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Analysis of combustion turbine inlet air cooling systems applied to an operating cogeneration power plant
Energy Conversion and Management 49 (2008) 2130–2141
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Analysis of combustion turbine inlet air cooling systems applied to an operating cogeneration power plant R. Chacartegui *, F. Jiménez-Espadafor, D. Sánchez, T. Sánchez Thermal Power Group (GMTS), Department of Energy Engineering, University of Seville, Camino de los Descubrimientos s/n, 41092 Sevilla, Spain
a r t i c l e
i n f o
Article history: Received 20 June 2007 Accepted 20 February 2008 Available online 14 April 2008
Keywords: Combustion turbine inlet air cooling Cogeneration Power plant modelling
a b s t r a c t In this work, combustion turbine inlet air cooling (CTIAC) systems are analyzed from an economic outlook, their effects on the global performance parameters and the economic results of the power plant. The study has been carried out on a combined cogeneration system, composed of a General Electric PG 6541 gas turbine and a heat recovery steam generator. The work has been divided into three parts. First, a revision of the present CTIAC technologies is shown, their effects on power plant performance and evaluation of the associated investment and maintenance costs. In a second phase of the work, the cogeneration plant was modelled with the objective of evaluating the power increase and the effects on the generated steam and the thermal oil. The cogeneration power plant model was developed, departing from the recorded operational data of the plant in 2005 and the gas turbine model offered by General Electric, to take into consideration that, in 2000, the gas turbine had been remodelled and the original performance curves should be corrected. The final objective of this model was to express the power plant main variables as a function of the gas turbine intake temperature, pressure and relative humidity. Finally, this model was applied to analyze the economic interest of different intake cooling systems, in different operative ranges and with different cooling capacities. Ó 2008 Elsevier Ltd. All rights reserved.
1. Introduction A combustion turbine inlet air cooling system (CTIAC) is a device, or a group of devices, that cools the gas turbine air intake. A gas turbine’s performance is highly dependent on the intake flow conditions. A gas turbine loses approximately 7% of its nominal power when the intake temperature increases from 15 °C, ISO conditions, to 25 °C, and in cases such as the power plant under consideration, when in summer, the ambient temperature increases above the 25 °C, the losses are still bigger, reaching even 15% of the power rating with 36 °C. Other effects of the higher intake air temperature are the increase of the heat rate (HR) and decrease of the compression ratio, as well as the increase of the gas turbine exhaust temperature and decrease of the exhaust gases mass flow. Therefore, the introduction of CTIAC systems will modify the heat delivery characteristics to the systems installed in the gas turbine exhaust, a steam generator in the case under study, and necessitate analysis of the effect on the performance of the whole power plant [1–3]. The introduction of an intake air cooler in an operating power plant allows producing a higher power than the nominal rating power by cooling the air below ISO conditions, depending on the conditions outside the plant [4]. This intake air cooling has limits; * Corresponding author. Tel.: +34 954 48 72 42; fax: +34 954 48 72 43. E-mail address: [email protected] (R. Chacartegui). 0196-8904/$ - see front matter Ó 2008 Elsevier Ltd. All rights reserved. doi:10.1016/j.enconman.2008.02.023
on the one hand, are the gas turbine mechanical characteristics and alternator power limitation, while on the other, ice crystals may appear in the gas turbine intake, which can generate erosion and wear of the intake vanes. So, in practice, intake temperature does not go below 43–45 °F (6.1–7.2 °C) [1,3,5]. Installation of CTIAC systems in operating plants presents some associated inconveniences, as they require investment in equipment and maintenance costs. In some cases, it even requires remodelling of some gas turbine sections. Besides, the equipments installed at the turbine air intake sections create additional pressure drops that will depend on the CTIAC technology installed. Also, the profitability of introducing a CTIAC system depends on many external factors that are not controllable, like climatology, electricity tariffs and fuel price. So, it requires a particular viability analysis for each power plant. 2. Combustion turbine inlet air cooling systems In summary, the main CTIAC technologies are the following [6–9]: 2.1. Evaporative coolers [10–13] They use water evaporation to cool the air. The water evaporates as it absorbs heat from the incoming air and reduces the
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Nomenclature Cp CH CDS COP CPI CTIAC EA FOG IGV GN HR HRSG Hp IIR m MH NPV OMEL P r RT SH T TES TIT W Wc w
constant pressure specific heat chiller percentage of cooling demand satisfied coefficient of performance consumer price index combustion turbine inlet air cooling absorption chillers fogging cooling system inlet guide vanes natural gas heat rate heat recovery steam generator lower calorific value internal rate of return mass flow (kg/s) evaporative media cooling system net present value electricity market operator price compression ratio refrigeration tons (1 RT equals 12,000 Btu/h or 3.52 kW) hybrid systems temperature thermal energy storage systems turbine inlet temperature power compression work humidity
dry bulb temperature. Evaporative systems can be classified as direct evaporative systems and indirect evaporative systems. In indirect evaporative systems, water is not brought in direct contact with the incoming air, water evaporation and the decrease in dry bulb temperature is done within an intermediate air stream flow, and it cools the compressor intake air in an additional heat exchanger [10]. In direct evaporative systems, water is brought in contact with the incoming air, and it evolves with the compressor intake stream flow. The main direct evaporative cooling systems are the evaporative media (MH) [11] and fogging systems (FOG). The minimum temperature is limited to 10 °C to avoid the risk of freezing at the intake of the compressor. They are capable of cooling inlet air to within 85–90% of saturation. In these evaporative systems, the evolution of fine water drops can produce an increase of extra power [14]. In addition to increasing the incoming mass flow, evaporation of drops will absorb latent heat, reducing the temperature of the mass that passes through the compressor [2], reducing the compression work that depends on the intake flow temperature, Eq. (1) [10] Wc ¼
C p T 01 ðr ðc1Þ=c c
1Þ
ð1Þ
Reduced NOx emissions with this CTIAC are attributed to the increased thermal heat capacity because of the water and the decreased compressor discharge temperature [5]. In evaporative media, the air passes over surfaces where water circulates. MH systems have the lowest installation costs of all the CTIAC systems. Also, they have a very low parasitic load (power consumed by the cooling system and auxiliary systems). Another additional advantage is that MH systems use potable water (cheaper than deionised water). Their main drawbacks are a high dependence on ambient conditions (temperature and relative humidity evolution), and the intake pressure drop even takes place when the cooling system is not working.
Subscripts 01 stagnation inlet a air amb ambient temperature c compressor cc combustion chamber cs single effect absorption cycle cd double effect absorption cycle disp available elec electric motor driven chiller f fuel g gas turbine exhaust h high pressure steam i initial conditions ice ice storage system mg gas engine driven chiller smed medium pressure steam t turbine tv steam turbine driven chiller w cold water storage system Greek symbols c specific heat ratio g efficiency
In fogging systems, high pressure water is injected [12,13]. These systems can produce nearly saturated air due to the very small drop size, even with reduced flow residence times, and so, they have a very high effectiveness, about 0.95–1. The grill of injectors also does not disturb the flow, and the pressures drops are very low. These systems have the possibility of overspray operation, or high fogging, with higher mass flow than would be needed to saturate the intake air, to force an ‘‘intercooling” of the air passing through the compressor [15]. Fogging systems have the second lowest acquisition costs of all CTIAC systems. Also, they have a low parasitic load. As drawbacks, these systems have a limited cooling capacity, down to the wet bulb temperature, and a high ambient dependence, they consume deionised water, are more expensive and present corrosion problems. Estimated installation costs and parasitic loads for evaporative systems are shown in Table 1. 2.2. Mechanical cooling compression or chillers (CH) They use a classical mechanical compression cycle, with an intermediate refrigerant fluid, which cools the intake air passing through heat exchangers located in the gas turbine inlet [7,16]. These systems can produce a stronger cooling effect than the evaporative systems, cooling the air down to 45 °F (7.2 °C). Estimated installation costs and mean parasitic load for these systems are compared in Table 2, differing by the form of driving the compressor among chillers with electric motor drive (CHelec), chillers with Table 1 Estimated investment costs and parasitic loads for evaporative systems
Investment cost (€/kW added) Parasitic load (% extra power generated)
MH
FOG
25–60 0.3–0.5
30–70 0.5–0.7
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Table 2 Estimated investment costs, parasitic loads and COP for mechanical and absorption chillers
Investment cost (€/RT) Parasitic load (kW/RT) COP
CHelec
CHmg
CHtv
EAcs
EAcd
870 0.81 2 and 6
1.225 0.187 1 and 2.5
1.300 0.195 1 and 2.5
1.000 0.31 0.7
1.130 0.265 1.2
gas engine drive (CHmg) and chillers with steam turbine drive (CHtv). Values are given as a function of the refrigeration capacity installed in RT (refrigeration tons). 2.3. Absorption chillers (EA) They use an absorption refrigeration cycle with an absorbent fluid and a refrigerating fluid. The working fluids used are water/ ammonia [17] or lithium bromide/water [3]. Although the ammonia presents better refrigerating properties than water, for toxicity problems, the technology of lithium bromide/water [6,18–20] is used. The evaporation temperature limits the inlet air cooling to 10 °C, higher than the temperature reached with mechanical chillers. Estimated mean installation costs and parasitic loads for the single effect LiBr–H2O absorption refrigeration cycle (EAcs) and the double effect LiBr–H2O absorption refrigeration cycle (EAcd) are shown in Table 2. The double effect LiBr–H2O absorption refrigeration cycle improves the performance of the single effect cycle. Also, different heat sources have been considered, direct heat source in an auxiliary boiler or taking advantage of the residual heat of the plant, for the single effect cycle with hot water at 115 °C and for the double effect cycle, medium pressure steam (8.5–10 bar). 2.4. Hybrid systems They are combinations of two or more of the previously mentioned systems. Their main purpose is to give operating flexibility to the cooling system to cover the demand, avoiding high parasitic load (electricity consumption) in periods with high electricity tariffs. The most usual combinations for hybrid systems are mechanical chillers (mainly electric motor driven chillers) in combination with absorption chillers. The use of evaporative coolers combined with mechanical chillers is not effective. 2.5. Thermal energy storage systems A thermal energy storage system (TES) allows storing cold (heat) so that a delay exists between the cold fluid storage (charging) and its consumption (discharging). Their feature is double, the main one is to avoid parasitic load consumption in ‘‘on-peak” periods, and the second one is to be able to cover cooling demand peaks with smaller cooling systems [13]. The use of continuous cooling is of interest when cooling the intake is needed for any one period greater than six hours [6]. They are usually installed with electric chillers, although they can also be used with hybrid systems. According to their storage form, these systems will be classified as cooled water storage systems (TESw), ice storage systems (TESice) and eutectic salt mixtures storage systems [21–24].
Cooled water storage systems store sensible heat energy; they are easy to install and have quick response on demand. They are more profitable for large storage sizes. Ice storage systems store energy in form of latent heat energy for which they are as much as six times smaller than a cooled water storage system of the same heat storage capacity. They work at a lower temperature of evaporation, and this penalizes the COP of the plant. A wide variety of types exists, but one of quicker response is the Ice Harvester. The estimated costs for each storage system are given in Table 3. 3. Power plant description The power plant under study is located in the south of Spain, in Cádiz, next to the sea. The cogeneration plant is inside a chemical plant, and it is composed of a General Electric gas turbine, model PG6541B, installed in 1995, and a heat recovery boiler. The turbine works in an open cycle and is composed of an axial compressor with 17 stages and a compression ratio of 1:11, and an axial turbine with 3 stages with a turbine inlet temperature (TIT) of 1100 °C. This gas turbine can be fuelled with natural gas or with naphtha. Their nominal characteristics values are Nominal Power ðGNÞ ¼ 38:340 kW Nominal Heat Rate ðGNÞ ¼ 11:460 kJ=kW h The gas turbine is coupled to a natural circulation heat recovery steam generator (HRSG), without post combustion, which takes advantage of the residual heat of the exhaust gases. In the HRSG, superheated steam is generated at two pressures (10 and 64 bars) and thermal oil is heated for its use in the chemical processes of the plant (290–345 °C). The HRSG generated mass flows and stream flow temperatures are given in Table 4. Of the electric power generated in the gas turbine, part is consumed in the chemical plant, and the remainder is exported to the electric network. Also, from the generated steam, part is used in their own chemical plant, and the remainder is exported to an adjacent refinery, while all the thermal oil is consumed in the chemical plant. The thermal oil mass flow and temperature are imposed by the demands of the chemical external process. In this part of the work, a model of the power plant was developed to be used to evaluate the economic effect of inserting different CTIAC systems for this power plant [7,3,25]. In 2000, were remodelled the gas turbine, changing the vanes of the second and third stages and the IGV (inlet guide vanes). As a consequence, neither the original curves of the gas turbine, GE PG6541B, nor those corresponding to a similar model after the remodelling, GE PG6551, were a valid solution for direct application in the power plant model [24]. After the modification, GE provided two correlations for the power and the heat rate, but they did not include the other necessary variables for study of the cogeneration power plant performance, like the exhaust flow gases and the exhaust temperature. On the other hand, for the HRSG, isolated and insufficient information was provided for its model. So, a model was generated through 2005 registered operational data. It was generated using a filtrate of the data, discarding anomalous or not valid data for the model that was desired. The data corresponding to operative
Table 4 HRSG generated mass flows and stream flow temperatures Table 3 Estimated investment costs for thermal storage systems
€/RT h
TESw
TESice
157.5–2.75 10-3 RT h for 2000 < RT h < 20,000
145 €/RT h
Steam 64 bar Steam 10 bar Thermal Oil
Production (T/h)
Temperature (°C)
40.7 13.7 395
480 187 290 and 345
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conditions in which the values of power delivery by the gas turbine that indicated start conditions or stopping conditions, that show operating with washing on line or that with strongly partial operating conditions were discarded. Also, the data of generation of steam that indicated a heating or stopping process in the HRSG or a stack diverter by pass were discarded. After this filtration, about 85% of the 2005 time registers were valid data to prepare the model.
The model of the gas turbine was prepared for base load operation, about 95% of the annual performance of the plant. The used expressions to model power, heat rate and exhaust temperature in the plant as functions of the intake temperature, relative humidity, nature of fuel and pressure drops are given in Appendix 1. Figs. 1–4 represent the values for these three parameters obtained from the operation registers as well as those obtained from the correlations given by the manufacturer.
Fig. 1. Registered power data as function of intake temperature.
44 Original manufacturer correlation
42
Correlation from measured power
Power (MW)
40 Manufacturer correlation after gas turbine modification
38 36 34 32 30 0
5
10
15
20
25
30
35
40
Intake Temperature (ºC) Fig. 2. Power calculated correlation from measured data and GE power correlations before and after gas turbine modifications.
Fig. 3. Heat rate obtained from 2005 registered data as function of intake temperature.
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13000
Heat Rate (kJ/kWh)
12500
12000
11500 Original Heat Rate correlation Heat Rate from measured data Heat Rate correlation after gas turbine modification
11000
10500 0
10
20
30
40
Intake Temperature (ºC) Fig. 4. Heat rate calculated correlation from measured data and GE HR correlations before and after gas turbine modifications.
With the measured values, or those estimated from them, the expressions for the model of the gas turbine were adjusted, having the gas turbine inlet temperature, relative humidity, fuel and pressure drops, as independent variables and obtaining the power and heat rate as a result of the inlet conditions from the expressions. The dispersion of data can be due to differences in the composition of the fuel from the one expressed in the correlations, pollution of filters, uncertainty in measurements, imperfect fit of the correlations given by the manufacturer, or safety regulation controls that continuously vary the gas turbine performance around a fixed operating point (gas generator speed, TIT, exhaust gas temperature etc.). For the joint model of the gas turbine–HRSG, the registered exhaust temperature was used, as is shown in Fig. 5, developing a correlation with a 95% confidence interval. Thus, obtaining the correlation given in Eq. (2) where the last term shows the confidence interval between both signs T g ð CÞ ¼ 0:722T amb þ 537:49 5:72
ð2Þ
The exhaust gases flow was correlated with the intake conditions through the global energy balance in the gas turbine, Eq. (3). The correlation obtained is shown in Eq. (5). The original correlation for the exhaust gases flow and the new correlation obtained are shown in Fig. 6
Fig. 5. Exhaust gases temperature correlation obtained from measured data.
ma C pa T amb þ mf Hp gcc ¼
W þ ðma þ mf ÞC pg T g gt
ð3Þ
where the efficiency of the turbine was correlated by the expression
Fig. 6. Exhaust gases mass flow, original correlation and new correlation obtained from measured data.
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gt ¼ 0:0008W þ 0:9203 mg ¼ 0:4489T amb þ 138:29 1:48
ð4Þ ð5Þ
W disp ðkWÞ ¼ 179:4 T amb þ 87080
ð6Þ
45
Steam mass flow (Tn/h)
With these expressions for the exhaust temperature and exhaust mass flow, the model of the HRSG and the power plant with base load operation was constructed with the independent variables being the mass flow and exhaust temperature, related directly with the gas turbine intake conditions [26]. As results, the model gives the steam production and the thermal oil production as a function of a unique variable that includes, simultaneously, the mass flow and the exhaust temperature in the form of available power in the exhaust gases, Fig. 7. The obtained correlation for this variable is shown in Eq. (6). Starting from the steam and thermal oil productions expressed as a function of available power, Fig. 7, as well as the relationship among the production of each one of the flows, Fig. 8, conditioned by the HRSG design, Fig. 9, and the heat demand of the chemical plant where the power plant is placed, the expressions shown in Appendix 1 were obtained. With this model, the economic study of the different CTIAC systems, as functions of the prices of fuel and electricity, as well as the costs of the equipments (investment, maintenance, others) can be conducted
50
40 35 30 25 20 15 10 High pressure steam Medium pressure steam
5 0 10000
11000
12000
13000
14000
The relationship among the power supplied with the thermal oil flow and the steam flows generated is shown in Fig. 10 where, due to the thermal oil temperature ranges, the thermal oil economizer is situated between the medium pressure steam generator and the high pressure steam generator. The HRSG control system adjusts
Available power inexhaust gases (kW)
87000 86000 85000 84000 83000 82000 81000 80000 5
10
15
20
25
30
35
40
IntakeTemperature (ºC) Fig. 7. Available exhaust gases heat power as function of intake temperature.
70000
HRSG Heat Balance (kW)
60000 Total Heat High pressure steam Thermal oil Medium pressure steam
50000 40000 30000 20000 10000 0 80000
81000
82000
83000
16000
Fig. 9. Relationship among steam mass flows and thermal oil heat demand.
88000
0
15000
Thermal oil power consumption (kW)
84000
85000
86000
Exhaust gases available heat power (kW) Fig. 8. HRSG heat balance.
87000
88000
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Fig. 10. HRSG scheme. Stack gases temperature (°C), high pressure steam flow and medium pressure steam flow evaporators and thermal oil economizer.
the thermal oil temperature, and when the thermal oil demand decreases, the high pressure sections absorb more heat and the disposal heat to the medium pressure sections of the steam generator decrease. 4. CTIAC systems modelling In this section, the treatment given to the different cooling systems, as well as the main hypotheses used, is discussed. Depending on the inlet air conditions, the cooling evolution in the chillers will be different. Different cooling possibilities are shown in the psychometric chart, Fig. 11. Three zones can be distinguished in this psychometric chart, delimited by two lines, one horizontal and one vertical, one line of dry bulb constant temperature and another line of constant relative humidity. The first is determined by the cooling temperature limit of the chiller, 7.2 °C in the case of mechanical chillers and 10 °C with absorption chillers, while the second line corresponds to constant absolute humidity with the value of absolute humidity given by the intersection point with the limit dry bulb temperature. Zone 1 corresponds to all the points with dry bulb temperature less than the limit temperature; therefore, the final conditions of any point with these initial conditions within this zone are the same as the initial conditions of temperature for any humidity. Zones 2 and 3 are those in which the temperature of the air is above the limit temperature, and therefore, they can be cooled by the chiller. The difference resides in the absolute humidity of the points of the zones, the first has a lower humidity than the suitable one and the second ha a greater humidity. The points within zone 2 evolve according to an absolute constant humidity line; they consume sensible heat but not latent heat, since they can never arrive at the saturation. Their limit is exactly the consigned temperature limit, although if the cooling capacity of the chiller is not enough, the final temperature will be greater than the limit temperature of the chiller. For the points belonging to zone 3, two different evolution forms exist, one for the initial conditions marked in the figure as type 1, in which the cooling capacity of the chiller is not able to cool the air sufficiently to take it to the sat-
uration line, and the other for the initial conditions marked in the figure as type 2, in which the cooling capacity of the chiller is able to take the initial point to the saturation and to continue cooling it, following the saturation line. Given the initial conditions (Ti, Wi) and the cooling capacity of the selected chiller, it is possible to calculate the maximum cooling that could be provided for any initial point and the final temperature that could be reached, depending upon which of the three ways of cooling is chosen. The intake pressure drop penalties in power and heat rate have been considered, and in the case of inserting a heat exchanger in the exhaust, as in the case of absorption chillers, the additional pressure drops have been included and the power and heat rate corrected using the expressions shown in Appendix 1. In the case of absorption chillers, the maximum heat that can be recovered from the stack is conditioned by the stack exhaust temperature that can be reached without corrosion problems due to condensation on the tubes. In the case of a Li–Br single effect absorption chiller, the maximum cooling capacity that can be used, using hot water, will be 300 RT, while for a Li–Br double effect absorption chiller, using steam, it can reach a maximum capacity of 500 RT. Hourly electricity tariffs have been considered. In Fig. 12, the hourly prices evolution of representative days taken for January and July are shown. In the viability analysis of hybrid systems, and later of the TES systems, this treatment of the electricity prices throughout the day is fundamental to the analysis of the chiller that will operate in determined periods to minimize the operating costs of the group, obtaining the maximum benefits of the power plant. For thermal energy storage systems, a daily storage system was chosen, developing a typical day model for each month, with the associated hourly electricity tariffs, taking a series of representative days such that the amount of cooling reached 80% of the cooling demand of the month. In Fig. 13, the mean chiller cooling capacity and minimum storage capacity obtained for the year using these representative days to cover the whole cooling demand are shown.
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Fig. 11. Chiller cooling treatment as function of the initial conditions and the chiller cooling capacity.
From Fig. 13, two main conclusions can be extracted. The first is obvious; the required mean cooling power is increased in summer, and the second is that the associated cooling storage minimum capacity does not follow the demand shape exactly, reaching its maximum values in months in which the relative time delay between the cooling demand and cooling supply is higher.
5. Economic considerations of the analysis The main aspects and hypotheses of the economic analysis are shown in this section. The phases of the economic analysis are summarized in the following steps: Fig. 12. Hourly electricity prices.
– Election of a cooling system with a refrigeration capacity. – Evaluation of the benefits of the group before and after implantation of the cooling system. – Differential evaluation of the costs and benefits of operation (electricity, steam, thermal oil, natural gas, deionised water). – Evaluation of the maintenance costs and investment. – Economic analysis of the solution. – Return to the first point, varying the capacity of the equipment.
Fig. 13. Evaluated monthly chiller mean cooling power and related minimum storage capacity.
Prices of steam, thermal oil and natural gas have been taken from 2005 data. Also, the electricity tariffs have been taken from the OMEL. The annual evaluation of costs and benefits has been divided into four periods, corresponding to the quarterly revisions of the rates of natural gas. A power plant operation factor of 0.94 has been considered over the total annual time. It affects the benefits, the operating costs and the maintenance costs. As for other economic aspects: the capital cost has been considered as 7.77%. The refund period and the
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considered redemption period have been taken as 10 years. Although this value is high for evaporative media, it is adjusted for the other cooling equipments, and it has been taken with the same time horizon for all them. A null salvage value and a yearly foreseen CPI of 2.9% have been considered . The tax rate on the economic activity is 33%; it affects the difference between the raw benefit and the depreciation of the installation. While the costs associated with the fuel are the most important operating cost, there are other operating costs that have been kept in mind also; among them are the natural water cost, deionised water cost and the cost of parasitic electricity consumption in the cooling equipment. The cost of natural water only has to be considered in wet media evaporative systems because, for the other systems, this consumption is null. For deionised water, necessary for the additional production of steam in the HRSG and for operation of the fogging system, its price is not only considerable, but in addition, it is necessary to keep in mind the maximum capacity of deionised water production in the plant because having to install new process lines for deionised water excessively diminishes the productiveness of the fogging system. In this case, the company has enough capacity to cover the demand of the fogging systems considered in this work with the increase in steam production, so the price of deionised water is taken as about 1.52 €/ Tm. In this plant, the increase of steam production can become an important part of the benefits (it reaches a value of 20% of the realised profit for electricity). The cost of using steam will take the same value as its selling cost. The price of the additional production of steam of high and medium pressure is given by Eqs. (7) and (8) 1:108P f ð664 15Þ þA 0:86 0:9 1:108Pf ð670 15Þ þA ð€=TmÞ ¼ 0:86 0:9
Psmed ð€=TmÞ ¼
ð7Þ
Psh
ð8Þ
where Pf is the price of fuel (€/kW h), A is the price of demineralised water (€/Tm).
6. Results As a result of this work, an extensive economic analysis was obtained for the cogeneration plant considered, showing the effects of the different CTIAC systems in the power plant performance. In Figs. 14–17, the results obtained for the best configuration studied for each CTIAC system are shown. In Fig. 14, the investments for the best configurations of each CTIAC system are compared; in Fig. 15, the annual benefits are shown for these configurations, and the internal rates of return (IIR) are shown in Fig. 16. The net present values (NPV) of the configurations are shown in Fig. 17. The analysis of evaporative cooling systems showed that, as a certain number of stages in the fogging system is surpassed, the raw annual benefit begins to diminish as well as the NPV, and then, the increase of stages is not translated into an increase of benefits. In these cases, the cooling system would be unnecessarily large. The most interesting evaporation system for this power plant is the fogging system with two stages. The analysis of mechanically driven chillers showed that steam turbine driven chillers have a very low productiveness. This is due to the appraised price of the steam in this installation being high, being more interesting in its direct sale than its use in a chiller to obtain extra additional power. With electrically driven chillers or gas engine driven chillers, we found the best results with refrigeration capacities around 750 RT for the electrically driven chiller and 700 RT for the gas engine driven chiller. With higher refrigeration capacities, although the percentage of cooling demand covered was increased, the annual incomes were not increased in the same manner, which indicates that most of the time, the CTIAC system was over designed. There is a better solution with intermediate refrigeration capacities due to two opposite effects; on one hand, as the cooling capacity of the chiller increases, a great part of this power is used in cooling saturated air, which gives a low ratio between sensible cooling and power consumption, and on the other hand, is the fact that not reaching the minimum temperature limit is a loss of additional electricity production and additional steam production.
Fig. 14. Equipment investment for the best economic configuration of every CTIAC system.
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Fig. 15. Annual benefits for the best economic configuration of every CTIAC system.
Fig. 16. Internal rate of return for the best economic configuration of every CTIAC system.
Fig. 17. Net present value for the best economic configuration of every CTIAC system.
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Table 5 Economic results obtained for the best configuration of each CTIAC system System Evaporative media Fogging of two stages Electric Chiller 850 RT Gas chiller 700 RT Steam chiller 250 RT Absorption chiller simple effect 300 RT Absorption chiller double effect 250 RT Absorption chiller simple effect and Electric chiller 850 RT Electric chiller – gas chiller 575–275 RT Electric chiller – gas chiller 250–500 RT Electric chiller – gas chiller 425–425 RT Electric chiller 600 RT – TES 500 RT h
Investment (€)
Incomes (€/year)
Expenses (€//year)
Benefits (€/year)
Payback (years)
IIR (%)
NPV (€)
CDS (%)
74,000 30,666 850,425 943,250 357,500 330,000 352,000 906,683
192,416 139,764 904,645 996,634 443,490 438,853 420,562 933,403
151,318 109,873 596,317 704,838 344,633 267,406 322,764 603,838
41,098 29,891 308,328 291,796 98,857 171,447 97,798 329,565
3 1.5 4.5 5 6 3 5.5 4.5
40.5 69 21.22 16.42 13.32 34.2 13.45 21.30
146,642 124,311 875,020 722,526 215,845 595,532 214,809 937,055
100 100 73.5 64.1 25.8 37.8 32.2 73.5
974,225 951,591 1,027,837 680,050
984,701 994,084 1,027,830 768,467
660,909 677,191 697,031 482,819
323,792 316,893 330,779 285,648
4.5 4.5 5 3.5
18.53 18.59 17.57 30.64
857,006 840,186 850,753 892,887
73.5 67.4 73.5 73.5
With absorption chillers, the double effect absorption chiller (direct fired) was shown not to be appropriate for this application, and it cannot compete with the other chillers of similar capacities. On the other hand, the single effect absorption chiller is a very interesting solution, although it is desirable that its cooling capacity would be bigger, being limited in its maximum capacity by the available heat in the stack exhaust compatible with the steam generation and thermal oil demand. This is due to its low parasitic load. For hybrid systems, two different combinations were checked, the first one was composed of an electric motor driven chiller and a single effect absorption chiller, which works with hot water produced with the HRSG stack exhaust gases. Given the high productiveness of the single effect absorption chiller, due mainly to its very low operating costs, in this hybrid system setup, the single effect absorption chiller will always work and the electric chiller will act when the demand requires it. The other combination for hybrid systems checked was composed of an electric motor driven chiller and a gas engine driven chiller, the operation modes of both are similar and simply vary the form of work. The main philosophy of this hybrid group was to use the gas engine driven chiller whenever the price of electricity was high to diminish the parasitic load electricity consumption and to maximize the electricity sale. From the results obtained with this configuration, the electric motor driven chiller alone gives better economic performance, and it is not worthwhile installing a hybrid system with this chillers combination. This is due fundamentally to the investment needed with the gas engine driven chiller, higher than that of an electric motor driven chiller, and the operating costs for natural gas, so the savings obtained using gas engine driven chillers, because of its very low parasitic load, is counteracted by the other two factors. Finally, the results of CTIAC systems composed of thermal energy storage combined with electric driven chillers were analyzed. The results obtained showed that the configurations with lower storage capacities were better than those with greater storage capacities due to the fact that in the summery periods, the TES system does not provide an advantage because the demand is higher than the chiller cooling capacity, while in some months, the demand is always below the cooling capacity of the chiller and, although the consumption is the same, the redistribution of parasitic load at lower electricity tariff hours does not compensate for the investment increase. Finally, in those periods in which an oscillation of demand occurs above and below the chiller cooling capacity, it was shown that with a 500 RT h storage capacity, it is enough to cover acceptably the cooling demand peaks. In fact, when the demand is satisfied with the TES system, to raise the storage size affects only the hourly redistribution of the parasitic load, diminishing it at the hours with higher electricity tariff.
In Table 5, the economic results obtained for the best configuration of each CTIAC system are shown. The values presented are investment, incomes, expenses, benefits, payback, IIR, NPV and percentage of cooling demand satisfied (CDS) by the cooling system installed.
7. Conclusions From this work the following conclusions are derived. The main features and costs of the different CTIAC systems have been identified, and this has allowed identifying the most feasible types of CTIAC systems to be introduced in this cogeneration power plant, ranging from very profitable systems with low initial investment, as the evaporative systems are, to others with higher investments but more profitable at the end of their capital refund period, as the hybrid systems with thermal storage are. With the model of the plant implemented, as developed from registered performance data, it has been possible to eliminate some uncertainty in the previous information concerning the power plant equipment. This model has allowed obtaining the global variables of interest for the economic model as functions of the gas turbine intake conditions. The statistical treatment of the operation registers of the power plant obtained in 2005, with the corresponding filtration of operative conditions that were outside the study conditions, has generated a precise model of the power plant for easy coupling of the gas turbine intake with the exit from different cooling systems. From the economic analysis, it is shown that installing a CTIAC system can be a very profitable solution, which also improves the gas turbine efficiency, saves fuel and reduces the CO2 and NOx emissions. From the data shown in Table 5, it is seen that the three most interesting intake cooling systems for this power plant are an electrically driven chiller of 850 RT cooling capacity, a hybrid system with an absorption chiller of 300 RT cooling capacity disposed sequentially with an electrically driven chiller of 450 RT cooling capacity and an electrically driven chiller of 600 RT cooling capacity combined with an Ice Harvester thermal energy storage system with 500 RT h of storage capacity. Of the comparison between the electric motor driven chiller and the electric motor driven chiller with TES system, it is concluded that both systems present very similar results with the exception that the TES system has a bigger NPV with a lower investment, and so, the system with storage would be a more interesting option. Taking the criteria of NPV as the main economic reference, the CTIAC system chosen would be the hybrid system with a value of the NPV 45.000 € greater than the NPV of the TES system, but with a worse IIR and payback period. However, other factors should also
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be contemplated, the main ones of them being the simplicity and installation time. Under these considerations, the system with thermal storage is clearly favourable relative to the hybrid system. The joint chiller plus TES is not a compact group, but it is simpler than the hybrid system with two chillers and will present a reasonable volume for an industrial plant; for a tank of 500 RT h, its volume would be about 35.5 m3. The installation time of the hybrid cooling system is bigger, not only because it is necessary to install two chillers and their interconnections and controls, but also because one of them is an absorption chiller working with hot water generated in the final section of the HRSG. All the necessary actions in the HRSG clearly exceed the installation time of the TES system. As we are studying the enhancement of an existing operating plant, the available time for the cooling system installation will be scheduled to coincide with the power plant planned overhauls. With the hybrid cooling system, the risk of not finishing the installation works in time is higher. The cost of one day with the chemical plant stopped is about 50,000 €, greater than the difference of final benefit between the two options. Also, the control of the plant is easier for the TES system than for the hybrid system, since the control becomes independent of the operation in the HRSG. Under all these considerations, the option of the electrically driven chiller with ice storage would be the CTIAC system chosen for this power plant. Appendix 1. Power plant model equations Gas turbine Power W W W ðkWÞ ¼ ð42487:878 236:51T a 0:256T 2a ÞF W wabs F Hf F Pc
ðA:1Þ
Power corrections FW wabs ¼ 1:00096 0:1519wabs FW Hf FW pc
ðA:2Þ
100 þ ð10:254 3:466 log Hf þ 0:0837 log H2f Þ ¼ 100 0:015D inch H2 O ¼1 4
ðA:3Þ ðA:4Þ
Heat rate HR ðkJ=kW hÞ ¼ ð2731:796 þ 2:859T a HR HR þ 0:064T 2a Þ4:18F HR wabs F Hf F Pc
ðA:5Þ
Heat rate corrections F HR wabs ¼ 0:9977 þ 0:35443wabs 100 þ ð4:50628 þ 1:31713 log Hf þ 100 0:005D inch H2 O ¼1þ 4
F HR Hf ¼ F HR pc
ðA:6Þ 0:01431 log H2f Þ
ðA:7Þ ðA:8Þ
Exhaust gas temperature T g ð CÞ ¼ 0:722T a þ 537:49 5:72
ðA:9Þ
Exhaust mass flow mg ðkg=sÞ ¼ 0:4489T a þ 138:29 1:48
ðA:10Þ
HRSG Exhaust gas available power W disp ðkWÞ ¼ 179:4T a þ 87080
ðA:11Þ
Total heat delivery to cogeneration W cons:tot ðkWÞ ¼ 0:5993W disp þ 6268:5
ðA:12Þ
Heat delivery to medium pressure steam W vapmed ðkWÞ ¼ 0:2339W disp 11503
ðA:13Þ
Heat delivery to high pressure steam W vaphigh ðkWÞ ¼ 0:3654W disp þ 17771:5 W th
oil
ðA:14Þ
References [1] Giampaolo T. The gas turbine handbook: principles and practices. Georgia: Fairmont Press; 2003. [2] Brooks FJ. GE gas turbine performance characteristics. GE Power Systems GER3567H. [3] Wang FJ, Chiou JS. Integration of steam injection and inlet air cooling for a gas turbine generation system. Energy Convers Manage 2004;45:15–26. [4] Kim TS, Ro ST. Power augmentation of combined cycle power plants using cold energy of liquefied natural gas. Energy 2000;25:841–56. [5] Jonhke T, Mast M. Power boosters. Siemens Power J 2002(May). [6] Sanjeev Jolly PE et al. Evaluation of combustion turbine inlet air cooling systems. In: Power Gen Asia, New Delhi, India – September 29–October 1, 1998. [7] Sanjeev Jolly PE et al. Inlet air cooling for a Frame 7EA based combined cycle power plant. In: Power Gen International, Dallas, Texas, December 1997. [8] Al-Amiri Abdalla M, Zamzam Montaser M. Systematic assessment of combustion turbine inlet air-cooling techniques. J Eng Gas Turbines Power 2005;127(January):159–69. [9] Dawoud B, Zurigat YH, Bortmany J. Thermodynamic assessment of power requirements and impact of different gas-turbine inlet air cooling techniques at two different locations in Oman. Appl Thermal Eng 2005;25:1579–98. [10] Bassily AM. Performance improvements of the intercooled reheat regenerative gas turbine cycles using indirect evaporative cooling of the inlet air and evaporative cooling of the compressor discharge. Proc Inst Mech Engrs, Part A: J Power Energy 2001;215(5):545–57. [11] Hosseini R, Beshkani A, Soltani M. Performance improvement of gas turbines of Fars (Iran) combined cycle power plant by intake air cooling using a media evaporative cooler. Energy Convers Manage 2007;48:1055–64. [12] Chaker M, Meher-Homji CB, Mee III T. Inlet fogging of gas turbine engines. Part I fog droplet thermodynamics, heat transfer, and practical considerations. J Gas Turbines Power 2004;126(3). [13] Sanjeev Jolly PE et al. Direct spray system for inlet air cooling W 501 B5. In: Power Gen International, Orlando, Florida, December 1998. [14] Utamura M, Takehara I, Karasawa H. MAT, a novel, open cycle gas turbine for power augmentation. Energy Convers Manage 1998;39(16–18):1631–42. [15] Cataldi Giovanni, Güntner Harald, Matz Charles, McKay Tom, Hoffmann Jürgen, Nemet Anton, et al. Influence of high fogging systems on gas turbine engine operation and performance. J Eng Gas Turbines Power 2006;128(1):135–43. [16] Alhazmy MM, Najjar YSH. Augmentation of gas turbine performance using air coolers. Appl Thermal Eng 2004;24:415–29. [17] Najjar Yousef SH. Enhancement of performance of gas turbine engines by inlet air cooling and cogeneration system. Appl Thermal Eng 1996;16(2):163–73. [18] Bassily AM. Performance improvements of the intercooled reheat recuperated gas-turbine cycle using absorption inlet-cooling and evaporative after-cooling. Appl Energy 2004;77:249–72. [19] Kakaras E, Doukelis A, Karellas S. Compressor intake-air cooling in gas turbine plants. Energy 2004;29:2347–58. [20] Ameri M, Hejazi SH. The study of capacity enhancement of the Chabahar gas turbine installation using an absorption chiller. Appl Thermal Eng 2004;24:59–68. [21] Sanjeev Jolly PE, et al. Capacity enhancement of ABB I INI with thermal energy storage. In: Power Gen International, New Orleans, Louisiana, November 1999. [22] Yokoyama Ryohei, Ito Koichi. Effect of inlet air cooling by ice storage on unit sizing of a gas turbine cogeneration plant. J Eng Gas Turbines Power 2004;126(2):351–7. [23] Hasnain SM, Alawaji SH, Al-Ibrahim AM, Smiai MS. Prospects of cool thermal storage utilization in Saudi Arabia. Energy Convers Manage 2000;41:1829–39. [24] Ameri Mohammad, Hejazi Seyed Hossein, Montaser Kourosh. Performance and economic of the thermal energy storage systems to enhance the peaking capacity of the gas turbines. Appl Thermal Eng 2005;25:241–51. [25] Gareta Raquel, Romeo Luis M, Gil Antonia. Methodology for the economic evaluation of gas turbine air cooling systems in combined cycle applications. Energy 2004;29:1805–18. [26] Horlock JH. Combined power plants. Oxford: Pergamon Press; 1992.
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Energy Conversion and Management journal homepage: www.elsevier.com/locate/enconman
Analysis of combustion turbine inlet air cooling systems applied to an operating cogeneration power plant R. Chacartegui *, F. Jiménez-Espadafor, D. Sánchez, T. Sánchez Thermal Power Group (GMTS), Department of Energy Engineering, University of Seville, Camino de los Descubrimientos s/n, 41092 Sevilla, Spain
a r t i c l e
i n f o
Article history: Received 20 June 2007 Accepted 20 February 2008 Available online 14 April 2008
Keywords: Combustion turbine inlet air cooling Cogeneration Power plant modelling
a b s t r a c t In this work, combustion turbine inlet air cooling (CTIAC) systems are analyzed from an economic outlook, their effects on the global performance parameters and the economic results of the power plant. The study has been carried out on a combined cogeneration system, composed of a General Electric PG 6541 gas turbine and a heat recovery steam generator. The work has been divided into three parts. First, a revision of the present CTIAC technologies is shown, their effects on power plant performance and evaluation of the associated investment and maintenance costs. In a second phase of the work, the cogeneration plant was modelled with the objective of evaluating the power increase and the effects on the generated steam and the thermal oil. The cogeneration power plant model was developed, departing from the recorded operational data of the plant in 2005 and the gas turbine model offered by General Electric, to take into consideration that, in 2000, the gas turbine had been remodelled and the original performance curves should be corrected. The final objective of this model was to express the power plant main variables as a function of the gas turbine intake temperature, pressure and relative humidity. Finally, this model was applied to analyze the economic interest of different intake cooling systems, in different operative ranges and with different cooling capacities. Ó 2008 Elsevier Ltd. All rights reserved.
1. Introduction A combustion turbine inlet air cooling system (CTIAC) is a device, or a group of devices, that cools the gas turbine air intake. A gas turbine’s performance is highly dependent on the intake flow conditions. A gas turbine loses approximately 7% of its nominal power when the intake temperature increases from 15 °C, ISO conditions, to 25 °C, and in cases such as the power plant under consideration, when in summer, the ambient temperature increases above the 25 °C, the losses are still bigger, reaching even 15% of the power rating with 36 °C. Other effects of the higher intake air temperature are the increase of the heat rate (HR) and decrease of the compression ratio, as well as the increase of the gas turbine exhaust temperature and decrease of the exhaust gases mass flow. Therefore, the introduction of CTIAC systems will modify the heat delivery characteristics to the systems installed in the gas turbine exhaust, a steam generator in the case under study, and necessitate analysis of the effect on the performance of the whole power plant [1–3]. The introduction of an intake air cooler in an operating power plant allows producing a higher power than the nominal rating power by cooling the air below ISO conditions, depending on the conditions outside the plant [4]. This intake air cooling has limits; * Corresponding author. Tel.: +34 954 48 72 42; fax: +34 954 48 72 43. E-mail address: [email protected] (R. Chacartegui). 0196-8904/$ - see front matter Ó 2008 Elsevier Ltd. All rights reserved. doi:10.1016/j.enconman.2008.02.023
on the one hand, are the gas turbine mechanical characteristics and alternator power limitation, while on the other, ice crystals may appear in the gas turbine intake, which can generate erosion and wear of the intake vanes. So, in practice, intake temperature does not go below 43–45 °F (6.1–7.2 °C) [1,3,5]. Installation of CTIAC systems in operating plants presents some associated inconveniences, as they require investment in equipment and maintenance costs. In some cases, it even requires remodelling of some gas turbine sections. Besides, the equipments installed at the turbine air intake sections create additional pressure drops that will depend on the CTIAC technology installed. Also, the profitability of introducing a CTIAC system depends on many external factors that are not controllable, like climatology, electricity tariffs and fuel price. So, it requires a particular viability analysis for each power plant. 2. Combustion turbine inlet air cooling systems In summary, the main CTIAC technologies are the following [6–9]: 2.1. Evaporative coolers [10–13] They use water evaporation to cool the air. The water evaporates as it absorbs heat from the incoming air and reduces the
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Nomenclature Cp CH CDS COP CPI CTIAC EA FOG IGV GN HR HRSG Hp IIR m MH NPV OMEL P r RT SH T TES TIT W Wc w
constant pressure specific heat chiller percentage of cooling demand satisfied coefficient of performance consumer price index combustion turbine inlet air cooling absorption chillers fogging cooling system inlet guide vanes natural gas heat rate heat recovery steam generator lower calorific value internal rate of return mass flow (kg/s) evaporative media cooling system net present value electricity market operator price compression ratio refrigeration tons (1 RT equals 12,000 Btu/h or 3.52 kW) hybrid systems temperature thermal energy storage systems turbine inlet temperature power compression work humidity
dry bulb temperature. Evaporative systems can be classified as direct evaporative systems and indirect evaporative systems. In indirect evaporative systems, water is not brought in direct contact with the incoming air, water evaporation and the decrease in dry bulb temperature is done within an intermediate air stream flow, and it cools the compressor intake air in an additional heat exchanger [10]. In direct evaporative systems, water is brought in contact with the incoming air, and it evolves with the compressor intake stream flow. The main direct evaporative cooling systems are the evaporative media (MH) [11] and fogging systems (FOG). The minimum temperature is limited to 10 °C to avoid the risk of freezing at the intake of the compressor. They are capable of cooling inlet air to within 85–90% of saturation. In these evaporative systems, the evolution of fine water drops can produce an increase of extra power [14]. In addition to increasing the incoming mass flow, evaporation of drops will absorb latent heat, reducing the temperature of the mass that passes through the compressor [2], reducing the compression work that depends on the intake flow temperature, Eq. (1) [10] Wc ¼
C p T 01 ðr ðc1Þ=c c
1Þ
ð1Þ
Reduced NOx emissions with this CTIAC are attributed to the increased thermal heat capacity because of the water and the decreased compressor discharge temperature [5]. In evaporative media, the air passes over surfaces where water circulates. MH systems have the lowest installation costs of all the CTIAC systems. Also, they have a very low parasitic load (power consumed by the cooling system and auxiliary systems). Another additional advantage is that MH systems use potable water (cheaper than deionised water). Their main drawbacks are a high dependence on ambient conditions (temperature and relative humidity evolution), and the intake pressure drop even takes place when the cooling system is not working.
Subscripts 01 stagnation inlet a air amb ambient temperature c compressor cc combustion chamber cs single effect absorption cycle cd double effect absorption cycle disp available elec electric motor driven chiller f fuel g gas turbine exhaust h high pressure steam i initial conditions ice ice storage system mg gas engine driven chiller smed medium pressure steam t turbine tv steam turbine driven chiller w cold water storage system Greek symbols c specific heat ratio g efficiency
In fogging systems, high pressure water is injected [12,13]. These systems can produce nearly saturated air due to the very small drop size, even with reduced flow residence times, and so, they have a very high effectiveness, about 0.95–1. The grill of injectors also does not disturb the flow, and the pressures drops are very low. These systems have the possibility of overspray operation, or high fogging, with higher mass flow than would be needed to saturate the intake air, to force an ‘‘intercooling” of the air passing through the compressor [15]. Fogging systems have the second lowest acquisition costs of all CTIAC systems. Also, they have a low parasitic load. As drawbacks, these systems have a limited cooling capacity, down to the wet bulb temperature, and a high ambient dependence, they consume deionised water, are more expensive and present corrosion problems. Estimated installation costs and parasitic loads for evaporative systems are shown in Table 1. 2.2. Mechanical cooling compression or chillers (CH) They use a classical mechanical compression cycle, with an intermediate refrigerant fluid, which cools the intake air passing through heat exchangers located in the gas turbine inlet [7,16]. These systems can produce a stronger cooling effect than the evaporative systems, cooling the air down to 45 °F (7.2 °C). Estimated installation costs and mean parasitic load for these systems are compared in Table 2, differing by the form of driving the compressor among chillers with electric motor drive (CHelec), chillers with Table 1 Estimated investment costs and parasitic loads for evaporative systems
Investment cost (€/kW added) Parasitic load (% extra power generated)
MH
FOG
25–60 0.3–0.5
30–70 0.5–0.7
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Table 2 Estimated investment costs, parasitic loads and COP for mechanical and absorption chillers
Investment cost (€/RT) Parasitic load (kW/RT) COP
CHelec
CHmg
CHtv
EAcs
EAcd
870 0.81 2 and 6
1.225 0.187 1 and 2.5
1.300 0.195 1 and 2.5
1.000 0.31 0.7
1.130 0.265 1.2
gas engine drive (CHmg) and chillers with steam turbine drive (CHtv). Values are given as a function of the refrigeration capacity installed in RT (refrigeration tons). 2.3. Absorption chillers (EA) They use an absorption refrigeration cycle with an absorbent fluid and a refrigerating fluid. The working fluids used are water/ ammonia [17] or lithium bromide/water [3]. Although the ammonia presents better refrigerating properties than water, for toxicity problems, the technology of lithium bromide/water [6,18–20] is used. The evaporation temperature limits the inlet air cooling to 10 °C, higher than the temperature reached with mechanical chillers. Estimated mean installation costs and parasitic loads for the single effect LiBr–H2O absorption refrigeration cycle (EAcs) and the double effect LiBr–H2O absorption refrigeration cycle (EAcd) are shown in Table 2. The double effect LiBr–H2O absorption refrigeration cycle improves the performance of the single effect cycle. Also, different heat sources have been considered, direct heat source in an auxiliary boiler or taking advantage of the residual heat of the plant, for the single effect cycle with hot water at 115 °C and for the double effect cycle, medium pressure steam (8.5–10 bar). 2.4. Hybrid systems They are combinations of two or more of the previously mentioned systems. Their main purpose is to give operating flexibility to the cooling system to cover the demand, avoiding high parasitic load (electricity consumption) in periods with high electricity tariffs. The most usual combinations for hybrid systems are mechanical chillers (mainly electric motor driven chillers) in combination with absorption chillers. The use of evaporative coolers combined with mechanical chillers is not effective. 2.5. Thermal energy storage systems A thermal energy storage system (TES) allows storing cold (heat) so that a delay exists between the cold fluid storage (charging) and its consumption (discharging). Their feature is double, the main one is to avoid parasitic load consumption in ‘‘on-peak” periods, and the second one is to be able to cover cooling demand peaks with smaller cooling systems [13]. The use of continuous cooling is of interest when cooling the intake is needed for any one period greater than six hours [6]. They are usually installed with electric chillers, although they can also be used with hybrid systems. According to their storage form, these systems will be classified as cooled water storage systems (TESw), ice storage systems (TESice) and eutectic salt mixtures storage systems [21–24].
Cooled water storage systems store sensible heat energy; they are easy to install and have quick response on demand. They are more profitable for large storage sizes. Ice storage systems store energy in form of latent heat energy for which they are as much as six times smaller than a cooled water storage system of the same heat storage capacity. They work at a lower temperature of evaporation, and this penalizes the COP of the plant. A wide variety of types exists, but one of quicker response is the Ice Harvester. The estimated costs for each storage system are given in Table 3. 3. Power plant description The power plant under study is located in the south of Spain, in Cádiz, next to the sea. The cogeneration plant is inside a chemical plant, and it is composed of a General Electric gas turbine, model PG6541B, installed in 1995, and a heat recovery boiler. The turbine works in an open cycle and is composed of an axial compressor with 17 stages and a compression ratio of 1:11, and an axial turbine with 3 stages with a turbine inlet temperature (TIT) of 1100 °C. This gas turbine can be fuelled with natural gas or with naphtha. Their nominal characteristics values are Nominal Power ðGNÞ ¼ 38:340 kW Nominal Heat Rate ðGNÞ ¼ 11:460 kJ=kW h The gas turbine is coupled to a natural circulation heat recovery steam generator (HRSG), without post combustion, which takes advantage of the residual heat of the exhaust gases. In the HRSG, superheated steam is generated at two pressures (10 and 64 bars) and thermal oil is heated for its use in the chemical processes of the plant (290–345 °C). The HRSG generated mass flows and stream flow temperatures are given in Table 4. Of the electric power generated in the gas turbine, part is consumed in the chemical plant, and the remainder is exported to the electric network. Also, from the generated steam, part is used in their own chemical plant, and the remainder is exported to an adjacent refinery, while all the thermal oil is consumed in the chemical plant. The thermal oil mass flow and temperature are imposed by the demands of the chemical external process. In this part of the work, a model of the power plant was developed to be used to evaluate the economic effect of inserting different CTIAC systems for this power plant [7,3,25]. In 2000, were remodelled the gas turbine, changing the vanes of the second and third stages and the IGV (inlet guide vanes). As a consequence, neither the original curves of the gas turbine, GE PG6541B, nor those corresponding to a similar model after the remodelling, GE PG6551, were a valid solution for direct application in the power plant model [24]. After the modification, GE provided two correlations for the power and the heat rate, but they did not include the other necessary variables for study of the cogeneration power plant performance, like the exhaust flow gases and the exhaust temperature. On the other hand, for the HRSG, isolated and insufficient information was provided for its model. So, a model was generated through 2005 registered operational data. It was generated using a filtrate of the data, discarding anomalous or not valid data for the model that was desired. The data corresponding to operative
Table 4 HRSG generated mass flows and stream flow temperatures Table 3 Estimated investment costs for thermal storage systems
€/RT h
TESw
TESice
157.5–2.75 10-3 RT h for 2000 < RT h < 20,000
145 €/RT h
Steam 64 bar Steam 10 bar Thermal Oil
Production (T/h)
Temperature (°C)
40.7 13.7 395
480 187 290 and 345
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conditions in which the values of power delivery by the gas turbine that indicated start conditions or stopping conditions, that show operating with washing on line or that with strongly partial operating conditions were discarded. Also, the data of generation of steam that indicated a heating or stopping process in the HRSG or a stack diverter by pass were discarded. After this filtration, about 85% of the 2005 time registers were valid data to prepare the model.
The model of the gas turbine was prepared for base load operation, about 95% of the annual performance of the plant. The used expressions to model power, heat rate and exhaust temperature in the plant as functions of the intake temperature, relative humidity, nature of fuel and pressure drops are given in Appendix 1. Figs. 1–4 represent the values for these three parameters obtained from the operation registers as well as those obtained from the correlations given by the manufacturer.
Fig. 1. Registered power data as function of intake temperature.
44 Original manufacturer correlation
42
Correlation from measured power
Power (MW)
40 Manufacturer correlation after gas turbine modification
38 36 34 32 30 0
5
10
15
20
25
30
35
40
Intake Temperature (ºC) Fig. 2. Power calculated correlation from measured data and GE power correlations before and after gas turbine modifications.
Fig. 3. Heat rate obtained from 2005 registered data as function of intake temperature.
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13000
Heat Rate (kJ/kWh)
12500
12000
11500 Original Heat Rate correlation Heat Rate from measured data Heat Rate correlation after gas turbine modification
11000
10500 0
10
20
30
40
Intake Temperature (ºC) Fig. 4. Heat rate calculated correlation from measured data and GE HR correlations before and after gas turbine modifications.
With the measured values, or those estimated from them, the expressions for the model of the gas turbine were adjusted, having the gas turbine inlet temperature, relative humidity, fuel and pressure drops, as independent variables and obtaining the power and heat rate as a result of the inlet conditions from the expressions. The dispersion of data can be due to differences in the composition of the fuel from the one expressed in the correlations, pollution of filters, uncertainty in measurements, imperfect fit of the correlations given by the manufacturer, or safety regulation controls that continuously vary the gas turbine performance around a fixed operating point (gas generator speed, TIT, exhaust gas temperature etc.). For the joint model of the gas turbine–HRSG, the registered exhaust temperature was used, as is shown in Fig. 5, developing a correlation with a 95% confidence interval. Thus, obtaining the correlation given in Eq. (2) where the last term shows the confidence interval between both signs T g ð CÞ ¼ 0:722T amb þ 537:49 5:72
ð2Þ
The exhaust gases flow was correlated with the intake conditions through the global energy balance in the gas turbine, Eq. (3). The correlation obtained is shown in Eq. (5). The original correlation for the exhaust gases flow and the new correlation obtained are shown in Fig. 6
Fig. 5. Exhaust gases temperature correlation obtained from measured data.
ma C pa T amb þ mf Hp gcc ¼
W þ ðma þ mf ÞC pg T g gt
ð3Þ
where the efficiency of the turbine was correlated by the expression
Fig. 6. Exhaust gases mass flow, original correlation and new correlation obtained from measured data.
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gt ¼ 0:0008W þ 0:9203 mg ¼ 0:4489T amb þ 138:29 1:48
ð4Þ ð5Þ
W disp ðkWÞ ¼ 179:4 T amb þ 87080
ð6Þ
45
Steam mass flow (Tn/h)
With these expressions for the exhaust temperature and exhaust mass flow, the model of the HRSG and the power plant with base load operation was constructed with the independent variables being the mass flow and exhaust temperature, related directly with the gas turbine intake conditions [26]. As results, the model gives the steam production and the thermal oil production as a function of a unique variable that includes, simultaneously, the mass flow and the exhaust temperature in the form of available power in the exhaust gases, Fig. 7. The obtained correlation for this variable is shown in Eq. (6). Starting from the steam and thermal oil productions expressed as a function of available power, Fig. 7, as well as the relationship among the production of each one of the flows, Fig. 8, conditioned by the HRSG design, Fig. 9, and the heat demand of the chemical plant where the power plant is placed, the expressions shown in Appendix 1 were obtained. With this model, the economic study of the different CTIAC systems, as functions of the prices of fuel and electricity, as well as the costs of the equipments (investment, maintenance, others) can be conducted
50
40 35 30 25 20 15 10 High pressure steam Medium pressure steam
5 0 10000
11000
12000
13000
14000
The relationship among the power supplied with the thermal oil flow and the steam flows generated is shown in Fig. 10 where, due to the thermal oil temperature ranges, the thermal oil economizer is situated between the medium pressure steam generator and the high pressure steam generator. The HRSG control system adjusts
Available power inexhaust gases (kW)
87000 86000 85000 84000 83000 82000 81000 80000 5
10
15
20
25
30
35
40
IntakeTemperature (ºC) Fig. 7. Available exhaust gases heat power as function of intake temperature.
70000
HRSG Heat Balance (kW)
60000 Total Heat High pressure steam Thermal oil Medium pressure steam
50000 40000 30000 20000 10000 0 80000
81000
82000
83000
16000
Fig. 9. Relationship among steam mass flows and thermal oil heat demand.
88000
0
15000
Thermal oil power consumption (kW)
84000
85000
86000
Exhaust gases available heat power (kW) Fig. 8. HRSG heat balance.
87000
88000
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Fig. 10. HRSG scheme. Stack gases temperature (°C), high pressure steam flow and medium pressure steam flow evaporators and thermal oil economizer.
the thermal oil temperature, and when the thermal oil demand decreases, the high pressure sections absorb more heat and the disposal heat to the medium pressure sections of the steam generator decrease. 4. CTIAC systems modelling In this section, the treatment given to the different cooling systems, as well as the main hypotheses used, is discussed. Depending on the inlet air conditions, the cooling evolution in the chillers will be different. Different cooling possibilities are shown in the psychometric chart, Fig. 11. Three zones can be distinguished in this psychometric chart, delimited by two lines, one horizontal and one vertical, one line of dry bulb constant temperature and another line of constant relative humidity. The first is determined by the cooling temperature limit of the chiller, 7.2 °C in the case of mechanical chillers and 10 °C with absorption chillers, while the second line corresponds to constant absolute humidity with the value of absolute humidity given by the intersection point with the limit dry bulb temperature. Zone 1 corresponds to all the points with dry bulb temperature less than the limit temperature; therefore, the final conditions of any point with these initial conditions within this zone are the same as the initial conditions of temperature for any humidity. Zones 2 and 3 are those in which the temperature of the air is above the limit temperature, and therefore, they can be cooled by the chiller. The difference resides in the absolute humidity of the points of the zones, the first has a lower humidity than the suitable one and the second ha a greater humidity. The points within zone 2 evolve according to an absolute constant humidity line; they consume sensible heat but not latent heat, since they can never arrive at the saturation. Their limit is exactly the consigned temperature limit, although if the cooling capacity of the chiller is not enough, the final temperature will be greater than the limit temperature of the chiller. For the points belonging to zone 3, two different evolution forms exist, one for the initial conditions marked in the figure as type 1, in which the cooling capacity of the chiller is not able to cool the air sufficiently to take it to the sat-
uration line, and the other for the initial conditions marked in the figure as type 2, in which the cooling capacity of the chiller is able to take the initial point to the saturation and to continue cooling it, following the saturation line. Given the initial conditions (Ti, Wi) and the cooling capacity of the selected chiller, it is possible to calculate the maximum cooling that could be provided for any initial point and the final temperature that could be reached, depending upon which of the three ways of cooling is chosen. The intake pressure drop penalties in power and heat rate have been considered, and in the case of inserting a heat exchanger in the exhaust, as in the case of absorption chillers, the additional pressure drops have been included and the power and heat rate corrected using the expressions shown in Appendix 1. In the case of absorption chillers, the maximum heat that can be recovered from the stack is conditioned by the stack exhaust temperature that can be reached without corrosion problems due to condensation on the tubes. In the case of a Li–Br single effect absorption chiller, the maximum cooling capacity that can be used, using hot water, will be 300 RT, while for a Li–Br double effect absorption chiller, using steam, it can reach a maximum capacity of 500 RT. Hourly electricity tariffs have been considered. In Fig. 12, the hourly prices evolution of representative days taken for January and July are shown. In the viability analysis of hybrid systems, and later of the TES systems, this treatment of the electricity prices throughout the day is fundamental to the analysis of the chiller that will operate in determined periods to minimize the operating costs of the group, obtaining the maximum benefits of the power plant. For thermal energy storage systems, a daily storage system was chosen, developing a typical day model for each month, with the associated hourly electricity tariffs, taking a series of representative days such that the amount of cooling reached 80% of the cooling demand of the month. In Fig. 13, the mean chiller cooling capacity and minimum storage capacity obtained for the year using these representative days to cover the whole cooling demand are shown.
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Fig. 11. Chiller cooling treatment as function of the initial conditions and the chiller cooling capacity.
From Fig. 13, two main conclusions can be extracted. The first is obvious; the required mean cooling power is increased in summer, and the second is that the associated cooling storage minimum capacity does not follow the demand shape exactly, reaching its maximum values in months in which the relative time delay between the cooling demand and cooling supply is higher.
5. Economic considerations of the analysis The main aspects and hypotheses of the economic analysis are shown in this section. The phases of the economic analysis are summarized in the following steps: Fig. 12. Hourly electricity prices.
– Election of a cooling system with a refrigeration capacity. – Evaluation of the benefits of the group before and after implantation of the cooling system. – Differential evaluation of the costs and benefits of operation (electricity, steam, thermal oil, natural gas, deionised water). – Evaluation of the maintenance costs and investment. – Economic analysis of the solution. – Return to the first point, varying the capacity of the equipment.
Fig. 13. Evaluated monthly chiller mean cooling power and related minimum storage capacity.
Prices of steam, thermal oil and natural gas have been taken from 2005 data. Also, the electricity tariffs have been taken from the OMEL. The annual evaluation of costs and benefits has been divided into four periods, corresponding to the quarterly revisions of the rates of natural gas. A power plant operation factor of 0.94 has been considered over the total annual time. It affects the benefits, the operating costs and the maintenance costs. As for other economic aspects: the capital cost has been considered as 7.77%. The refund period and the
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considered redemption period have been taken as 10 years. Although this value is high for evaporative media, it is adjusted for the other cooling equipments, and it has been taken with the same time horizon for all them. A null salvage value and a yearly foreseen CPI of 2.9% have been considered . The tax rate on the economic activity is 33%; it affects the difference between the raw benefit and the depreciation of the installation. While the costs associated with the fuel are the most important operating cost, there are other operating costs that have been kept in mind also; among them are the natural water cost, deionised water cost and the cost of parasitic electricity consumption in the cooling equipment. The cost of natural water only has to be considered in wet media evaporative systems because, for the other systems, this consumption is null. For deionised water, necessary for the additional production of steam in the HRSG and for operation of the fogging system, its price is not only considerable, but in addition, it is necessary to keep in mind the maximum capacity of deionised water production in the plant because having to install new process lines for deionised water excessively diminishes the productiveness of the fogging system. In this case, the company has enough capacity to cover the demand of the fogging systems considered in this work with the increase in steam production, so the price of deionised water is taken as about 1.52 €/ Tm. In this plant, the increase of steam production can become an important part of the benefits (it reaches a value of 20% of the realised profit for electricity). The cost of using steam will take the same value as its selling cost. The price of the additional production of steam of high and medium pressure is given by Eqs. (7) and (8) 1:108P f ð664 15Þ þA 0:86 0:9 1:108Pf ð670 15Þ þA ð€=TmÞ ¼ 0:86 0:9
Psmed ð€=TmÞ ¼
ð7Þ
Psh
ð8Þ
where Pf is the price of fuel (€/kW h), A is the price of demineralised water (€/Tm).
6. Results As a result of this work, an extensive economic analysis was obtained for the cogeneration plant considered, showing the effects of the different CTIAC systems in the power plant performance. In Figs. 14–17, the results obtained for the best configuration studied for each CTIAC system are shown. In Fig. 14, the investments for the best configurations of each CTIAC system are compared; in Fig. 15, the annual benefits are shown for these configurations, and the internal rates of return (IIR) are shown in Fig. 16. The net present values (NPV) of the configurations are shown in Fig. 17. The analysis of evaporative cooling systems showed that, as a certain number of stages in the fogging system is surpassed, the raw annual benefit begins to diminish as well as the NPV, and then, the increase of stages is not translated into an increase of benefits. In these cases, the cooling system would be unnecessarily large. The most interesting evaporation system for this power plant is the fogging system with two stages. The analysis of mechanically driven chillers showed that steam turbine driven chillers have a very low productiveness. This is due to the appraised price of the steam in this installation being high, being more interesting in its direct sale than its use in a chiller to obtain extra additional power. With electrically driven chillers or gas engine driven chillers, we found the best results with refrigeration capacities around 750 RT for the electrically driven chiller and 700 RT for the gas engine driven chiller. With higher refrigeration capacities, although the percentage of cooling demand covered was increased, the annual incomes were not increased in the same manner, which indicates that most of the time, the CTIAC system was over designed. There is a better solution with intermediate refrigeration capacities due to two opposite effects; on one hand, as the cooling capacity of the chiller increases, a great part of this power is used in cooling saturated air, which gives a low ratio between sensible cooling and power consumption, and on the other hand, is the fact that not reaching the minimum temperature limit is a loss of additional electricity production and additional steam production.
Fig. 14. Equipment investment for the best economic configuration of every CTIAC system.
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Fig. 15. Annual benefits for the best economic configuration of every CTIAC system.
Fig. 16. Internal rate of return for the best economic configuration of every CTIAC system.
Fig. 17. Net present value for the best economic configuration of every CTIAC system.
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Table 5 Economic results obtained for the best configuration of each CTIAC system System Evaporative media Fogging of two stages Electric Chiller 850 RT Gas chiller 700 RT Steam chiller 250 RT Absorption chiller simple effect 300 RT Absorption chiller double effect 250 RT Absorption chiller simple effect and Electric chiller 850 RT Electric chiller – gas chiller 575–275 RT Electric chiller – gas chiller 250–500 RT Electric chiller – gas chiller 425–425 RT Electric chiller 600 RT – TES 500 RT h
Investment (€)
Incomes (€/year)
Expenses (€//year)
Benefits (€/year)
Payback (years)
IIR (%)
NPV (€)
CDS (%)
74,000 30,666 850,425 943,250 357,500 330,000 352,000 906,683
192,416 139,764 904,645 996,634 443,490 438,853 420,562 933,403
151,318 109,873 596,317 704,838 344,633 267,406 322,764 603,838
41,098 29,891 308,328 291,796 98,857 171,447 97,798 329,565
3 1.5 4.5 5 6 3 5.5 4.5
40.5 69 21.22 16.42 13.32 34.2 13.45 21.30
146,642 124,311 875,020 722,526 215,845 595,532 214,809 937,055
100 100 73.5 64.1 25.8 37.8 32.2 73.5
974,225 951,591 1,027,837 680,050
984,701 994,084 1,027,830 768,467
660,909 677,191 697,031 482,819
323,792 316,893 330,779 285,648
4.5 4.5 5 3.5
18.53 18.59 17.57 30.64
857,006 840,186 850,753 892,887
73.5 67.4 73.5 73.5
With absorption chillers, the double effect absorption chiller (direct fired) was shown not to be appropriate for this application, and it cannot compete with the other chillers of similar capacities. On the other hand, the single effect absorption chiller is a very interesting solution, although it is desirable that its cooling capacity would be bigger, being limited in its maximum capacity by the available heat in the stack exhaust compatible with the steam generation and thermal oil demand. This is due to its low parasitic load. For hybrid systems, two different combinations were checked, the first one was composed of an electric motor driven chiller and a single effect absorption chiller, which works with hot water produced with the HRSG stack exhaust gases. Given the high productiveness of the single effect absorption chiller, due mainly to its very low operating costs, in this hybrid system setup, the single effect absorption chiller will always work and the electric chiller will act when the demand requires it. The other combination for hybrid systems checked was composed of an electric motor driven chiller and a gas engine driven chiller, the operation modes of both are similar and simply vary the form of work. The main philosophy of this hybrid group was to use the gas engine driven chiller whenever the price of electricity was high to diminish the parasitic load electricity consumption and to maximize the electricity sale. From the results obtained with this configuration, the electric motor driven chiller alone gives better economic performance, and it is not worthwhile installing a hybrid system with this chillers combination. This is due fundamentally to the investment needed with the gas engine driven chiller, higher than that of an electric motor driven chiller, and the operating costs for natural gas, so the savings obtained using gas engine driven chillers, because of its very low parasitic load, is counteracted by the other two factors. Finally, the results of CTIAC systems composed of thermal energy storage combined with electric driven chillers were analyzed. The results obtained showed that the configurations with lower storage capacities were better than those with greater storage capacities due to the fact that in the summery periods, the TES system does not provide an advantage because the demand is higher than the chiller cooling capacity, while in some months, the demand is always below the cooling capacity of the chiller and, although the consumption is the same, the redistribution of parasitic load at lower electricity tariff hours does not compensate for the investment increase. Finally, in those periods in which an oscillation of demand occurs above and below the chiller cooling capacity, it was shown that with a 500 RT h storage capacity, it is enough to cover acceptably the cooling demand peaks. In fact, when the demand is satisfied with the TES system, to raise the storage size affects only the hourly redistribution of the parasitic load, diminishing it at the hours with higher electricity tariff.
In Table 5, the economic results obtained for the best configuration of each CTIAC system are shown. The values presented are investment, incomes, expenses, benefits, payback, IIR, NPV and percentage of cooling demand satisfied (CDS) by the cooling system installed.
7. Conclusions From this work the following conclusions are derived. The main features and costs of the different CTIAC systems have been identified, and this has allowed identifying the most feasible types of CTIAC systems to be introduced in this cogeneration power plant, ranging from very profitable systems with low initial investment, as the evaporative systems are, to others with higher investments but more profitable at the end of their capital refund period, as the hybrid systems with thermal storage are. With the model of the plant implemented, as developed from registered performance data, it has been possible to eliminate some uncertainty in the previous information concerning the power plant equipment. This model has allowed obtaining the global variables of interest for the economic model as functions of the gas turbine intake conditions. The statistical treatment of the operation registers of the power plant obtained in 2005, with the corresponding filtration of operative conditions that were outside the study conditions, has generated a precise model of the power plant for easy coupling of the gas turbine intake with the exit from different cooling systems. From the economic analysis, it is shown that installing a CTIAC system can be a very profitable solution, which also improves the gas turbine efficiency, saves fuel and reduces the CO2 and NOx emissions. From the data shown in Table 5, it is seen that the three most interesting intake cooling systems for this power plant are an electrically driven chiller of 850 RT cooling capacity, a hybrid system with an absorption chiller of 300 RT cooling capacity disposed sequentially with an electrically driven chiller of 450 RT cooling capacity and an electrically driven chiller of 600 RT cooling capacity combined with an Ice Harvester thermal energy storage system with 500 RT h of storage capacity. Of the comparison between the electric motor driven chiller and the electric motor driven chiller with TES system, it is concluded that both systems present very similar results with the exception that the TES system has a bigger NPV with a lower investment, and so, the system with storage would be a more interesting option. Taking the criteria of NPV as the main economic reference, the CTIAC system chosen would be the hybrid system with a value of the NPV 45.000 € greater than the NPV of the TES system, but with a worse IIR and payback period. However, other factors should also
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be contemplated, the main ones of them being the simplicity and installation time. Under these considerations, the system with thermal storage is clearly favourable relative to the hybrid system. The joint chiller plus TES is not a compact group, but it is simpler than the hybrid system with two chillers and will present a reasonable volume for an industrial plant; for a tank of 500 RT h, its volume would be about 35.5 m3. The installation time of the hybrid cooling system is bigger, not only because it is necessary to install two chillers and their interconnections and controls, but also because one of them is an absorption chiller working with hot water generated in the final section of the HRSG. All the necessary actions in the HRSG clearly exceed the installation time of the TES system. As we are studying the enhancement of an existing operating plant, the available time for the cooling system installation will be scheduled to coincide with the power plant planned overhauls. With the hybrid cooling system, the risk of not finishing the installation works in time is higher. The cost of one day with the chemical plant stopped is about 50,000 €, greater than the difference of final benefit between the two options. Also, the control of the plant is easier for the TES system than for the hybrid system, since the control becomes independent of the operation in the HRSG. Under all these considerations, the option of the electrically driven chiller with ice storage would be the CTIAC system chosen for this power plant. Appendix 1. Power plant model equations Gas turbine Power W W W ðkWÞ ¼ ð42487:878 236:51T a 0:256T 2a ÞF W wabs F Hf F Pc
ðA:1Þ
Power corrections FW wabs ¼ 1:00096 0:1519wabs FW Hf FW pc
ðA:2Þ
100 þ ð10:254 3:466 log Hf þ 0:0837 log H2f Þ ¼ 100 0:015D inch H2 O ¼1 4
ðA:3Þ ðA:4Þ
Heat rate HR ðkJ=kW hÞ ¼ ð2731:796 þ 2:859T a HR HR þ 0:064T 2a Þ4:18F HR wabs F Hf F Pc
ðA:5Þ
Heat rate corrections F HR wabs ¼ 0:9977 þ 0:35443wabs 100 þ ð4:50628 þ 1:31713 log Hf þ 100 0:005D inch H2 O ¼1þ 4
F HR Hf ¼ F HR pc
ðA:6Þ 0:01431 log H2f Þ
ðA:7Þ ðA:8Þ
Exhaust gas temperature T g ð CÞ ¼ 0:722T a þ 537:49 5:72
ðA:9Þ
Exhaust mass flow mg ðkg=sÞ ¼ 0:4489T a þ 138:29 1:48
ðA:10Þ
HRSG Exhaust gas available power W disp ðkWÞ ¼ 179:4T a þ 87080
ðA:11Þ
Total heat delivery to cogeneration W cons:tot ðkWÞ ¼ 0:5993W disp þ 6268:5
ðA:12Þ
Heat delivery to medium pressure steam W vapmed ðkWÞ ¼ 0:2339W disp 11503
ðA:13Þ
Heat delivery to high pressure steam W vaphigh ðkWÞ ¼ 0:3654W disp þ 17771:5 W th
oil
ðA:14Þ
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