efficiency condition and for maximum power condition. The maximum power can be obtained relative at lower separator temperature compared to the maximum efficiency conditions at a fixed strong solution concentration. The turbine inlet concentration influences the high pressure whereas the strong solution concentration influences the low pressure. The minimum collector cost can be obtained with high solar beam radiation and optimized cycle conditions. The thermodynamic model has been validated by comparing the literature and the existed plant readings. References [1] Kalina IA. Combined cycle system with novel bottoming cycle. ASME J Eng Gas Turb Power 1984;106:737–42. [2] Sayed YMEI, Tribus M. A theoretical comparison of Rankine and Kalina cycles. ASME publication; 1985. AES: 1. [3] Marston CH. Parametric analysis of the Kalian cycle. ASME J Eng Gas Turb Power 1990;112:107–16. [4] Rogdakis ED, Antonopoulos KA. A high efficiency NH3–H2O absorption power cycle. Heat Recov Syst CHP 1991;11:263–75. [5] Rogdakis ED. Thermodynamic analysis, parametric study and optimum operation of the Kalina cycle. Int J Energy Res 1996;20(4):359–70. [6] Ziegler B, Trepp C. Equation of state for ammonia–water mixtures. Int J Refrig 1984;7(2):101–6. [7] Leibowitz H, Mirolli M. First Kalina combined cycle plant tested successfully. Power Eng 1997;101(5). [8] Tamm G, Goswami DY, Lu S, Hasan AA. Novel combined power and cooling thermodynamic cycle for low temperature heat sources, part 1: theoretical investigation. ASME J Sol Energy Eng 2003;125(2):218–22. [9] Tamm G, Goswami DY. Novel combined power and cooling thermodynamic cycle for low temperature heat sources, part 2: experimental investigation. ASME J Sol Energy Eng 2003;125(2):223–9. [10] Hasan AA, Goswami DY. Exergy analysis of a combined power and refrigeration thermodynamic cycle driven by a solar heat source. ASME J Sol Energy Eng 2003;125(1):55–60. [11] Tamm G, Goswami DY, Lu S, Hasan AA. Theoretical and experimental investigation of an ammonia–water power and refrigeration thermodynamic cycle. Sol Energy 2004;76(1–3):217–28. [12] Zheng D, Chen B, Qi Y, Jin H. Thermodynamic analysis of a novel absorption power/cooling combined-cycle. Appl Energy 2006;83(4):311–23.
[13] Valdimarsson P, Eliasson L. Factors influencing the economics of the Kalina power cycle and situations of superior performance. In: International geothermal conference 2003, Reykjavik; 2003. p. 32–40 Paper No. 79. [14] Desideri U, Bidini G. Study of possible optimization criteria for geothermal power plants. Energy Convers Manage 1997;38(15–17):1681–91. [15] Dejfors C, Svedberg G. Second law analysis of ammonia–water power cycle for direct-fired cogeneration application. Int J Appl Thermodyn 1999;2(3): 125–31. [16] Hettiarachchi HDM, Golubovic M, Worek WM, Ikegami Y. The performance of the Kalina cycle system 11 (KCS-11) with low-temperature heat sources. ASME J Energy Resour Technol 2007;129(3):243–7. [17] Murugan RS, Subbarao PMV. Thermodynamic analysis of Rankine–Kalina combined cycle. Int J Thermodyn 2008;11(3):133–41. [18] Srinivas T, Gupta AVSSKS, Reddy BV. Performance simulation of combined cycle with Kalina bottoming cycle. Cogener Distrib Gener J 2008;23(1):6–20. [19] Galanis N, Cayer E, Roy P, Denis ES, Desilets M. Electricity generation from low temperature sources. J Appl Fluid Mech 2009;2(2):55–67. [20] Hua E, Yang YP, Nishimura A, Yilmaz F, Kouzani A. Solar thermal aided power generation. Appl Energy 2010;87(9):2881–5. [21] Chacartegui R, Sanchez D, Muñoz JM, Sanchez T. Alternative ORC bottoming cycles for combined cycle power plants. Appl Energy 2009;86(10):2162–70. [22] Cayer E, Galanis N, Desilets M, Nesreddine H, Roy P. Analysis of a carbon dioxide transcritical power cycle using a low temperature source. Appl Energy 2009;86(7–8):1055–63. [23] Baik YJ, Kim M, Chang KC, Kim SJ. Power-based performance comparison between carbon dioxide and R125 transcritical cycles for a low-grade heat source. Appl Energy 2011;88(3):892–8. [24] Wang J, Dai Y, Gao L. Parametric analysis and optimization for a combined power and refrigeration cycle. Appl Energy 2008;85(11):1071–85. [25] Valan Arasu A, Sornakumar T. Design, manufacture and testing of fiberglass reinforced parabola trough for parabolic trough solar collectors. Sol Energy 2007;81(10):1273–9. [26] Pouraghaie M, Atashkari K, Besarati SM, Nariman-zadeh N. Thermodynamic performance optimization of a combined power/cooling cycle. Energy Convers Manage 2010;51(1):204–11. [27] Lolos PA, Rogdakis ED. A Kalina power cycle driven by renewable energy sources. Energy 2009;34(4):457–64. [28] Bombarda P, Invernizzi CM, Pietra C. Heat recovery from diesel engines: a thermodynamic comparison between Kalina and ORC cycles. Appl Therm Eng 2010;30(2–3):212–9. [29] Mirolli M, Hjartarson H, Mlcak H, Ralph M. Testing and operating experience of the 2 MW Kalina cycle geothermal power plant in Husavik, Iceland. In: Proceedings of the world renewable energy congress VII, Cologne, Germany; 2002. [30] Nasruddin, Usvika R, Rifaldi M, Noor A. Energy and energy analysis of Kalina cycle system (KCS) 34 with mass fraction ammonia–water mixture variation. J Mech Sci Technol 2009;23(7):1871–6.
Contents lists available at SciVerse ScienceDirect
Applied Energy journal homepage: www.elsevier.com/locate/apenergy
Design and modeling of low temperature solar thermal power station N. Shankar Ganesh, T. Srinivas ⇑ Energy Division, School of Mechanical and Building Sciences, Vellore Institute of Technology (VIT) University, Vellore 632 014, Tamil Nadu, India
a r t i c l e
i n f o
Article history: Received 16 March 2011 Received in revised form 19 August 2011 Accepted 12 September 2011 Available online 13 October 2011 Keywords: Ammonia–water mixture Heat recovery Kalina cycle Solar energy Thermodynamic analysis
a b s t r a c t During the heat recovery in a Kalina cycle, a binary aqua–ammonia mixture changes its state from liquid to vapor, the more volatile ammonia vaporizes first and then the water starts vaporization to match temperature profile of the hot fluid. In the present work, a low temperature Kalina cycle has been investigated to optimize the heat recovery from solar thermal collectors. Hot fluid coming from solar parabolic trough collector with vacuum tubes is used to generate ammonia rich vapor in a boiler for power generation. The turbine inlet conditions are optimized to match the variable hot fluid temperature with the intermittent nature of the solar radiation. The key parameters discussed in this study are strong solution concentration, separator temperature which affects the hot fluid inlet temperature and turbine ammonia concentration. Solar parabolic collector system with vacuum tubes has been designed at the optimized power plant conditions. This work can be used in the selection of boiler, separator and turbine conditions to maximize the power output as well as efficiency of power generation system. The current model results a maximum limit temperature for separator as 150 °C at the Indian climatic conditions. A maximum specific power of 105 kW per kg/s of working fluid can be obtained at 80% of strong solution concentration with 140 °C separator temperature. The corresponding plant and cycle efficiencies are 5.25% and 13% respectively. But the maximum efficiencies of 6% and 15% can be obtained respectively for plant and Kalina cycle at 150 °C of separator temperature. Ó 2011 Elsevier Ltd. All rights reserved.
1. Introduction The Kalina cycle is a thermodynamic cycle for converting thermal energy to mechanical power which utilizes working fluid comprised of at least two different components and a ratio between those components is varied in different parts of the system to increase thermodynamic reversibility and therefore increase overall thermody namic efficiency. The Kalina cycle was first invented by the Russian Engineer Alexander Kalina [1]. When a binary aqua–ammonia mixture changes its state from liquid to vapor, the ammonia having low boiling point vaporizes first and then the water. Hence, a better match with the temperature profile of the flue gas ensues. Basically this concept is suitable for low and medium temperature heat recovery systems. Waste heat recovery plants using the Kalina cycle technology are in operation at Sumitomo metal steel works and Fuji oil’s refinery in Tokyo Bay. Geothermal plants exist in Husavik, Iceland, and Unterhaching, Germany, recently built by Siemens. The Kalina cycle trademark and patents are owned by Global Geothermal Ltd., the parent of Recurrent Engineering Inc. Sayed and Tribus [2] compared the performance of Kalina cycle with the Rankine cycle. But the configurations developed by them were very much complicated because several heat exchangers had
⇑ Corresponding author. Tel.: +91 924 4557787; fax: +91 416 2243092. E-mail address: [email protected] (T. Srinivas). 0306-2619/$ - see front matter Ó 2011 Elsevier Ltd. All rights reserved. doi:10.1016/j.apenergy.2011.09.021
more than two streams. Later, Marston [3] modified the Sayed and Tribus configuration with a simple two stream heat exchangers. Rogdakis and Antonopolos [4] showed that for fixed upper (i.e. superheating) and lower (i.e. condensation) temperatures, the Kalina cycle shows 20% higher efficiency than that of the Rankine cycle. This cycle was also optimized by Rogdakis [5] who developed correlations describing the optimum operation of the cycle. Thermodynamic properties of ammonia–water system developed by Ziegler and Trepp [6] are used in the current work to develop the thermodynamic model. In 1992, a Kalina demonstration plant started operation at the US Department of Energy’s Energy Technology Engineering Center in California. At first the plant used waste heat at a temperature of approximately 540 °C to generate 3 MWe of power and accumulated 5200 operation hours. In this plant the maximum pressure and temperature of the Kalina cycle were 110 bar and 516 °C [7]. The Kalina cycle turbine was less expensive than the steam turbine, since the volumetric flow rate in the low pressure part of the ammonia–water turbine was much smaller than in the steam turbine, while the Kalina cycle heat exchangers were more expensive. However, the additional power output of the Kalina cycle gave an economic benefit compared with the steam cycle. Power plants using conventional processes and unconventional fluids have a significant potential for the valorization of low and medium temperature renewable energy sources as well as waste heat from industrial, commercial or institutional installations.
N. Shankar Ganesh, T. Srinivas / Applied Energy 91 (2012) 180–186
181
Nomenclature A AP f F h L m N PP T w W x
g /
area, m2 approach point focal length, m vapor mass fraction in separator specific enthalpy, kJ/kg length, m mass, kg number pinch point, K temperature, K specific work output, kJ/kg width of collector, m mass fraction of ammonia, kg/kg mixture efficiency rim angle
bp c g ge fi h KC m p ps rs tro t tot w
bubble point collector global generator fluid inlet hot Kalina cycle mechanical pump parallel segments reflective sheet tracking, receiver and other turbine total water
Subscripts a ambient b beam
Recently researchers are paying interest towards Kalina power generation due to the suitability of this system in conventional and non-conventional areas of energy. Tamm and Goswami [8,9] and Hasan and Goswami [10] proposed a combined power/refrigeration cycle using ammonia–water as the mixed working fluids, and investigated its performance. Tamm et al. [11] showed the feasibility of the vapor generation and absorption condensation processes experimentally. Zheng et al. [12] proposed an absorption power/cooling combined cycle and a thermodynamic analysis of the cycle is performed using log p–T, log p–h and T–s diagrams. Valdimarsson and Eliasson [13] showed the cost benefit of Kalina cycle over organic Rankine cycle (ORC) and found from their contour diagrams, that the best power and best-cost points are different. The potential use of non-conventional fluids in Rankine cycles and the Kalina cycle, to improve the performance with respect to conventional single and dual flash steam power plants has been proved by Desideri and Bidini [14]. Dejfors and Svedberg [15] showed that the vapor absorption cycle has a higher net power output and consequently a lower total exergy loss compared to other Rankine cycles. Hettiarachchi et al. [16] analyzed the performance of the Kalina cycle system for low-temperature geothermal heat sources and is compared with an ORC performance. They investigated the effect of the ammonia fraction and turbine inlet pressure on the cycle performance. Murugan and Subbarao [17] performed the exergy and energy analysis for the proposed Rankine–Kalina combined cycle. Srinivas et al. [18] studied the heat recovery from gas turbine exhaust with Kalina bottoming cycle and also highlighted the benefit over the steam bottoming cycle. Galanis et al. [19] reviewed some prototypes of power plants with unconventional fluids (refrigerants, CO2, binary mixtures) and summarized some of the relevant scientific and technical work. Solar energy is a clean, environmental friendly energy source for power generation, and it results higher efficiency and low cost compared to solar photovoltaic systems at certain conditions [20]. Recently more efficient way of heat recovery technologies are developed from hot gases or solar thermic fluid. Chacartegui et al. [21] developed ORC for combined bottoming cycles. These developments can be shared with other power generators, like thermal solar facilities. Cayer et al. [22] developed the carbon dioxide transcritical cycles for the lower temperature heat recovery. Baik et al. [23] compared carbon dioxide and R125 transcritical cycles for a low grate heat source and proved that the R125
transcritical cycle produced 14% more power than the carbon dioxide transcritical cycle. Wang et al. [24] proposed a combined power and refrigeration cycle, which combines the Rankine cycle and the absorption refrigeration cycle with binary ammonia–water mixture as the working fluid. The main objective of the current work is to examine the effect of the separator temperature which influences the source inlet temperature at fixed degree of superheat and terminal temperature difference, strong solution concentration, turbine concentration and solar direct beam radiation on performance of the solar power plant. A parametric investigation has been carried out to maximize the efficiency and power output. This work also focused on cost minimization of solar parabolic collectors with the heat source conditions. 2. Thermodynamic modeling and analysis of solar thermal power plant Assumptions used in the thermodynamic analysis: Atmospheric condition is taken as 1.01325 bar and 25 °C. Terminal temperature difference (TTD) at heat recovery vapor generator (boiler) inlet with respect to the collector’s hot fluid is taken at 5 °C. Pinch point (PP) in boiler is 5 °C. Approach point (AP) in the boiler is 10 °C. The isentropic efficiency of solution pump and mixture turbine is considered as 75%. The mechanical efficiency of the solution pump (gm,p) and mixture turbine (gm,t) is taken at 96%. Electrical generator efficiency (gge) is taken as 98%. The condensate leaving the condenser is assumed as saturated liquid. Pressure drop and heat loss in pipe lines are neglected. The schematic flow diagram of the Kalina cycle with hot fluid (glycol based water) as heat source coming from a solar concentrating collectors is shown in Fig. 1. The heat from the hot fluid (14) coming from solar collectors is recovered in the superheater and boiler. In separator, the working fluid is separated into rich ammonia water vapor (10) and weak liquid mixture (11). The rich ammonia water vapor’s temperature is increased in the superheater
182
N. Shankar Ganesh, T. Srinivas / Applied Energy 91 (2012) 180–186
is the vapor portion and (1 F) kg/s is the liquid portion to be separated. After applying lever rule for separation process,
Array of solar collectors 16
F¼
T 8 ¼ T bp AP
10 9
1
11
MXT
HT RGN
12 THR 13
8 Circulating pump 6 CFP
2 MXR
T 16 ¼ T bp þ PP
LT RGN CND
5
CW in
G
3
7
ð4Þ
The unknown properties i.e. temperature, concentration and mass flow rates are determined by mass, concentration and energy balance equations. The turbine exit temperature can be determined by entropy equalization for isentropic expansion and the actual temperature by the isentropic efficiency relation. The same can be applied for solution pump also. The energy interactions in the plant components are: Mixture turbine output,
4
CW out
Fig. 1. Schematic flow diagram of the Kalina cycle with hot fluid coming from solar concentrating collectors; CFP: condensate feed pump; CND: condenser; CW: cooling water; HT: high temperature; LT: low temperature; MXR: mixture; MXT: mixture turbine; RGN: regenerator; SEP: separator; SH: superheater; THR: throttling.
W t ¼ m1 ðh1 h2 Þgm;t gge kW
before entering the inlet of the turbine (1). The vapor (1) is expanded in mixture turbine to generate power and it is diluted with a weak solution (13). The liquid weak solution coming from separator has been throttled (12–13) after rejecting heat (11–12) at high temperature regenerator and mixed with turbine exit fluid (2). The mixture (3) again rejects heat (3–4) at low temperature regenerator and condenses to a saturated liquid state (5). The condensate is pumped to separator pressure (6) and heated in a low temperature regenerator (6–7) and high temperature regenerator (7–8). The preheated liquid mixture (8) is converted into liquid vapor mixture (9) in the evaporator of the boiler. The saturated vapor (10) is heated in a superheater before entering into the mixture turbine. This cycle repeats for continuous power generation. Hot fluid exit temperature from the collectors,
ð1Þ
where the temperature, T1 is maintained about the separator temperature which is equal to the degree of superheat. The working fluid at the separator exit (10) is in a saturated vapor condition. The high pressure (P10) can be determined from the separator temperature and vapor concentration (T10, x10), since it is the function of temperature and concentration at the saturated vapor state. Other wise if there is no superheater, the separator pressure (high pressure at turbine inlet) is determined from the inlet conditions of turbine i.e. ammonia concentration (x1) and temperature (T1). The separator temperature, T11 at the liquid outlet is the bubble point temperature at the high pressure and concentration, x11. From this relation only, the liquid portion concentration, x11 can be determined through the iteration. The Kalina cycle is solved with 1 kg/s of strong solution at the separator inlet (9). In separator, out of one kg/s of mixture, F kg/s
ð5Þ
Work input to pump,
Wp ¼
T 14 ¼ T 1 þ TTDSH
ð3Þ
where Tbp is the bubble point temperature at boiler pressure and concentration. The low pressure is determined from mixture ammonia concentration (x5) and temperature (T5) at condenser outlet. The temperature of working fluid between low temperature and high temperature regenerator, T7 is maintained below the temperature, T3. The hot fluid temperature at the evaporator section of boiler,
SEP
Boiler
ð2Þ
To avoid the evaporation in the high temperature regenerator, the boiler liquid inlet temperature is maintained below the bubble point temperature of the solution. The temperature of strong solution at the boiler inlet,
Hot fluid from collectors 14
SH
15
x9 x11 x10 x11
m5 ðh6 h5 Þ
gm;p
kW
ð6Þ
Net output from Kalina cycle,
W net ¼ ðW t W p Þ kW
ð7Þ
Heat supply in boiler,
Q boiler ¼ m8 ðh9 h8 Þ
ð8Þ
Heat supply in superheater,
Q SH ¼ m1 ðh1 h10 Þ
ð9Þ
Design features of parabolic trough collector with vacuum tube at the focal line (specifications are taken from the manufacturer’s data):
Width of the collector;W ¼ 3 m Parabolic rim angle;/rim ¼ 80 2
The focal length is calculated from parabola equation, y ¼ x4f at 80° rim angle.
Focal length;f ¼
W 4 tanð0:5/rim Þ
ð10Þ
Number of parallel sections, nps = 2. The arc length for parabolic reflector sheet,
/ / / / tan þ ln sec þ tan Larc ¼ 2f sec 2 2 2 2
ð11Þ
The outlet temperature of hot fluid in the boiler, Amount of hot fluid to be circulated in the solar concentrating collectors for 1 kg/s of strong solution,
N. Shankar Ganesh, T. Srinivas / Applied Energy 91 (2012) 180–186
mh ¼ m14 ¼
m8 ðh9 h8 Þ þ m10 ðh1 h10 Þ kg=s cph ðT 14 T 16 Þ
ð12Þ
Mass flow rate of hot fluid in each row of parallel collectors,
mh;ps ¼
mh nps
ð13Þ
where nps is number of parallel sections. The collector length in row can be determined from the collector’s efficiency and outlet temperature of the fluid. Parabolic trough collector efficiency [25],
T fi T a Ib
gc ¼ 0:642 0:441
ð14Þ
Length of each parallel line,
L¼
m14 cph ðT 14 T 16 Þ gc Ib Wnps
ð15Þ
Total area of collection,
Ac
tot
¼ nps WL
ð16Þ
Kalina cycle efficiency,
gKC ¼
W net 100 Q boiler þ Q SH
ð17Þ
Solar plant energy efficiency,
g1 ¼
W net 100 Ig Ac tot
ð18Þ
Cost of solar parabolic concentrating collector: Cost of reflective sheet, Crs = Rs. 1000/m2 surface area. Cost of tracking, receiver tube and other equipments, Ctro = Rs. 25,000/m length. Total cost of the collectors,
C tot ¼ C rso nps Larc L þ C tro nps L
ð19Þ
3. Results and discussion The performance of solar thermal power plant is thermodynamically investigated under the operation conditions. The influences of the turbine inlet concentration, strong solution concentration, separator temperature and solar beam radiation have been examined on the performance, solar concentrator area and cost.
183
Fig. 2 shows the effect of strong solution concentration (0.65– 0.8) and separator temperature (110–150 °C) on (a) solar plant efficiency Fig. 2 shows the effect of strong solution concentration (0.65–0.8) and separator temperature (110–150 °C) on (a) solar plant efficiency – specific power and (b) Kalina cycle efficiency – specific power. Under the specified limits of operating conditions, the plant results 2.5–6% of solar power plant efficiency, 6–15% cycle efficiency and 30–110 kW of power output. The maximum possible temperature for separator is found as 150 °C from the parametric simulation at the Indian climatic conditions. Considering the degree of super heat and terminal temperature difference in the boiler, the required hot fluid supply temperature is 170 °C at this conditions. The plant efficiency increases with an increment in the separator temperature up to the maximum value. At 0.65 strong solution concentration, the efficiency maximizes at 140 °C of the separator temperature. The optimum separator temperature increases with an increase in solution concentration. The power output maximizes with 125 °C, 130 °C, 135 °C and 140 °C respectively at 0.65, 0.7, 0.75 and 0.8 solution concentration. They are 60 kW, 75 kW, 85 kW and 105 kW respectively for one kg/s of strong solution flow rate. With the increase in the separator temperature, the high pressure i.e. turbine inlet pressure increases which increases the turbine work. Simultaneously the pump input also increases due to a rise in high pressure at a fixed low pressure. The vapor fraction at the separator decreases with an increase in the separator temperature. On overall basis, the net output increases with the turbine effect and then decreases due to pump work and vapor fraction changes. The maximum power conditions are different from the maximum efficiency conditions. The maximum power can be obtained relatively at lower separator temperature compared to the maximum efficiency conditions. The low pressure i.e. turbine exit pressure increases with the increase in strong solution concentration at a fixed high pressure. It decreases the turbine expansion and also decreases the pump supply. The vapor fraction at the separator increases with increase in the strong solution concentration. These factors lead to increase the net output at a variable phase. Increase in strong solution results higher efficiency only at the high separator temperature i.e. 120 °C. The efficiency decreases with an increase in the strong solution concentration below this temperature (120 °C). On an average basis the specific power output increases from 0.5 to 1.5 kW per 1 °C rise in the separator temperature. Similarly, there is a 0.1% rise in plant efficiency and 0.25% rise in the cycle efficiency per 1 °C rise in the separator temperature. Hettiarachchi et al. [16] reported power output of 60–70 kW per kg/s of strong solution at a heat source
Fig. 2. Variation of plant and cycle performance with strong solution concentration and separator temperature at the turbine inlet concentration of 0.95.
184
N. Shankar Ganesh, T. Srinivas / Applied Energy 91 (2012) 180–186
Fig. 3. Influence of strong solution and turbine inlet concentrations on performance of plant at the separator temperature of 125 °C.
temperature of 90 °C. These are given at a PP of 2 °C. The current model results 50–70 kW per kg/s at 5 °C PP and 110 °C of separator temperature. Pouraghaie et al. [26] resulted the power outputs from 60 to 110 kW per kg/s at the superheater temperature of 125–225 °C. Lolos and Rogdakis [27] gave a turbine output of 216.7 kJ/kg vapor for maximum work condition at the superheated temperature of 130 °C. The current model results 198 kJ/kg vapor at the same superheated temperature. This is lower than the reported result due to the higher sink temperature in the current model compared to the reported sink temperature of 20 °C. They resulted the cycle efficiency from 8.1 to 8.3% and the same for the current model is 8.6%. Fig. 3a and b depicts the effect of strong solution concentration and turbine inlet concentration on the performance of the solar thermal power plant. The efficiencies (plant and cycle) are increasing with an increase in turbine concentration. As per Bombarda et al. [28] results, the optimum strong solution concentration increases with an increase in separator temperature to get a maximum net power output. The current work results the maximum power outputs at 70 kW, 76 kW, 83 kW and 93 kW per kg/s of strong solution respectively at the strong solution concentrations of 0.65, 0.7, 0.75 and 0.8. Turbine inlet concentration influences the high pressure but not the low pressure. Similarly, the strong solution concentration effects the low pressure only. A rise in
turbine concentration increases the high pressure and decreases both the vapor fraction and heat supply in the boiler. The combined influence of these parameters, increases the net power output and then decreases with a rise in turbine inlet concentration. It also causes a rise in plant efficiency. A rise in strong solution concentration increases the low pressure, vapor fraction and heat supply. All these parameters increases the power output with a rise in strong solution concentration but the rise in efficiency takes place above the 0.95 concentration only. The specific power output increases from 1.5 to 3 kW per 1% rise in the strong solution concentration. There is a 0.2% rise in the plant efficiency and 1% rise in the cycle efficiency with a 1% rise in the turbine inlet concentration. Fig. 4 has been plotted for (a) plant performance and (b) cost of solar collectors at various beam radiations and separator temperatures. The solar direct beam radiation is varied from 400 W/m2 to 700 W/m2 as it varies from place to place and also with the time. The separator temperature is varied from 110 °C to 150 °C. The results are plotted based on the fixation of cycle conditions with a suitable collector area as per the available beam radiation. The plant results a maximum power at 135 °C of separator temperature and 0.75 strong solution concentration. The efficiency increases with an increment in sun beam radiation. The solar collector efficiency increases with an increase in beam radiation even through the cycle efficiency is kept constant. The specific power
Fig. 4. Influence of solar beam radiation with separator temperature on plant performance and collectors cost at the strong solution concentration of 0.75.
185
N. Shankar Ganesh, T. Srinivas / Applied Energy 91 (2012) 180–186 Table 1 Solar thermal power plant material flow details with respect to Fig. 1 at Indian climatic conditions and Tsep = 125 °C. State
Pressure (bar)
Temperature (°C)
Ammonia concentration
Flow rate (kg/s)
Specific enthalpy (kJ/kg)
1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16
35.6 7.5 7.5 7.5 7.5 35.6 35.6 35.6 35.6 35.6 35.6 35.6 7.5 – – –
135.0 66.4 57.6 53.4 30.0 31.8 44.6 82.5 125.0 125.0 125.0 49.6 50.1 145.0 143.9 92.5
0.95 0.95 0.70 0.70 0.70 0.70 0.70 0.70 0.70 0.95 0.47 0.47 0.47 – – –
0.48 0.48 1.00 1.00 1.00 1.00 1.00 1.00 1.00 0.48 0.52 0.52 0.52 3.47 3.47 3.47
1545.2 1363.6 651.9 590.1 55.5 45.2 14.6 198.1 943.8 1513.3 342.2 14.3 14.4 501.6 497.2 282.1
Table 2 Specifications of the solar thermal power plant at the hot fluid inlet temperature of 145 °C. Description
Result
Vapor fraction in separator, % Heat load in boiler, kW Heat load in superheater, kW Heat load in low temperature regenerator, kW Heat load in high temperature regenerator, kW Heat load in condenser, kW Work output of mixture turbine, kW Work input to solution pump, kW Net electricity output, kW Kalina cycle energy efficiency, % Solar plant efficiency, % Total length of receiver tubes, m Total area of collectors, m2 Cost of collectors, Rs. in lakhs
48 746 15 62 184 646 86.6 10.6 76 10 4 2722 2042 193
Table 3 Comparison of the present work with the existing plant readings at 75% turbine efficiency (Tsep = 120 °C, x1 = 0.95, T5 = 12 °C, Welectric = 1.82 MW). Sl. no. 1 2 3 4 5 6 7 8 9 10 11 12 12 13 14
Description
Plant result
Predicted result
Hot water requirement, kg/s Hot water inlet temperature, °C Hot water outlet temperature, °C Separator pressure, bar Low pressure, bar Temperature of working fluid at boiler inlet, °C Strong solution concentration Separated liquid concentration Vapor in separator, kg/s Ammonia–water mixture before separation, kg/s Temperature after expansion, °C Temperature after pumping, °C Turbine output, kW Pump input, kW Cycle energy efficiency, %
90 124 80 31 5.5 67
74.2 127 77 30.2 5.3 67
0.81 0.5 11.2 16.3
0.8 0.45 10.6 15.1
60 13 1950 130 11.8
55 14.8 1968 148 11.7
output is constant with the solar radiation as it depends on the cycle conditions in the current study. The variations in collector area, collector cost with separator temperature and direct beam radiation have been plotted in Fig. 4b. The economic conditions are considered with the government subsidy of Rs. 5500/m2 of collection area. At the strong solution of 0.75, the plant efficiency increases with the separator temperature. Therefore the minimum cost can be obtained at higher solar radiation and separator temperature.
The collector cost obtained is matched with the manufacturers cost quotation with the government subsidy. At the predefined conditions and a change in the parameters, 2.5–6% plant efficiency, 65–95 kW of specific electric power, 1500–5000 m2 collector area and Rs. 150 lakhs to Rs. 450 lakhs of collectors cost are resulted with the parametric variations. Table 1 shows the properties of the working fluid and hot fluid at the state points, defined in Fig. 1. The results are plotted at the separator temperature of 125 °C, 0.7 strong solution concentration, 1 kg/s of working fluid and the turbine concentration of 0.95. It results 145° hot fluid inlet temperature and 135 °C separator temperature. A unit mass of working fluid in the power circuit demands 3.47 units of hot fluid at the above said conditions. It also shows 48% of vapor fraction in the separator. The unknown properties at the state points are determined from the equation outlined from Eqs. (1)–(13). Table 2 gives the results of the solar thermal power plant at the operating conditions stated in earlier sections. The rating of heat exchangers, power, efficiency and cost details are developed at the unit mass of the working fluid. These specifications are developed at the same conditions defined for Table 1. The heat load in the high temperature regenerator is nearly three times more than the low temperature regenerator. The present thermodynamic evaluation has been validated by comparing the existing Kalina cycle power plant located in Husavik, Iceland running from hot water (geo thermal resource). Table 3 compares the present simulated results with the plant readings [29]. The plant and current model are solved with the working conditions i.e. at strong solution concentration = 0.81, T15 = 121 °C, x1 = 0.95 and t5 = 12 °C. The considered real power plant is operating without a superheater; therefore in the present simulated model the superheater is bypassed for this study. Most of the calculated results in this work are closely matched with the Husavik power plant conditions. The strong solution concentration is considered at 0.8 in place of the Husavik value of 0.81. The resulted low pressure is 5 bar where as 5.5 bar is the plant reading. The simulated separator liquid concentration is comparatively weak (0.45) over the plant reading (0.5). These results also matched with the literature results reported by Nasruddin et al. [30] who reported 11.4% of energy net efficiency. 4. Conclusions A solar thermal power plant based on Kalina power system has been modeled thermodynamically and analyzed parametrically. The performance of the plant has been studied at the separator temperature of 110–150 °C, strong solution concentration of
186
N. Shankar Ganesh, T. Srinivas / Applied Energy 91 (2012) 180–186
0.65–0.8, turbine inlet concentration of 0.92–0.955 and solar beam radiation of 400–700 W/m2. The optimum separator temperature is different for maximum efficiency condition and for maximum power condition. The maximum power can be obtained relative at lower separator temperature compared to the maximum efficiency conditions at a fixed strong solution concentration. The turbine inlet concentration influences the high pressure whereas the strong solution concentration influences the low pressure. The minimum collector cost can be obtained with high solar beam radiation and optimized cycle conditions. The thermodynamic model has been validated by comparing the literature and the existed plant readings. References [1] Kalina IA. Combined cycle system with novel bottoming cycle. ASME J Eng Gas Turb Power 1984;106:737–42. [2] Sayed YMEI, Tribus M. A theoretical comparison of Rankine and Kalina cycles. ASME publication; 1985. AES: 1. [3] Marston CH. Parametric analysis of the Kalian cycle. ASME J Eng Gas Turb Power 1990;112:107–16. [4] Rogdakis ED, Antonopoulos KA. A high efficiency NH3–H2O absorption power cycle. Heat Recov Syst CHP 1991;11:263–75. [5] Rogdakis ED. Thermodynamic analysis, parametric study and optimum operation of the Kalina cycle. Int J Energy Res 1996;20(4):359–70. [6] Ziegler B, Trepp C. Equation of state for ammonia–water mixtures. Int J Refrig 1984;7(2):101–6. [7] Leibowitz H, Mirolli M. First Kalina combined cycle plant tested successfully. Power Eng 1997;101(5). [8] Tamm G, Goswami DY, Lu S, Hasan AA. Novel combined power and cooling thermodynamic cycle for low temperature heat sources, part 1: theoretical investigation. ASME J Sol Energy Eng 2003;125(2):218–22. [9] Tamm G, Goswami DY. Novel combined power and cooling thermodynamic cycle for low temperature heat sources, part 2: experimental investigation. ASME J Sol Energy Eng 2003;125(2):223–9. [10] Hasan AA, Goswami DY. Exergy analysis of a combined power and refrigeration thermodynamic cycle driven by a solar heat source. ASME J Sol Energy Eng 2003;125(1):55–60. [11] Tamm G, Goswami DY, Lu S, Hasan AA. Theoretical and experimental investigation of an ammonia–water power and refrigeration thermodynamic cycle. Sol Energy 2004;76(1–3):217–28. [12] Zheng D, Chen B, Qi Y, Jin H. Thermodynamic analysis of a novel absorption power/cooling combined-cycle. Appl Energy 2006;83(4):311–23.
[13] Valdimarsson P, Eliasson L. Factors influencing the economics of the Kalina power cycle and situations of superior performance. In: International geothermal conference 2003, Reykjavik; 2003. p. 32–40 Paper No. 79. [14] Desideri U, Bidini G. Study of possible optimization criteria for geothermal power plants. Energy Convers Manage 1997;38(15–17):1681–91. [15] Dejfors C, Svedberg G. Second law analysis of ammonia–water power cycle for direct-fired cogeneration application. Int J Appl Thermodyn 1999;2(3): 125–31. [16] Hettiarachchi HDM, Golubovic M, Worek WM, Ikegami Y. The performance of the Kalina cycle system 11 (KCS-11) with low-temperature heat sources. ASME J Energy Resour Technol 2007;129(3):243–7. [17] Murugan RS, Subbarao PMV. Thermodynamic analysis of Rankine–Kalina combined cycle. Int J Thermodyn 2008;11(3):133–41. [18] Srinivas T, Gupta AVSSKS, Reddy BV. Performance simulation of combined cycle with Kalina bottoming cycle. Cogener Distrib Gener J 2008;23(1):6–20. [19] Galanis N, Cayer E, Roy P, Denis ES, Desilets M. Electricity generation from low temperature sources. J Appl Fluid Mech 2009;2(2):55–67. [20] Hua E, Yang YP, Nishimura A, Yilmaz F, Kouzani A. Solar thermal aided power generation. Appl Energy 2010;87(9):2881–5. [21] Chacartegui R, Sanchez D, Muñoz JM, Sanchez T. Alternative ORC bottoming cycles for combined cycle power plants. Appl Energy 2009;86(10):2162–70. [22] Cayer E, Galanis N, Desilets M, Nesreddine H, Roy P. Analysis of a carbon dioxide transcritical power cycle using a low temperature source. Appl Energy 2009;86(7–8):1055–63. [23] Baik YJ, Kim M, Chang KC, Kim SJ. Power-based performance comparison between carbon dioxide and R125 transcritical cycles for a low-grade heat source. Appl Energy 2011;88(3):892–8. [24] Wang J, Dai Y, Gao L. Parametric analysis and optimization for a combined power and refrigeration cycle. Appl Energy 2008;85(11):1071–85. [25] Valan Arasu A, Sornakumar T. Design, manufacture and testing of fiberglass reinforced parabola trough for parabolic trough solar collectors. Sol Energy 2007;81(10):1273–9. [26] Pouraghaie M, Atashkari K, Besarati SM, Nariman-zadeh N. Thermodynamic performance optimization of a combined power/cooling cycle. Energy Convers Manage 2010;51(1):204–11. [27] Lolos PA, Rogdakis ED. A Kalina power cycle driven by renewable energy sources. Energy 2009;34(4):457–64. [28] Bombarda P, Invernizzi CM, Pietra C. Heat recovery from diesel engines: a thermodynamic comparison between Kalina and ORC cycles. Appl Therm Eng 2010;30(2–3):212–9. [29] Mirolli M, Hjartarson H, Mlcak H, Ralph M. Testing and operating experience of the 2 MW Kalina cycle geothermal power plant in Husavik, Iceland. In: Proceedings of the world renewable energy congress VII, Cologne, Germany; 2002. [30] Nasruddin, Usvika R, Rifaldi M, Noor A. Energy and energy analysis of Kalina cycle system (KCS) 34 with mass fraction ammonia–water mixture variation. J Mech Sci Technol 2009;23(7):1871–6.