- Email: [email protected]
off-design performance simulation and discussion for the gas turbine combined cycle with inlet air heating
Energy 178 (2019) 386e399
Contents lists available at ScienceDirect
Energy journal homepage: www.elsevier.com/locate/energy
Design/off-design performance simulation and discussion for the gas turbine combined cycle with inlet air heating Yongping Yang a, *, Ziwei Bai a, Guoqiang Zhang a, *, Yongyi Li a, Ziyu Wang b, Guangying Yu b a
National Thermal Power Engineering and Technology Research Center, Beijing Key Laboratory of Emission Surveillance and Control for Thermal Power Generation, North China Electric Power University, Changping District, Beijing, 102206, China Department of Mechanical and Industrial Engineering, Northeastern University, Boston, MA, 02115-5000, USA
b
a r t i c l e i n f o
a b s t r a c t
Article history: Received 24 April 2018 Received in revised form 17 April 2019 Accepted 21 April 2019 Available online 25 April 2019
A novel method for the combined cycle retrofitted with compressor inlet air heating (HEAT) process is proposed in this work for part load performance improvement. The performance of the new cycle is researched with two regulations with HEAT and compared with another two strategies with compressor inlet guide vane (IGV). For the novel method, the compressor inlet air is preheated by part of the exhaust flue gas from the heat recovery steam generator. Moreover, comparation result indicates that the novel strategy with HEAT to maintain design turbine inlet temperature (T3) and then to keep turbine exhaust temperature (T4) at its maximum value (HEAT-T3-T4) is suggested for load arrangement. If being compared with the cycle using IGV control to maintain design T3 then keep extreme T4 (IGV-T3-T4), a maximum combined cycle efficiency increment as 1.7% pt. could be obtained by the cycle with HEAT-T3T4. Both energy analysis and exergy analysis are given in this paper. The simulation results indicate that higher compressor outlet temperature (T2) decreases the extra exergy destruction during combustion, so that the performance of the cycle with HEAT is improved. Meanwhile the specific work of each cycle has also been presented. Furthermore the pressure loss affection of the compressor inlet air heat exchanger on the combined cycle performance is investigated in range of 0%e5%. Each 1% pressure loss is calculated to decrease 1.4% on the design power output, and 0.25% pt. on the combined cycle efficiency, which will weaken the performance improvement. To summarize, this paper proposes a novel method for the combined cycle power output adjustment to obviously improve its part load efficiency. This innovative solution has never been suggested before and it could be combined with other methods for better result. © 2019 Elsevier Ltd. All rights reserved.
Keywords: Gas turbine Compressor inlet heating Combined cycle Off-design performance Thermodynamic analysis
1. Introduction Gas turbine combined cycle (GTCC) has been worldwide researched [1] for performance improvement because of its largescale application [2]. The design compressor pressure ratio (PR) and T3 of heavy-duty gas turbines are continuously increased for design efficiency and power output increment [3]. Consequently, triple-pressure reheat heat recovery steam generator (HRSG) is normally utilized due to continuously increasing T4. Thermal power plants work under off-design (namely part-load) situations frequently to satisfy the requirement of the power grid. While the thermodynamic optimization is addressed under design condition, * Corresponding authors. E-mail addresses: [email protected] (G. Zhang).
(Y.
Yang),
https://doi.org/10.1016/j.energy.2019.04.136 0360-5442/© 2019 Elsevier Ltd. All rights reserved.
[email protected]
so that the combined cycle operation drifts away from benchmark and the cycle efficiency deteriorates under part load conditions. Along with the commercialization efforts, a number of researches have been extensively conducted on operation regulations and part load performance improvement. IGV control is widely applied for improving the part-load performance of heavyeduty gas turbine combined cycles [4], which results in increasing T4 with constant T3 demand. Aguilar et al. [5] compared the effects of several operation strategies on cycle efficiency, and the turbine variable area nozzle strategy leads to the best part load performance. Except for strategies only, the traditional combined cycle configuration has also been modified for improving the performance of GTCC under off-design conditions. An additional compressor was proposed before the compressor of GTCC for sliding pressurization, in which way the part load efficiency of the combined cycle was increased [6]. Zhang et al. [7] studied a GTCC
Y. Yang et al. / Energy 178 (2019) 386e399
Nomenclature
Symbols A
Cp
E h
L LHV m
Ma n
p Q T
U w W △p
△T
turbine inlet area [m2 ] specific heat capacity at constant pressure [J/(kg $ K)] exergy of a given working fluid [MW] enthalpy [kJ/kg] the theoretical air quantity [kg/kg fuel] the lower heating value [kJ/kg] mass flow rate [kg/s] the Mach number rotational speed [r/min] pressure [kPa] heat transfer capacity [kJ] temperature [K] overall heat transfer coefficient [W/(m2 $ K)] specific work [kJ/kg] power output [MW] pressure loss [kPa] Log-mean temperature difference [ C]
Greek Letters a vane outlet absolute flow angle b the excess air coefficient ε HEAT rate g the specific heat capacity 4 flow coefficient z the mixing loss coefficient k the comprehensive parameter (set as 0.404 in this work) h thermal efficiency
h j
exergy efficiency pressure coefficient
Subscripts a ca cc d f g
and superscripts air the cooling air combustion chamber design condition fuel flue gas
with extra recuperation cycle inserting into the HRSG for off-design performance improvement. Liu et al. [8] investigated a GTCC with partially recuperative HRSG which could receive a higher thermal efficiency but lower specific power output, comparing with traditional GTCC. Li et al. [9] found a novel method to improve part-load efficiency of GTCC by adjusting the backpressure of HRSG with corresponding operation strategies which could improve the cycle efficiency of 1.76% pt. at most. It is well known that the performance of a gas turbine cycle (namely topping cycle), with [10] or without HRSG [11], is directly affected by the compressor inlet air temperature (T1). The effects of ambient temperature (T0) has been evaluated on micro gas turbines [12], gas turbines [13] and gas turbine combined cycles [14]. Similar conclusion has been conducted as the cycle performances are degraded when T1 increases [15], because the air density reduces and the compressor inlet mass flow decreases. Considering the significant impact of T1 on the cycle performance, compressor inlet air cooling technology is considered as an effective method for
gt in out s st total 0 1 2 3 4 5 * Acronyms C CC EC EV GT HEAT HP HPT HRSG IGV IP IPT LP LPT PR RH SH ST T0 T1 T2 T3 T4 T5 *
387
gas turbine inlet outlet steam/water steam turbine combined cycle ambient condition compressor inlet compressor outlet/combustion chamber inlet combustion chamber outlet/gas turbine inlet gas turbine outlet/HRSG inlet HRSG outlet stagnation value
compressor combustion chamber economizer evaporator gas turbine compressor inlet air heating high pressure high-pressure steam turbine heat recover steam generator inlet guide vane intermediate pressure intermediate-pressure steam turbine low pressure low-pressure steam turbine pressure ratio reheater superheater steam turbine ambient temperature compressor inlet temperature compressor outlet temperature gas turbine inlet temperature gas turbine exhaust temperature HRSG exhaust temperature All the operating conditions were assumed to be under steady state and every parameters were referring to total condition during simulation
enhancing the performance of gas turbine power plants [16]. The generated power for a selected gas turbine power plant is expected to be increased by 1% via decreasing T1 for 1 C [17]. Wang and Chiou [18] claimed that the thermal efficiency and power output of a given gas turbine cycle (namely topping cycle) could be increased by 5.16% and 12% with inlet absorption cooling system. LNG cryogenic energy was suggested to be indirectly utilized in the compressor inlet air cooling process to further improve the part load performance of the GTCC [19]. Hosseini et al. [20] investigated the influence of T1 on the operation of power plants with different evaporative media in the intake-air cooler system. It should be mentioned that the impact of T1 on the GTCC cycle performance has only been widely researched related to the inlet air cooling system and the seasonal change affection. It has never been employed as an operation strategy for cycle load adjustment or applied for off-design performance improvement yet. In this paper, a novel method to improve the part load thermal efficiency of GTCC by heating compressor inlet air is presented to
388
Y. Yang et al. / Energy 178 (2019) 386e399
adjust the cycle power output capacity while maintain cycle efficiency. The thermodynamic analysis including exergy analysis for the two pairs of strategies are carried out through modeling and simulation to highlight the improvement of HEAT. Then, critical characteristics parameters in terms of T3, T4, PR, mass flow, specific work and cycle efficiencies are compared and discussed. Finally, the effect of the pressure loss of the compressor inlet heat exchanger on the cycle performance is discussed in range of 0%e5%. 2. Characteristics of the GTCC with HEAT/IGV 2.1. Description of the combined cycles 2.1.1. Unit with IGV The traditional S109FA gas turbine combined cycle is selected as the basic unit in this study, which consists of a PG9351FA gas turbine and a matching triple-pressure reheat HRSG. The selected topping cycle is composed of an 18 stage axial flow compressor, a low NOx permission combustion and a 3 stage gas turbine (GT) from GE. A triple-pressure, reheat, horizontal, natural circulation HRSG without supplementary fire is selected as the matching steam turbine cycle (namely bottoming cycle) for the basic system which comprises high-pressure (HP), intermediate-pressure (IP), and low-pressure (LP) steam subsystems. Each subsystem consists of a feed water pump, an economizer (EC), an evaporator (EV), a superheater (SH) and a corresponding steam turbine (ST). In addition, attemperators are installed between superheaters and reheaters (RH) for safe operation. The schematic diagram of the conventional GTCC is shown in Fig. 1. Firstly, inlet air is compressed and then mixed with fuel. Next, they are burned in the combustion chamber. The burned gas with high temperature drives the GT to generate electrical power. Finally, the exhaust gas from the gas turbine outlet is utilized to generate electricity by the ST with the help of HRSG. The main design parameters of the traditional combined cycle are summarized in Table 1. 2.1.2. Unit with HEAT The proposed GTCC with HEAT is studied with the retrofitting
Table 1 Main parameters of the basic combined cycle. Facility (Unit) Compressor Pressure ratio Inlet air mass flow (kgas1) Efficiency (%) Outlet temperature ( C) Gas turbine Combustion pressure loss (%) Combustion heat loss (%) Fuel lower heating value (kJakg1) Inlet temperature ( C) Exhaust temperature ( C) Power output (MW) Topping cycle efficiency (%) HRSG Main steam temperature ( C) Main steam pressure (kPa) Steam turbine power output (MW) Exhaust gas temperature ( C) Bottoming cycle efficiency (%) GTCC Mechanical loss (%) Generator loss (%) Power outputa (MW) Combined cycle efficiencya (%)
Value 15.4 641.97 88.1 391.8 3.5 0.5 48989.915 1327 614.9 262.3 37.5 567.5 9800 147.1 99.7 33.9 1 1 401.25 57.34
a Mechanical loss and generator loss are considered during the combined cycle power output calculation.
layout as presented in Fig. 2. The compressor inlet air is heated by part of the HRSG exhaust gas in the compressor inlet heat exchanger. The rest part of the combined cycle hardware parameters are designed to have similar geometry parameters with the traditional combined cycle. The flow rate of the heat exchanger hot side can be adjusted to modify T1. The HEAT rate ε is defined as the proportion of the HRSG exhaust gas flow into the heat exchanger. The logarithmic heat transfer temperature difference of the compressor inlet heat exchanger is set as more than 25 C in whole load conditions, resulting in the T1 range of 15 Ce65 C at standard environment conditions.
Fig. 1. Schematic of the basic combined cycle.
Y. Yang et al. / Energy 178 (2019) 386e399
389
Fig. 2. Scheme of the compressor inlet air heating process.
2.2. Part-load operation strategy of the combined cycle with HEAT/ IGV For further study the influence of HEAT/IGV on combined cycle and detailed performance comparison, both the traditional unit and the modified unit are proposed with two corresponding regulations. T4 is significant both to the topping and bottoming cycle performances. Thus T4 is suggested to maintain at its limit value, resulting in strategy of IGV-T3-T4 as well as HEAT-T3-T4.
manner is adopted to decrease the load. Similarly, T3 is firstly maintained at 1327 C for the strategy of IGV-T3-T4 with variable IGV. And then T4 is kept at its maximum constantly by adjusting the fuel consumption rate and further decreasing the IGV angle from 71.3 (83.3%) to 49 (60%), in which process the power output of the combined cycle is reduced from 84.6% to 49.1%. Finally, the IGV angle is fixed and the fuel-only load manner is adopted for load decrement. 3. Combined cycle modeling
2.2.1. Strategy with HEAT Two strategies with HEAT are named as HEAT-T3 (use HEAT control to increase T1 and maintain T3 to decrease load þ keep T1 steady and only decrease fuel to decrease load) and HEAT-T3-T4 (use HEAT control to increase T1 and maintain T3 to decrease load þ use HEAT control to increase T1 and maintain T4 to decrease load þ keep T1 steady and only decrease fuel to decrease load) to highlight the characteristics of the method. IGV angle is kept constant during these strategies with HEAT control. In details for the strategy of HEAT-T3, T3 is firstly maintained at the initial temperature (1327 C) with T1 increasing from 15 C to 57 C by the heat exchanger and T4 increasing from 614.9 C to 650.9 C (maximum allowable temperature) at the same time. Simultaneously, HEAT rate ε is changed from 0% to 80%, in which process the combined cycle power output is reduced from 100% to 81%. Then, T1 is fixed and only the fuel consumption rate is regulated to further reduce the cycle load. As for the strategy of HEATT3-T4: Firstly, T4 is increasing from 614.9 C to 650.9 C when T1 increases from 15 C to 57 C. And then T4 remains unchanged by adjusting the fuel consumption rate. Consequently T1 is further increased from 57 C to 65 C, in which process T3 decreases and the power output drops from 81% to 76.1%. Finally, εreaches its limit (100%) and fuel only control is applied for the cycle load regulation. 2.2.2. Strategy with IGV To verify the accuracy of the models and for comparative purposes, the traditional PG9351FA combined cycle with IGV-T3 [21] and IGV-T3-T4 are also simulated by the gas turbine combined cycle models under design/off-design conditions. To be more specific, for the strategy of IGV-T3: The combined cycle is first operated with IGV angle adjustment to keep T3 at 1327 C constantly, leading to T4 increases from 615 C to 651 C. In which process the IGV angle varies from 88 (100%) to 71.3 (83.3%) when the combined cycle load is 84.6%e100%. Once T4 reaches its limit, the IGV angle will maintained constant and the fuel-only load
In this paper, the components had two kinds of operational modes: design and off-design. The gas turbine combined cycle was first simulated with design models in order to size the units of the system. This was the model parameterization step, so that key parameters such as the stage outlet parameters and overall parameters of compressor could been got. Subsequently, the offdesign model of the system was utilized in which the sizes of the process units were assigned to the values determined in the design model. The combined cycle models were modified in order to predict their behaviors at off-design conditions. Excel (for the topping cycle) and the commercial software Aspen Plus V11.1 [22] (for the bottoming cycle and the recuperator) were utilized to connect the components and calculate the design models. The offdesign performance was modeled using software Matlab [2]. The gas turbine combined cycle was simulated according to the mass/energy balance and chemical equilibrium based on the process layout and most important thermodynamic design parameters reported in the previous literature [23]. Each of the component under full load conditions was modeled respectively，and the overall representation of the gas turbine combined cycle was carried out by connecting them appropriately by means of thermodynamic and mechanic links. It was assumed that ambient air was at standard condition (15 C, 1.013 bar and 60% relative humidity). The fuel was assumed to be injected into the combustion chamber with the same temperature of the compressor inlet air. The rotational speed of the gas turbine and steam turbines were stable at 3000 r/min in all conditions. The fuel lower heating value was assumed as 48989.915 kJ/ kg for all systems. The off-design parameters for the units were assumed to be under steady state and obtained by simulation results from design models. The thermo-physical properties of air and flue gas were calculated by the method proposed by Zhang [24]. In this paper, it should be mentioned that the specific heat capacity of air and flue gas were determined by temperature, component, and
390
Y. Yang et al. / Energy 178 (2019) 386e399
pressure. The detailed design/off-design calculation process of the basic combined cycle can be referred to our earlier work [25]. The cycle efficiency and specific work were defined as follows:
hgt ¼
Wgt mf LHV
hst ¼
(4)
1 ¼ dðtan aÞ d f
(2)
Wtotal ¼ Wgt þ Wst
htotal ¼
(3)
The compressor performance under off-design condition was modeled based on the one-dimensional sequential stage-stacking method [28], with parameters obtained from the design model. The main function of IGV was adjusted to a certain limit to generate arbitrary power during the part-load performance state [29]. The variable compressor inlet mass flow affected by the inlet temperature was selected by the matching relationship between the compressor and the gas turbine during off-design condition with modification based on the characteristics curves mentioned in Ref. [30]. The details of the compressor simulation and IGV control were presented in the previous literature [2] with following functions [31]:
(1)
Wst h4g h0g mg
Wtotal mf LHV
3.1. Compressor modeling
j Wgt þ Wst w¼ ma þ mf
(5)
where W is the power output, h is the thermal efficiency, m is the mass flow, LHV is the fuel lower heating value, h is the enthalpy, P is the pressure, T is the temperature, and w is the specific work. Subscript and Superscript: gt means gas turbine cycle, st means steam turbine cycle, total means combined cycle, f means fuel, a means air, g means flue gas, whereas 0 represents the ambient condition，1 represents the compressor inlet, 2 represents the compressor outlet, 3 represents the gas turbine inlet, 4 represents the gas turbine outlet. Furthermore, the exergy analysis was presented based on Ref [26] and Ref [27] in this paper. The physical exergy of each working fluid was obtained by simulating through Aspen Plus based on its temperature, pressure and component. The exergy efficiencies were defined as follows. In this paper, each thermal efficiency is given as efficiency for short, and the exergy efficiency will be directly named if needed.
hcc
¼
E3g
(6)
f
¼ constant
¼
3.2. Combustion chamber modeling The thermodynamic characteristics of the combustion chamber were simulated with focus on the determination of the excess air coefficient and combustion temperature. The pressure loss during combustion was assumed as 3.5% of the combustion chamber inlet pressure under design/off-design conditions as mentioned in our previous work [2]. Fuel was assumed to be heated into the same temperature as the compressor inlet air before combustion. The thermo-physical properties of gas turbine flue gas mainly relied on its composition and temperature. The method proposed by Zhang [24] was used to calculate the thermo-physical properties of flue gas in different states in the system. The thermal balance equation [32] was shown as following:
ð1 þ bLÞ h3g h1g ¼ h2f h1f þ bL h2a h1a þ hcc LHV
Ef
hst ¼
(12)
Wtur Wcom
E4g
htotal ¼
Wst
(11)
where a is the vane outlet absolute flow angle, 4 is the flow coefficient, j is the pressure coefficient.
E2a þ Ef hgt
(10)
(7)
E4g
Where b is the excess air coefficient, L is the theoretical air quantity, h is the enthalpy, h is the efficiency, and LHV is the fuel lower heating value. Subscripts and Superscripts: cc means the combustion chamber, g means gas, f means fuel, a means air, 1 represents the compressor inlet, 2 represents the compressor outlet, and 3 presents the combustion chamber outlet.
(8)
E5g
3.3. Gas turbine modeling
Wtur Wcom þ Wst
(9)
Ef
Where W is the power output or the power consumption, h is the exergy efficiency, E is the exergy of a given working fluid, Subscript and Superscript: gt means gas turbine cycle, st means steam turbine cycle, total means combined cycle, f means fuel, a means air, g means flue gas, whereas 0 represents the ambient condition，1 represents the compressor inlet, 2 represents the compressor outlet, 3 represents the gas turbine inlet, 4 represents the gas turbine outlet, 5 represents the HRSG outlet.
A gas turbine model was considered with three separate turbine stage cooling assumption [2] as shown in Fig. 3, considering inlet pressure losses [33]. Each stage had a pair of nozzle vane and rotor blade rows. The gas turbine exhaust pressure was set as 0.1038 MPa in all conditions for safety operation. The turbine expansion ratios of the first stage was assumed to be constant; however, the turbine expansion ratios of the rest two stages changed based on the characteristics curve shown in Ref. [32] under part load conditions. In the off-deign conditions, the relationship between the turbine inlet pressure, temperature and mass flow were described as the Flügel equation [34].
Y. Yang et al. / Energy 178 (2019) 386e399
mca ¼ mca;d
391
p2* p2* d
!
T 2* d T 2*
!0:5 (14)
where m is the mass flow rate, p is the pressure, and T is the temperature. Subscripts and Superscripts: d means design condition, ca means the cooling air, 2 represents the compressor outlet, and * represents the stagnation condition. The cooling air was accelerated by the main flue gas, thus its velocity and direction were consistent with those of the main flue gas. The pressure drop during coolant mixing was calculated as [33]:
Dp* p*gt;in
Fig. 3. Schematic of the gas turbine cooled air.
pffiffiffi m T ¼ constant AP
(13)
where m is the inlet mass flow rate of gas turbine, A is the turbine inlet area, T is turbine initial temperature and P is the turbine inlet pressure. The blade cooling model was assumed as shown in Fig. 4. The cooling air quantity at each turbine stage at design condition was estimated by Zhang [35]. The cooling air of the nozzle vanes was mixed with the main flue gas at the inlet. The mixed flue gas expanded in both the nozzle vanes and the rotor blade rows. Finally, the cooling air of the corresponding rotor blade rows was mixed with the expanded flue gas at the outlet of this stage. The cooling air quantity of each turbine stage at off-design conditions can be adjusted according to the temperature and pressure of the bleeding stage [36].
mca m ¼ g Ma2g z ¼ ca k; Dp* < 0 mg g mg
where△p is the pressure loss, p is the pressure, m is the mass flow, g is the specific heat capacity, Ma is the Mach number, z is the mixing loss coefficient, and k is the comprehensive parameter which is set as 0.404 in this work [35]. Subscripts: gt is the gas turbine, g is the gas, ca is the cooling air, and in is the inlet. Superscripts: * is the stagnation condition.
3.4. HRSG modeling The triple-pressure bottoming cycle was modeled by modifying the simplified method for single-pressure HRSG proposed by Ganapathy [37] as summarized in previous literature [25]. The energy balance equation and heat transfer equilibrium equation were used for model calculation as following. In details, the evaporators were simulated as shown in Equation (16), while the rest of heat exchangers in HRSG were calculated by Equations 16e18 together.
Q ¼ mg cp tg;in tg;out ¼ ms hs;out hs;in
(16)
Q ¼ UA D T
(17)
DT ¼
Fig. 4. Blade cooling model adopted for the cooled stages.
(15)
tg;out ts;in tg;in ts;out ln tg;out ts;in tg;in ts;out
(18)
where m is the mass flow, Cp is the specific heat capacity at constant pressure, t is the temperature, h is the enthalpy, U is the overall heat transfer coefficient and A is heat transferring area, △T is the log-mean temperature difference, Q is the heat transfer capacity. Subscript: in means the inlet of heat exchanger, out means the outlet, s means the steam/water side, and g means the gas side. The off-design characteristic of the steam turbine can be described by the Flügel formula [38]. The exhaust temperature of HRSG in design condition was maintained at 100 C for both traditional cycle and retrofitted cycle, by adjusting the design thermal parameters of the bottoming cycle system in reasonable ranges. The simulation followed the flow sequence to analyses the parameter characteristics of the combined cycle under full load conditions. The design water/steam/gas side pressure loss of each heater was set between 2% and 5% and corrected under the offdesign condition [39]. It should be mentioned that since the design parameters of HRSG configuration cannot be adjusted under off-design conditions, the exhaust temperature of HRSG changed with its inlet temperature (T4). This value was set as 99.7 C under design condition and calculated by the off-design modeling.
392
Y. Yang et al. / Energy 178 (2019) 386e399
3.5. Recuperator modeling In this study, the entire heat transfer process of the heat exchange at the inlet of compressor was assumed stable under full load. The pressure loss in the heat exchanger was neglected during Section 4.1-4.4 and discussed in range of 0%e5% in Section 4.5. The mass flow rate of the heat exchanger inlet flue gas was adjusted to balance the heat transfer equilibrium equation [37]:
mg cp;g Tg;in Tg;out ¼ ma cp;a Ta;out Ta;in
(19)
where m is the mass flow, Cp is the specific heat at constant pressure, Tis the temperature. Subscripts and Superscripts: g means gas, a means air, and in means the inlet of recuperator, whereas out means the outlet of recuperator. 4. Results and discussions 4.1. Simulation verification Fig. 5. The influence of inlet air temperature on the gas turbine power output.
Table 2 shows several characteristics parameters of five key operating points (A/B/C/D/E operation points in the following figures) in different strategies. The design gas turbine power output and efficiency of PG9351FA unit are 262.3 MW and 37.5% which are similar with the data mentioned in Gas Turbine World 2004e2005 handbook [40]. Fig. 5 compares the simulation result and the data in Ref. [30] and Ref [41] of the gas turbine power output capacity (with T3 at 1327 C constantly) affected by T1 to validate the models. Gas turbine performance highly depends on the compressor inlet temperature [41]. The simulation result indicates that the gas turbine can produce more power with lower compressor inlet temperature. It is calculated that the difference of the gas turbine cycle power output capacity between 5 C and 55 C is 106.7 MW. 4.2. Off-design characteristic analysis of the topping cycle Fig. 6 presents the running lines of the two strategies with IGV control. The characteristic map of the compressor with inlet angle ranging from 88 (100%) to 49 (60%) is simulated by the stagestacking method as mentioned before. Depending on the matching relationship between the compressor and the turbine, the working points can be specified in the models by the Flügel formula. In addition, each characteristic curve in the figure possesses its own IGV opening with constant rotational speed. The running lines of strategies with HEAT method are shown in Fig. 22 for clearly discussion.
Fig. 6. Operation characteristic curve of compressor with IGV control.
Fig. 7 presents the variation of the gas turbine cycle thermal efficiency and T4 with various compressor inlet temperature at full load capacity, at which T3 is maintained at its design value during simulation. The curves indicate that T1 negatively affects the cycle
Table 2 The key parameters of the main operating conditions in the four strategiesa. Description
Design condition
Inflexion of HEAT-T3
Inflexion of HEAT-T3-T4
Inflexion of IGV-T3
Inflexion of IGV-T3-T4
Point T1 ( C) IGV Compressor pressure ratio Inlet air mass flow (kg/s) Gas turbine power output (MW) Gas turbine cycle efficiency (%) T3 ( C) T4 ( C) Main steam mass flow (kg/s) Steam turbine Power output (MW) Steam turbine cycle efficiency (%) Combined cycle power outputa (MW) Combined cycle Efficiencya (%)
A 15 88 (100%) 15.4 642.0 262.29 37.48 1326.8 614.9 80 147.1 33.90 401.25 57.34
B 57 88 (100%) 13.2 549.9 194.73 34.34 1327.1 650.8 72.1 137.0 34.76 325.14 57.33
C 65 88 (100%) 12.7 532.9 179.02 33.46 1313.4 651.3 69.6 132.7 34.77 305.54 57.1
D 15 71.3 (83.3%) 12.8 535.8 212.25 35.24 1327.0 652.0 70.2 134.0 34.62 339.34 56.34
E 15 49 (60%) 8.8 390.0 105.90 27.36 1176.6 651.5 51.7 95.0 34.03 196.9 50.88
a
Mechanical loss and generator loss are considered during the combined cycle power output calculation.
Y. Yang et al. / Energy 178 (2019) 386e399
393
Fig. 9. Variation in turbine inlet temperature with combined cycle power output.
Fig. 7. The effect of inlet air temperature on the gas turbine cycle.
efficiency which can be explained by the fact that PR decreases with the increasing T1, leading to the decrement of turbine expansion ratio and the increment of T4. This conclusion is in agreement with previous statements [42]. Considering HRSG can recover the energy of the turbine exhaust flue gas, the bottoming cycle power ability is gradually improved by the increasing T4. The variation trends of the gas turbine combined cycles with the four operation strategies in terms of T2, T3, compressor PR, T4 and turbine exhaust mass flow are shown in Figs. 8e12, respectively. A phenomenon of “parallel line” can be observed in these figures under relative low load, because the same regulation of “fuel-only control” is employed in this range. As expected, T2 is affected by T1 with a rising trend for combined cycles with HEAT control under decreasing load conditions. T3 is firstly maintained constant with decreasing load for all strategies. When T4 reaches its limit, the combined cycle load with IGV control decreases to 84.6% (point D), while the combined cycle load with HEAT control is declined to 81% (point B). Then T3 is constantly decreased to 1176.6 C for IGV-T3-T4 and to 1313.4 C for HEAT-T3-T4 under cycle load 49.1% (point E)
Fig. 10. Variation in compressor pressure ratio with combined cycle power output.
Fig. 11. Variation in turbine exhaust temperature with combined cycle power output. Fig. 8. Variation in compressor outlet temperature with combined cycle power output.
394
Y. Yang et al. / Energy 178 (2019) 386e399
Fig. 12. Variation in turbine exhaust mass flow with combined cycle power output.
much worse. Particularly, the gap between IGV-T3-T4 and HEAT-T3T4 is firstly enlarged and then disappeared. At combined cycle load of 84.6% (point D), the topping cycle efficiency of IGV-T3/IGV-T3-T4 is 35.24% and that of HEAT-T3/HEAT-T3-T4 is 35.19%. The Inferiority of the performance of IGV-T3-T4/HEAT-T3-T4, comparing with that of IGV-T3/HEAT-T3 at relative low load (45%e70%), is mainly due to the decrement of PR which plays a significant role in gas turbine cycle performance. From a thermodynamic perspective, the results in Fig. 13 indicate that the cycles with IGV control result in relative higher topping cycle efficiency comparing with the cycles with HEAT control with load of 70%e100%. Furthermore, the operation strategy of HEAT-T3 is treated as an advisable choice only next to IGV-T3 to improve the performance of the gas turbine cycle in low load condition (45%e70%). Moreover, the exergy analysis on the topping cycle is presented in Figs. 14 and 15, respectively. The irreversibility loss in the combustion chamber is the largest part among the gas turbine cycle which is mainly affected by T2 [26]. Thanks to the increasing T2 resulting from the HEAT process, the combustion chamber exergy efficiency is improved and the fuel consumption rate is decreased. As a result, the topping cycle exergy efficiency with HEAT strategy is higher than that with IGV control under relative high load conditions (70%e100%).
and 76.1% (point C), respectively. The compressor PR for the strategy of HEAT-T3/IGV-T3 related to the T4 peak value is 13.2 (point B)/12.8 (point D), respectively. The PR of IGV-T3 is performed the steepest slope (A-D) when the cycle load decreases from 100% (point A) to 84.6% (point D), as presented in Fig. 10. This can be explained by the fact that the PR positively affects the temperature difference between T3 and T4. Furthermore, the relationship between the turbine inlet pressure, temperature and mass flow rate is described as the Flügel formula. Under the assumption of constant T3 (A-B&A-D), the curve trends of the mass flow rates would be similar with that of PRs. When the load reduces from 81% to 76.1% (from 84.6% to 49.1%) for cycle with HEAT-T3-T4 (IGV-T3-T4), the PR would continually drops to 12.7 (8.8) but the slope would be more gentle influenced by the drop of T3 (BeC & D-E). Worthy of note is that, under the fuel only control condition, T4 drops constantly and the turbine exhaust flow increases slightly under relative low load for all strategies, as shown in Fig. 12. The variation of the topping cycle efficiency curves are summarized in Fig. 13. It is observed that the curve trends of the four strategies are similar, and IGV-T3 presents the highest topping cycle efficiency, while the performance of HEAT-T3-T4/IGV-T3-T4 are
In this study, the inlet temperature of the HRSG is directly assumed to be T4 and the HRSG exhaust flue gas is connected to the heat exchanger for the systems with HEAT-T3/HEAT-T3-T4. The heat loss is neglected in these processes. The variation curves of the bottoming cycle efficiencies are presented in Fig. 16. For the combined cycle load range of 45%e 100%, the variation intervals of the bottoming cycle efficiencies of unit with IGV-T3 (IGV-T3-T4) and HEAT-T3 (HEAT-T3-T4) are 27.6%e34.6% (31.4%e34.6%) and 26.4%e34.8% (28.3%e34.8%), respectively. It is well known that the bottoming cycle efficiency is highly dependent on T4 under part load conditions. As shown in Fig. 13, IGV-T3-T4 (HEAT-T3-T4) obtains higher T4 at the expense of reducing the topping cycle performance under low load condition, comparing with IGV-T3 (HEAT-T3), which results in higher bottoming cycle performance.
Fig. 13. Variation in topping cycle efficiency with combined cycle power output.
Fig. 14. Variation in combustion chamber exergy efficiency with combined cycle power output.
4.3. Off-design characteristic analysis of the bottoming cycle
Y. Yang et al. / Energy 178 (2019) 386e399
Fig. 15. Variation in topping cycle exergy efficiency with combined cycle power output.
395
Fig. 17. Variation in HRSG exhaust temperature with combined cycle power output.
Fig. 18. Variation in bottoming cycle exergy efficiency with combined cycle power output. Fig. 16. Variation in bottoming cycle efficiency with combined cycle power output.
4.4. Combined cycle design/off-design characteristics With HEAT process, the waste energy in the HRSG exhaust gas has been partly recovered. Additionally, the compressor inlet air is heated to reduce its mass flow rate because the inlet area is constant for a given compressor. The HEAT regulation presents a similar operating characteristics with IGV control but with narrower range of adjustment. The performance would obviously changes when the HEAT adjustment ends and switches to fuel only control. The outlet temperature of the HRSG (T5) is shown in Fig. 17 with range among 91 C-107 C. The exergy efficiencies of the bottoming cycles are presented in Fig. 18. As aformentioned, attemperators are utilized when T4 is higher than the design value for bottoming cycle safety operation. So the bottoming cycle exergy efficiency decreases with increasing T4, and this parameter recovers with decreasing T4 if it is still higher than the design value (615 C). After the operation strategy changes back to the fuel only control and T4 is lower than 615 C, the attemperators are closed and the bottoming cycle performance decreases with the decreasing T4 as presented.
The comparison of the combined cycle efficiencies for the two units with different operation strategies are demonstrated in Fig. 19. The minimum values of the combined cycle thermal efficiencies of systems with IGV-T3/IGV-T3-T4/HEAT-T3/HEAT-T3-T4 are 46.9%/49.7%/49.0%/49.7% within the load range of 45%e100% respectively. Evidently, the combined cycle efficiency curves show that the unit with HEAT-T3-T4 exhibits the best part load performance comparing with other strategies when the combined cycle load is above 48%. The difference between cycle performance of HEAT-T3T4 (IGV-T3-T4) and HEAT-T3 (IGV-T3) is 0.7% pt. (2.5% pt.) under low load. The difference between that of HEAT-T3 and IGV-T3 becomes 1.1% pt. at 85% cycle load and 1.9% at 50% cycle load. This is because that the irreversibility loss during combustion is reduced by the HEAT process and the fuel consumption rate is decreased. While the gap between HEAT-T3-T4 and IGV-T3-T4 is firstly increased to 1.7% pt. at 76.1% combined cycle load (point B) and then gradually diminished at 48% combined cycle load because of the narrow
396
Y. Yang et al. / Energy 178 (2019) 386e399
Fig. 19. Variation in combined cycle efficiency with combined cycle power output.
range of the HEAT process. During HEAT process, T1 increases, leading to the decrement of inlet air density. As the inlet area held constant for a given compressor, the mass flow rate would be expected gradually decrease. The compressor outlet air will receive a relative higher temperature and lower compressor pressure ratio which will decreases the fuel consumption rate with a constant or a suitable T3 for the requirement of the gas turbine. Considering the constant turbine exhaust pressure assumption, the decrement of turbine expand rate leads to the decreasing turbine power output and increasing T4. The strategy with HEAT decreases the irreversibility loss in topping cycle and reduces fuel consumption rate by increasing T2, resulting in a good load regulating method. The IGV operation range of compressor with IGV-T3-T4 is 60%e 100%, related to the combined cycle load range from 49.1% (point E) to100% (point A). But for the combined cycle unit with HEAT-T3-T4 this total load range regulated by the HEAT process could only varied from 100% to 76.1% (point C). The heat transfer temperature difference limits the range of the strategy. Thus fuel only control is applied at 76.1% combined cycle load and the combined cycle efficiency reduces rapidly. And, it should be mentioned that, we just compared the two pairs of operating strategy IGV-T3-T4 (and IGVT3) and HEAT-T3-T4 (and HEAT-T3). If the HEAT method is combined with IGV, there will be a foreseeable best operating strategy, which may be researched in the future. On the other hand, for combined cycle units that always participate in the pitching peak with power output larger than 48%, HEAT-T3-T4 strategy is highly recommended due to its high part load efficiency. The combined cycle exergy efficiency is shown in Fig. 20. The exergy efficiency curves show similar trend as for the combined cycle efficiency. As previously mentioned, the advantage of the HEAT strategy over the IGV method is mainly due to the improvement of the combustion chamber performance. Higher T2 would decreases the exergy loss during combustion and saves the fuel consumption rate with the same performance requirement, resulting in a better combine cycle efficiency and combine cycle exergy efficiency. From the view of the Second Law of Thermodynamic, the higher temperature of the outlet air from the compressor will greatly reduces the exergy destruction during the combustion process, which results in a higher cycle exergy efficiency. In Fig. 21, the specific works of the four strategies are presented. It is shown that the specific works of cycles with HEAT-T3/HEAT-T3-
Fig. 20. Variation in combined cycle exergy efficiency with combined cycle power output.
Fig. 21. Variation in the specific work with combined cycle power output.
T4 is decreased from 611.4 kJ/kg to 579.1 kJ/kg at load range from 100% (point A) to 81% (point B) due to HEAT process. The compressor can absorbs more than half of the total work produced by the gas turbine, and the specific work of the compressor would be gradually enhanced by the increasing T1. It is also well known that relative low operating T3 results in low specific work. For the combined cycle with HEAT-T3-T4, the specific work is decreased into 561.84 kJ/kg at cycle load 76.1% (point C) affected by the increasing T1 and decreasing T3. The specific power output of the cycle with IGV control is kept almost stable in relative high load. As for IGV-T3/IGV-T3-T4 control, the specific power output of the combined cycle slightly increases from 611.41 kJ/kg to 619.17 kJ/kg when the cycle load is above 84.6% (point D). As for the cycle with IGV-T3-T4 at 49.1% (point E)-84.6% (point D) load, the variation interval of the specific work is 494.94 kJ/kg-619.17 kJ/kg resulting from the changing T3. “Four parallel lines” could be observed in the figure because the same rule (fuel only control) is applied under relative low load. To reduce the load of a combined cycle, normally either T3 or the flue gas mass flow of the working medium would be reduced. But as for the unit with HEAT-T3/HEAT-T3-T4, the specific work is reduced
Y. Yang et al. / Energy 178 (2019) 386e399
397
by increasing the compressor consumption specific work, therefore the cycle performance keeps preferably within corresponding load range. 4.5. Parametric analysis of the pressure loss of the compressor inlet heat exchanger After discussing the combined cycle performances of the novel HEAT process, the effect of the pressure loss of the compressor inlet heat exchanger is analyzed in this part. Previous discussion on Figs. 6e21 are based on the simulated results assuming that the pressure loss of the compressor inlet heat exchanger is ignorable. However, this hypothesis is good for discussion but not very close to reality. In this part, the simulation results of a retrofitted combined cycle using a heat exchanger with 2% or 5% inlet pressure loss at each side are simulated and compared as shown in Figs. 22e25. During the simulation process, the pressure drop ratios remain unchanged for all systems under part load conditions. The performance effects of the pressure loss on the compressor running lines with two different strategies (HEAT-T3&HEAT-T3-T4) are illustrated in Fig. 22. The characteristic map of the compressor inlet temperature ranging from 15 C to 65 C with constant rotational speed is simulated by the off-design compressor model [32] with variable inlet air pressure and density. The flow characteristic and the volume flow rate of the compressor are not affected by the
Fig. 23. Variation in turbine expansion ratio with combined cycle power output.
changing inlet pressure, but the compressor inlet mass flow rate decreases resulting from the reducing inlet air density. As can be seen from the figure, the running lines are parallel and the pressure loss increases PR. But the inlet pressure is affected so that the
Fig. 22. Operation characteristic curves of compressor with HEAT control.
398
Y. Yang et al. / Energy 178 (2019) 386e399
Fig. 24. Variation in combined cycle efficiency with combined cycle power output for strategies with HEAT.
Fig. 25. Variation in combined cycle exergy efficiency with combined cycle power output for strategies with HEAT.
curves of the turbine expansion ratio coincide as compared in Fig. 23. The system with larger pressure loss related to higher compressor PR but similar compressor outlet pressure/turbine inlet pressure for a given compressor inlet mass flow condition. In addition, the inflexion of HEAT-T3 (point B) is declined from 57 C to 55 C and then to 49 C caused by the increasing pressure loss from 0% to 2% and then to 5%. This is because that the turbine expansion ratio decreases under the same cycle load with constant T3, so that T4 raises for a given operation condition. But the turbine exhaust temperature limit is steady all the time, thus the inflexion value changes. The pressure loss shows a negative effect on the topping cycle performance and increases T4, which will enhance the bottoming cycle efficiency slightly but still cannot compensated the performance decline. Considering the negative effect of the increasing backpressure on the cycle performance [9], the HRSG exhaust pressure is set as constant by adding a fan behind the heat exchanger hot side. The extra power consumption by the fan is simulated by the commercial software Aspen Plus V 11.1 [22] and deducted from the combined cycle power output. As depicted in Figs. 24 and 25, the combined cycle efficiency and exergy efficiency of cycles with two different strategies (HEAT-T3 & HEAT-T3-T4) and three different compressor inlet pressure losses (0%, 2%, 5%) are compared. The simulation results convey that the compressor inlet heat exchanger pressure loss of 2% could decreases the combined cycle efficiency for about 0.5% pt. and that of 5% decreases the efficiency for about 1.2% pt. This negative affection remains identical when T3 or T4 is stable and alleviates if fuel only control is utilized for cycle load control. It should be mentioned that the total power output ability is also decreased by the inlet pressure loss, which will affects the load
range of a certain combined cycle unit. Each 1% pressure loss decreases about 5.7 MW on the design combined cycle power output ability. As a result, the combined cycle performance prominently affected by the compressor inlet heat exchanger characteristics. The pressure loss shows a negative effects both on the combined cycle power output ability and cycle efficiency. Certainly, the heat exchanger for commercial utilization will not expected to have such large pressure drop equal to or higher than 2% [43], therefore the novel strategy will still expected to perform a satisfactory cycle efficiency with reasonable heat exchanger pressure drop assumption. 5. Conclusion Basis on the design parameters of the given gas turbine combined cycle, a special method named HEAT (heating the compressor inlet air by part of the HRSG exhaust gas) is purposed in this paper. Topping, bottoming, and combined cycle performances of the two pairs of operation strategies (HEAT-T3/IGV-T3 & HEAT-T3-T4/IGVT3-T4) are compared with both energy analysis and exergy analysis to highlight the advantage of the novel method on part load performance improvement. The cycle with HEAT-T3-T4 is suggested for its significant part load performance. A maximum efficiency increment of 1.7% pt. could be received comparing the cycle performance with HEAT-T3-T4 and the cycle with IGV-T3-T4 under part load conditions. Even taking the cycle with IGV-T3 for comparation, a maximum performance improvement of 1.1% pt. could still be obtained for that with HEAT-T3. The simulation result indicates that the combined cycle efficiency of the novel cycle with HEAT-T3-T4 is almost unchanged with the increasing T1 and constant T3, because the specific work is gradually reduced (from
Y. Yang et al. / Energy 178 (2019) 386e399
611.41 kJ/kg to 579.1 kJ/kg) for load adjustment. This specific work decrement is due to the increasing specific work consumed by the compressor rather than lowering T3 which would leads to a remarkably combined cycle efficiency drop. The efficiency improvement by HEAT control is mainly because the waste energy from the HRSG exhaust gas is absorbed by the compressor inlet air to increase the compressor outlet air temperature. Higher temperature of the compressor outlet air can effectively reduces the extra exergy destruction during the combustion process, so that the cycle performance is improved and the fuel consumption rate is saved. For restriction of the temperature difference in the compressor inlet heat exchanger, the adjustment range of the novel regulation is much narrower than that of the IGV control, resulting in the diminishment of the advantage on combined cycle performance when the cycle load is lower than 48%. Moreover, the affection of the pressure loss is discussed within range of 0%e5%. Each 1% pressure loss caused by the compressor inlet heat exchanger is found to decrease 1.4% on the design cycle power output ability, and 0.25% pt. on the combined cycle efficiency. Besides, the HEAT method can be combined with IGV method, which is a foreseeable best operating strategy that can be researched in the future. In all, the novel method HEAT is suggested as an innovative solution for gas turbine combined cycle load management. Acknowledgements This study was supported by National Natural Science Foundation of China of China (Grant No.51436006); National Natural Science Foundation of China of China (Grant No.51306049); Fundamental Research Funds for the Central Universities (No.2017MS15). References [1] Bakhshmand Sina Kazemi, Khoshbakhti Saray Rahim, Bahlouli Keyvan, et al. Exergoeconomic analysis and optimization of a triple-pressure combined cycle plant using evolutionary algorithm. Energy 2015;93:555e67. [2] Zhang Guoqiang, Zheng Jiongzhi, Xie Angjun, et al. Thermodynamic analysis of combined cycle under design/off-design conditions for its efficient design and operation. Energy Convers Manag 2016;126:76e88. [3] Matta RK, Mercer GD, Tuthill RS. Power systems for the 21st century-" H" gas turbine combined-cycles. NY: GE Power systems schenectady; 2000. [4] Haglind F. Variable geometry gas turbines for improving the part-load performance of marine combined cyclesecombined cycle performance. Appl Therm Eng 2011;31(4):467e76. [5] Aguilar Francisco Jimenez-Espadafor, Rodriguez Quintero R, Carvajal Trujillo E, et al. Analysis of regulation methods of a combined heat and power plant based on gas turbines. Energy 2014;72:574e89. [6] Barelli Linda, Ottaviano Andrea. Supercharged gas turbine combined cycle: an improvement in plant flexibility and efficiency. Energy 2015;81:615e26. [7] Zhang Guoqiang, Bai Ziwei, Yang Yongping. Analysis of recuperated combined cycle with small temperature rise under design/off-design conditions. Energy Procedia 2017;105:1507e12. [8] Liu Taohong, Zhang Guoqiang, Li Yongyi, et al. Performance analysis of partially recuperative gas turbine combined cycle under off-design conditions. Energy Convers Manag 2018;162:55e65. [9] Li Yongyi, Zhang Guoqiang, Bai Ziwei, et al. Backpressure adjustable gas turbine combined cycle: a method to improve part-load efficiency. Energy Convers Manag 2018;174:739e54. [10] De Lucia Maurizio, Lanfranchi Carlo, Boggio Vanni. Benefits of compressor inlet air cooling for gas turbine cogeneration plants. In: ASME 1995 International gas turbine and aeroengine congress and exposition. American Society of Mechanical Engineers; 1995. [11] Farzaneh-Gord Mahmood, Deymi-Dashtebayaz Mahdi. Effect of various inlet air cooling methods on gas turbine performance. Energy 2011;36(2): 1196e205. [12] Basrawi Firdaus, Yamada Takanobu, Nakanishi Kimio, et al. Effect of ambient temperature on the performance of micro gas turbine with cogeneration system in cold region. Appl Therm Eng 2011;31(6e7):1058e67. [13] De Sa Ashley, Zubaidy Sarim Al. Gas turbine performance at varying ambient temperature. Appl Therm Eng 2011;31(14e15):2735e9.
399
[14] Lee Jae Hong, Tong Seop Kim, Eui-hwan Kim. Prediction of power generation capacity of a gas turbine combined cycle cogeneration plant. Energy 2017;124:187e97. [15] Gareta R, Romeo LM, Gil A. Economic optimization of gas turbine inlet aircooling systems in combined cycle applications. In: 15th International conference on efficiency, costs, optimization, simulation and environmental impact of energy systems; 2002. [16] Ali Sohani, Farasati Yashar, Sayyaadi Hoseyn. A systematic approach to find the best road map for enhancement of a power plant with dew point inlet air pre-cooling of the air compressor. Energy Convers Manag 2017;150:463e84. [17] Thamir K Ibrahim, Rahman MM, Abdalla Ahmed N. Improvement of gas turbine performance based on inlet air cooling systems: a technical review. Int J Phys Sci 2011;6(4):620e7. [18] Wang FJ, Chiou JS. Integration of steam injection and inlet air cooling for a gas turbine generation system. Energy Convers Manag 2004;45(1):15e26. [19] Zhang Guoqiang, Zheng Jiongzhi, Yang Yongping, et al. A novel LNG cryogenic energy utilization method for inlet air cooling to improve the performance of combined cycle. Appl Energy 2016;179:638e49. [20] 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 Manag 2007;48(4):1055e64. [21] Kim JH, Song TW, Kim TS, et al. Model development and simulation of transient behavior of heavy duty gas turbines. J Eng Gas Turbines Power 2001;123(3):589e94. [22] Liu Zuming, Karimi Iftekhar A. New operating strategy for a combined cycle gas turbine power plant. Energy Convers Manag 2018;171:1675e84. [23] Wang Haozhi. The method of the S109FA unit‘s operation about deeply unloading in the abnormal working conditions Master’degree thesis. Guangzhou (China): South China University of Technology; 2010. [24] Zhang Shizheng. Polynomial expressions of thermodynamic properties of the products of combustion of fuel with air. J Eng Thermophys 1980;1:10e6 [in Chinese]. [25] Zhang Guoqiang, Zheng Jiongzhi, Yang Yongping, et al. Thermodynamic performance simulation and concise formulas for triple-pressure reheat HRSG of gasesteam combined cycle under off-design condition. Energy Convers Manag 2016;122:372e85. [26] Song TW, Sohn JL, Kim JH, et al. Exergy-based performance analysis of the heavy-duty gas turbine in part-load operating conditions[J]. Exergy. Int J 2002;2(2):105e12. [27] Zhu Mingshan. The exergy analysis of the energy system[M]. Tsinghua University; 1985 [in Chinese]. [28] Stone Aubrey. Effects of stage characteristics and matching on axial-flowcompressor performance. ASME; 1957. [29] Mehrpanahi A, Payganeh GH. Multi-objective optimization of IGV position in a heavy-duty gas turbine on part-load performance. Appl Therm Eng 2017;125: 1478e89. [30] Brooks Frank J. GE gas turbine performance characteristics. Schenectady, NY: GE Power Systems; 2000. [31] Song TW, Kim TS, Kim JH, et al. Performance prediction of axial flow compressors using stage characteristics and simultaneous calculation of interstage parameters. Proc IME J Power Energy 2001;215(1):89e98. [32] Jiao Shujian, Sun Shoulin, Zhang Yanchun. Gas turbine and gas-steam combined cycle system. first ed. Beijing: China Electric Power Press; 2007. p. 41e191 [in Chinese]. [33] Jonsson Maria, Bolland Olav, Bücker Dominikus, et al. Gas turbine cooling model for evaluation of novel cycles. Proceedings of ECOS 2005:20e2. [34] Palmer Carl A, Erbes Michael R, Pechtl Peter A. Gatecycle performance analysis of the LM 2500 gas turbine utilizing low heating value fuels, vol. 8. NEW YORK, NY(USA): ASME; 1993. p. 69e76. [35] Zhang Lili. Modeling and performance analysis for coal gas fired gas turbine. Institute of Engineering Thermophysics Chinese Academy of Sciences; 2010 [in Chinese]. [36] Gate Cycle. Gate cycle for windows version 5.51. r. General Electric Company; 2003. [37] Ganapathy V. Simplify heat recovery steam generator evaluation. Hydrocarb Process 1990;69(3). [38] Cui Ning. Study and application on real-time dynamic simulation model for heavy-duty gas turbine combined cycle power unit. 2008. PhD thesis North China Electric Power University, Baoding (China). [in Chinese]. [39] Wu Xiaobing. Parameter optimization and performance analysis on largescale combined-cycle cogeneration systems. Beijing (China): Tsinghua University; 2005. [40] Handbook. Gas turbine world 2004-2005. USA: Gas Turbine World; 2005. [41] Antonio Peretto, Bhargava Rakesh. A parametric evaluation of fogging technology for gas turbine performance enhancement. Bologna, Italy: Alma Mater Studiorum Universita’ Degli Studi Di Bologna; 2000. lez-Díaz Abigail, Alcara z-Caldero n Agustin M, Gonza lez-Díaz Maria [42] Gonza Ortencia, et al. Effect of the ambient conditions on gas turbine combined cycle power plants with post-combustion CO2 capture. Energy 2017;134:221e33. [43] Kim TS, Ro ST. Power augmentation of combined cycle power plants using cold energy of liquefied natural gas. Energy 2000;9(25):841e56.
Contents lists available at ScienceDirect
Energy journal homepage: www.elsevier.com/locate/energy
Design/off-design performance simulation and discussion for the gas turbine combined cycle with inlet air heating Yongping Yang a, *, Ziwei Bai a, Guoqiang Zhang a, *, Yongyi Li a, Ziyu Wang b, Guangying Yu b a
National Thermal Power Engineering and Technology Research Center, Beijing Key Laboratory of Emission Surveillance and Control for Thermal Power Generation, North China Electric Power University, Changping District, Beijing, 102206, China Department of Mechanical and Industrial Engineering, Northeastern University, Boston, MA, 02115-5000, USA
b
a r t i c l e i n f o
a b s t r a c t
Article history: Received 24 April 2018 Received in revised form 17 April 2019 Accepted 21 April 2019 Available online 25 April 2019
A novel method for the combined cycle retrofitted with compressor inlet air heating (HEAT) process is proposed in this work for part load performance improvement. The performance of the new cycle is researched with two regulations with HEAT and compared with another two strategies with compressor inlet guide vane (IGV). For the novel method, the compressor inlet air is preheated by part of the exhaust flue gas from the heat recovery steam generator. Moreover, comparation result indicates that the novel strategy with HEAT to maintain design turbine inlet temperature (T3) and then to keep turbine exhaust temperature (T4) at its maximum value (HEAT-T3-T4) is suggested for load arrangement. If being compared with the cycle using IGV control to maintain design T3 then keep extreme T4 (IGV-T3-T4), a maximum combined cycle efficiency increment as 1.7% pt. could be obtained by the cycle with HEAT-T3T4. Both energy analysis and exergy analysis are given in this paper. The simulation results indicate that higher compressor outlet temperature (T2) decreases the extra exergy destruction during combustion, so that the performance of the cycle with HEAT is improved. Meanwhile the specific work of each cycle has also been presented. Furthermore the pressure loss affection of the compressor inlet air heat exchanger on the combined cycle performance is investigated in range of 0%e5%. Each 1% pressure loss is calculated to decrease 1.4% on the design power output, and 0.25% pt. on the combined cycle efficiency, which will weaken the performance improvement. To summarize, this paper proposes a novel method for the combined cycle power output adjustment to obviously improve its part load efficiency. This innovative solution has never been suggested before and it could be combined with other methods for better result. © 2019 Elsevier Ltd. All rights reserved.
Keywords: Gas turbine Compressor inlet heating Combined cycle Off-design performance Thermodynamic analysis
1. Introduction Gas turbine combined cycle (GTCC) has been worldwide researched [1] for performance improvement because of its largescale application [2]. The design compressor pressure ratio (PR) and T3 of heavy-duty gas turbines are continuously increased for design efficiency and power output increment [3]. Consequently, triple-pressure reheat heat recovery steam generator (HRSG) is normally utilized due to continuously increasing T4. Thermal power plants work under off-design (namely part-load) situations frequently to satisfy the requirement of the power grid. While the thermodynamic optimization is addressed under design condition, * Corresponding authors. E-mail addresses: [email protected] (G. Zhang).
(Y.
Yang),
https://doi.org/10.1016/j.energy.2019.04.136 0360-5442/© 2019 Elsevier Ltd. All rights reserved.
[email protected]
so that the combined cycle operation drifts away from benchmark and the cycle efficiency deteriorates under part load conditions. Along with the commercialization efforts, a number of researches have been extensively conducted on operation regulations and part load performance improvement. IGV control is widely applied for improving the part-load performance of heavyeduty gas turbine combined cycles [4], which results in increasing T4 with constant T3 demand. Aguilar et al. [5] compared the effects of several operation strategies on cycle efficiency, and the turbine variable area nozzle strategy leads to the best part load performance. Except for strategies only, the traditional combined cycle configuration has also been modified for improving the performance of GTCC under off-design conditions. An additional compressor was proposed before the compressor of GTCC for sliding pressurization, in which way the part load efficiency of the combined cycle was increased [6]. Zhang et al. [7] studied a GTCC
Y. Yang et al. / Energy 178 (2019) 386e399
Nomenclature
Symbols A
Cp
E h
L LHV m
Ma n
p Q T
U w W △p
△T
turbine inlet area [m2 ] specific heat capacity at constant pressure [J/(kg $ K)] exergy of a given working fluid [MW] enthalpy [kJ/kg] the theoretical air quantity [kg/kg fuel] the lower heating value [kJ/kg] mass flow rate [kg/s] the Mach number rotational speed [r/min] pressure [kPa] heat transfer capacity [kJ] temperature [K] overall heat transfer coefficient [W/(m2 $ K)] specific work [kJ/kg] power output [MW] pressure loss [kPa] Log-mean temperature difference [ C]
Greek Letters a vane outlet absolute flow angle b the excess air coefficient ε HEAT rate g the specific heat capacity 4 flow coefficient z the mixing loss coefficient k the comprehensive parameter (set as 0.404 in this work) h thermal efficiency
h j
exergy efficiency pressure coefficient
Subscripts a ca cc d f g
and superscripts air the cooling air combustion chamber design condition fuel flue gas
with extra recuperation cycle inserting into the HRSG for off-design performance improvement. Liu et al. [8] investigated a GTCC with partially recuperative HRSG which could receive a higher thermal efficiency but lower specific power output, comparing with traditional GTCC. Li et al. [9] found a novel method to improve part-load efficiency of GTCC by adjusting the backpressure of HRSG with corresponding operation strategies which could improve the cycle efficiency of 1.76% pt. at most. It is well known that the performance of a gas turbine cycle (namely topping cycle), with [10] or without HRSG [11], is directly affected by the compressor inlet air temperature (T1). The effects of ambient temperature (T0) has been evaluated on micro gas turbines [12], gas turbines [13] and gas turbine combined cycles [14]. Similar conclusion has been conducted as the cycle performances are degraded when T1 increases [15], because the air density reduces and the compressor inlet mass flow decreases. Considering the significant impact of T1 on the cycle performance, compressor inlet air cooling technology is considered as an effective method for
gt in out s st total 0 1 2 3 4 5 * Acronyms C CC EC EV GT HEAT HP HPT HRSG IGV IP IPT LP LPT PR RH SH ST T0 T1 T2 T3 T4 T5 *
387
gas turbine inlet outlet steam/water steam turbine combined cycle ambient condition compressor inlet compressor outlet/combustion chamber inlet combustion chamber outlet/gas turbine inlet gas turbine outlet/HRSG inlet HRSG outlet stagnation value
compressor combustion chamber economizer evaporator gas turbine compressor inlet air heating high pressure high-pressure steam turbine heat recover steam generator inlet guide vane intermediate pressure intermediate-pressure steam turbine low pressure low-pressure steam turbine pressure ratio reheater superheater steam turbine ambient temperature compressor inlet temperature compressor outlet temperature gas turbine inlet temperature gas turbine exhaust temperature HRSG exhaust temperature All the operating conditions were assumed to be under steady state and every parameters were referring to total condition during simulation
enhancing the performance of gas turbine power plants [16]. The generated power for a selected gas turbine power plant is expected to be increased by 1% via decreasing T1 for 1 C [17]. Wang and Chiou [18] claimed that the thermal efficiency and power output of a given gas turbine cycle (namely topping cycle) could be increased by 5.16% and 12% with inlet absorption cooling system. LNG cryogenic energy was suggested to be indirectly utilized in the compressor inlet air cooling process to further improve the part load performance of the GTCC [19]. Hosseini et al. [20] investigated the influence of T1 on the operation of power plants with different evaporative media in the intake-air cooler system. It should be mentioned that the impact of T1 on the GTCC cycle performance has only been widely researched related to the inlet air cooling system and the seasonal change affection. It has never been employed as an operation strategy for cycle load adjustment or applied for off-design performance improvement yet. In this paper, a novel method to improve the part load thermal efficiency of GTCC by heating compressor inlet air is presented to
388
Y. Yang et al. / Energy 178 (2019) 386e399
adjust the cycle power output capacity while maintain cycle efficiency. The thermodynamic analysis including exergy analysis for the two pairs of strategies are carried out through modeling and simulation to highlight the improvement of HEAT. Then, critical characteristics parameters in terms of T3, T4, PR, mass flow, specific work and cycle efficiencies are compared and discussed. Finally, the effect of the pressure loss of the compressor inlet heat exchanger on the cycle performance is discussed in range of 0%e5%. 2. Characteristics of the GTCC with HEAT/IGV 2.1. Description of the combined cycles 2.1.1. Unit with IGV The traditional S109FA gas turbine combined cycle is selected as the basic unit in this study, which consists of a PG9351FA gas turbine and a matching triple-pressure reheat HRSG. The selected topping cycle is composed of an 18 stage axial flow compressor, a low NOx permission combustion and a 3 stage gas turbine (GT) from GE. A triple-pressure, reheat, horizontal, natural circulation HRSG without supplementary fire is selected as the matching steam turbine cycle (namely bottoming cycle) for the basic system which comprises high-pressure (HP), intermediate-pressure (IP), and low-pressure (LP) steam subsystems. Each subsystem consists of a feed water pump, an economizer (EC), an evaporator (EV), a superheater (SH) and a corresponding steam turbine (ST). In addition, attemperators are installed between superheaters and reheaters (RH) for safe operation. The schematic diagram of the conventional GTCC is shown in Fig. 1. Firstly, inlet air is compressed and then mixed with fuel. Next, they are burned in the combustion chamber. The burned gas with high temperature drives the GT to generate electrical power. Finally, the exhaust gas from the gas turbine outlet is utilized to generate electricity by the ST with the help of HRSG. The main design parameters of the traditional combined cycle are summarized in Table 1. 2.1.2. Unit with HEAT The proposed GTCC with HEAT is studied with the retrofitting
Table 1 Main parameters of the basic combined cycle. Facility (Unit) Compressor Pressure ratio Inlet air mass flow (kgas1) Efficiency (%) Outlet temperature ( C) Gas turbine Combustion pressure loss (%) Combustion heat loss (%) Fuel lower heating value (kJakg1) Inlet temperature ( C) Exhaust temperature ( C) Power output (MW) Topping cycle efficiency (%) HRSG Main steam temperature ( C) Main steam pressure (kPa) Steam turbine power output (MW) Exhaust gas temperature ( C) Bottoming cycle efficiency (%) GTCC Mechanical loss (%) Generator loss (%) Power outputa (MW) Combined cycle efficiencya (%)
Value 15.4 641.97 88.1 391.8 3.5 0.5 48989.915 1327 614.9 262.3 37.5 567.5 9800 147.1 99.7 33.9 1 1 401.25 57.34
a Mechanical loss and generator loss are considered during the combined cycle power output calculation.
layout as presented in Fig. 2. The compressor inlet air is heated by part of the HRSG exhaust gas in the compressor inlet heat exchanger. The rest part of the combined cycle hardware parameters are designed to have similar geometry parameters with the traditional combined cycle. The flow rate of the heat exchanger hot side can be adjusted to modify T1. The HEAT rate ε is defined as the proportion of the HRSG exhaust gas flow into the heat exchanger. The logarithmic heat transfer temperature difference of the compressor inlet heat exchanger is set as more than 25 C in whole load conditions, resulting in the T1 range of 15 Ce65 C at standard environment conditions.
Fig. 1. Schematic of the basic combined cycle.
Y. Yang et al. / Energy 178 (2019) 386e399
389
Fig. 2. Scheme of the compressor inlet air heating process.
2.2. Part-load operation strategy of the combined cycle with HEAT/ IGV For further study the influence of HEAT/IGV on combined cycle and detailed performance comparison, both the traditional unit and the modified unit are proposed with two corresponding regulations. T4 is significant both to the topping and bottoming cycle performances. Thus T4 is suggested to maintain at its limit value, resulting in strategy of IGV-T3-T4 as well as HEAT-T3-T4.
manner is adopted to decrease the load. Similarly, T3 is firstly maintained at 1327 C for the strategy of IGV-T3-T4 with variable IGV. And then T4 is kept at its maximum constantly by adjusting the fuel consumption rate and further decreasing the IGV angle from 71.3 (83.3%) to 49 (60%), in which process the power output of the combined cycle is reduced from 84.6% to 49.1%. Finally, the IGV angle is fixed and the fuel-only load manner is adopted for load decrement. 3. Combined cycle modeling
2.2.1. Strategy with HEAT Two strategies with HEAT are named as HEAT-T3 (use HEAT control to increase T1 and maintain T3 to decrease load þ keep T1 steady and only decrease fuel to decrease load) and HEAT-T3-T4 (use HEAT control to increase T1 and maintain T3 to decrease load þ use HEAT control to increase T1 and maintain T4 to decrease load þ keep T1 steady and only decrease fuel to decrease load) to highlight the characteristics of the method. IGV angle is kept constant during these strategies with HEAT control. In details for the strategy of HEAT-T3, T3 is firstly maintained at the initial temperature (1327 C) with T1 increasing from 15 C to 57 C by the heat exchanger and T4 increasing from 614.9 C to 650.9 C (maximum allowable temperature) at the same time. Simultaneously, HEAT rate ε is changed from 0% to 80%, in which process the combined cycle power output is reduced from 100% to 81%. Then, T1 is fixed and only the fuel consumption rate is regulated to further reduce the cycle load. As for the strategy of HEATT3-T4: Firstly, T4 is increasing from 614.9 C to 650.9 C when T1 increases from 15 C to 57 C. And then T4 remains unchanged by adjusting the fuel consumption rate. Consequently T1 is further increased from 57 C to 65 C, in which process T3 decreases and the power output drops from 81% to 76.1%. Finally, εreaches its limit (100%) and fuel only control is applied for the cycle load regulation. 2.2.2. Strategy with IGV To verify the accuracy of the models and for comparative purposes, the traditional PG9351FA combined cycle with IGV-T3 [21] and IGV-T3-T4 are also simulated by the gas turbine combined cycle models under design/off-design conditions. To be more specific, for the strategy of IGV-T3: The combined cycle is first operated with IGV angle adjustment to keep T3 at 1327 C constantly, leading to T4 increases from 615 C to 651 C. In which process the IGV angle varies from 88 (100%) to 71.3 (83.3%) when the combined cycle load is 84.6%e100%. Once T4 reaches its limit, the IGV angle will maintained constant and the fuel-only load
In this paper, the components had two kinds of operational modes: design and off-design. The gas turbine combined cycle was first simulated with design models in order to size the units of the system. This was the model parameterization step, so that key parameters such as the stage outlet parameters and overall parameters of compressor could been got. Subsequently, the offdesign model of the system was utilized in which the sizes of the process units were assigned to the values determined in the design model. The combined cycle models were modified in order to predict their behaviors at off-design conditions. Excel (for the topping cycle) and the commercial software Aspen Plus V11.1 [22] (for the bottoming cycle and the recuperator) were utilized to connect the components and calculate the design models. The offdesign performance was modeled using software Matlab [2]. The gas turbine combined cycle was simulated according to the mass/energy balance and chemical equilibrium based on the process layout and most important thermodynamic design parameters reported in the previous literature [23]. Each of the component under full load conditions was modeled respectively，and the overall representation of the gas turbine combined cycle was carried out by connecting them appropriately by means of thermodynamic and mechanic links. It was assumed that ambient air was at standard condition (15 C, 1.013 bar and 60% relative humidity). The fuel was assumed to be injected into the combustion chamber with the same temperature of the compressor inlet air. The rotational speed of the gas turbine and steam turbines were stable at 3000 r/min in all conditions. The fuel lower heating value was assumed as 48989.915 kJ/ kg for all systems. The off-design parameters for the units were assumed to be under steady state and obtained by simulation results from design models. The thermo-physical properties of air and flue gas were calculated by the method proposed by Zhang [24]. In this paper, it should be mentioned that the specific heat capacity of air and flue gas were determined by temperature, component, and
390
Y. Yang et al. / Energy 178 (2019) 386e399
pressure. The detailed design/off-design calculation process of the basic combined cycle can be referred to our earlier work [25]. The cycle efficiency and specific work were defined as follows:
hgt ¼
Wgt mf LHV
hst ¼
(4)
1 ¼ dðtan aÞ d f
(2)
Wtotal ¼ Wgt þ Wst
htotal ¼
(3)
The compressor performance under off-design condition was modeled based on the one-dimensional sequential stage-stacking method [28], with parameters obtained from the design model. The main function of IGV was adjusted to a certain limit to generate arbitrary power during the part-load performance state [29]. The variable compressor inlet mass flow affected by the inlet temperature was selected by the matching relationship between the compressor and the gas turbine during off-design condition with modification based on the characteristics curves mentioned in Ref. [30]. The details of the compressor simulation and IGV control were presented in the previous literature [2] with following functions [31]:
(1)
Wst h4g h0g mg
Wtotal mf LHV
3.1. Compressor modeling
j Wgt þ Wst w¼ ma þ mf
(5)
where W is the power output, h is the thermal efficiency, m is the mass flow, LHV is the fuel lower heating value, h is the enthalpy, P is the pressure, T is the temperature, and w is the specific work. Subscript and Superscript: gt means gas turbine cycle, st means steam turbine cycle, total means combined cycle, f means fuel, a means air, g means flue gas, whereas 0 represents the ambient condition，1 represents the compressor inlet, 2 represents the compressor outlet, 3 represents the gas turbine inlet, 4 represents the gas turbine outlet. Furthermore, the exergy analysis was presented based on Ref [26] and Ref [27] in this paper. The physical exergy of each working fluid was obtained by simulating through Aspen Plus based on its temperature, pressure and component. The exergy efficiencies were defined as follows. In this paper, each thermal efficiency is given as efficiency for short, and the exergy efficiency will be directly named if needed.
hcc
¼
E3g
(6)
f
¼ constant
¼
3.2. Combustion chamber modeling The thermodynamic characteristics of the combustion chamber were simulated with focus on the determination of the excess air coefficient and combustion temperature. The pressure loss during combustion was assumed as 3.5% of the combustion chamber inlet pressure under design/off-design conditions as mentioned in our previous work [2]. Fuel was assumed to be heated into the same temperature as the compressor inlet air before combustion. The thermo-physical properties of gas turbine flue gas mainly relied on its composition and temperature. The method proposed by Zhang [24] was used to calculate the thermo-physical properties of flue gas in different states in the system. The thermal balance equation [32] was shown as following:
ð1 þ bLÞ h3g h1g ¼ h2f h1f þ bL h2a h1a þ hcc LHV
Ef
hst ¼
(12)
Wtur Wcom
E4g
htotal ¼
Wst
(11)
where a is the vane outlet absolute flow angle, 4 is the flow coefficient, j is the pressure coefficient.
E2a þ Ef hgt
(10)
(7)
E4g
Where b is the excess air coefficient, L is the theoretical air quantity, h is the enthalpy, h is the efficiency, and LHV is the fuel lower heating value. Subscripts and Superscripts: cc means the combustion chamber, g means gas, f means fuel, a means air, 1 represents the compressor inlet, 2 represents the compressor outlet, and 3 presents the combustion chamber outlet.
(8)
E5g
3.3. Gas turbine modeling
Wtur Wcom þ Wst
(9)
Ef
Where W is the power output or the power consumption, h is the exergy efficiency, E is the exergy of a given working fluid, Subscript and Superscript: gt means gas turbine cycle, st means steam turbine cycle, total means combined cycle, f means fuel, a means air, g means flue gas, whereas 0 represents the ambient condition，1 represents the compressor inlet, 2 represents the compressor outlet, 3 represents the gas turbine inlet, 4 represents the gas turbine outlet, 5 represents the HRSG outlet.
A gas turbine model was considered with three separate turbine stage cooling assumption [2] as shown in Fig. 3, considering inlet pressure losses [33]. Each stage had a pair of nozzle vane and rotor blade rows. The gas turbine exhaust pressure was set as 0.1038 MPa in all conditions for safety operation. The turbine expansion ratios of the first stage was assumed to be constant; however, the turbine expansion ratios of the rest two stages changed based on the characteristics curve shown in Ref. [32] under part load conditions. In the off-deign conditions, the relationship between the turbine inlet pressure, temperature and mass flow were described as the Flügel equation [34].
Y. Yang et al. / Energy 178 (2019) 386e399
mca ¼ mca;d
391
p2* p2* d
!
T 2* d T 2*
!0:5 (14)
where m is the mass flow rate, p is the pressure, and T is the temperature. Subscripts and Superscripts: d means design condition, ca means the cooling air, 2 represents the compressor outlet, and * represents the stagnation condition. The cooling air was accelerated by the main flue gas, thus its velocity and direction were consistent with those of the main flue gas. The pressure drop during coolant mixing was calculated as [33]:
Dp* p*gt;in
Fig. 3. Schematic of the gas turbine cooled air.
pffiffiffi m T ¼ constant AP
(13)
where m is the inlet mass flow rate of gas turbine, A is the turbine inlet area, T is turbine initial temperature and P is the turbine inlet pressure. The blade cooling model was assumed as shown in Fig. 4. The cooling air quantity at each turbine stage at design condition was estimated by Zhang [35]. The cooling air of the nozzle vanes was mixed with the main flue gas at the inlet. The mixed flue gas expanded in both the nozzle vanes and the rotor blade rows. Finally, the cooling air of the corresponding rotor blade rows was mixed with the expanded flue gas at the outlet of this stage. The cooling air quantity of each turbine stage at off-design conditions can be adjusted according to the temperature and pressure of the bleeding stage [36].
mca m ¼ g Ma2g z ¼ ca k; Dp* < 0 mg g mg
where△p is the pressure loss, p is the pressure, m is the mass flow, g is the specific heat capacity, Ma is the Mach number, z is the mixing loss coefficient, and k is the comprehensive parameter which is set as 0.404 in this work [35]. Subscripts: gt is the gas turbine, g is the gas, ca is the cooling air, and in is the inlet. Superscripts: * is the stagnation condition.
3.4. HRSG modeling The triple-pressure bottoming cycle was modeled by modifying the simplified method for single-pressure HRSG proposed by Ganapathy [37] as summarized in previous literature [25]. The energy balance equation and heat transfer equilibrium equation were used for model calculation as following. In details, the evaporators were simulated as shown in Equation (16), while the rest of heat exchangers in HRSG were calculated by Equations 16e18 together.
Q ¼ mg cp tg;in tg;out ¼ ms hs;out hs;in
(16)
Q ¼ UA D T
(17)
DT ¼
Fig. 4. Blade cooling model adopted for the cooled stages.
(15)
tg;out ts;in tg;in ts;out ln tg;out ts;in tg;in ts;out
(18)
where m is the mass flow, Cp is the specific heat capacity at constant pressure, t is the temperature, h is the enthalpy, U is the overall heat transfer coefficient and A is heat transferring area, △T is the log-mean temperature difference, Q is the heat transfer capacity. Subscript: in means the inlet of heat exchanger, out means the outlet, s means the steam/water side, and g means the gas side. The off-design characteristic of the steam turbine can be described by the Flügel formula [38]. The exhaust temperature of HRSG in design condition was maintained at 100 C for both traditional cycle and retrofitted cycle, by adjusting the design thermal parameters of the bottoming cycle system in reasonable ranges. The simulation followed the flow sequence to analyses the parameter characteristics of the combined cycle under full load conditions. The design water/steam/gas side pressure loss of each heater was set between 2% and 5% and corrected under the offdesign condition [39]. It should be mentioned that since the design parameters of HRSG configuration cannot be adjusted under off-design conditions, the exhaust temperature of HRSG changed with its inlet temperature (T4). This value was set as 99.7 C under design condition and calculated by the off-design modeling.
392
Y. Yang et al. / Energy 178 (2019) 386e399
3.5. Recuperator modeling In this study, the entire heat transfer process of the heat exchange at the inlet of compressor was assumed stable under full load. The pressure loss in the heat exchanger was neglected during Section 4.1-4.4 and discussed in range of 0%e5% in Section 4.5. The mass flow rate of the heat exchanger inlet flue gas was adjusted to balance the heat transfer equilibrium equation [37]:
mg cp;g Tg;in Tg;out ¼ ma cp;a Ta;out Ta;in
(19)
where m is the mass flow, Cp is the specific heat at constant pressure, Tis the temperature. Subscripts and Superscripts: g means gas, a means air, and in means the inlet of recuperator, whereas out means the outlet of recuperator. 4. Results and discussions 4.1. Simulation verification Fig. 5. The influence of inlet air temperature on the gas turbine power output.
Table 2 shows several characteristics parameters of five key operating points (A/B/C/D/E operation points in the following figures) in different strategies. The design gas turbine power output and efficiency of PG9351FA unit are 262.3 MW and 37.5% which are similar with the data mentioned in Gas Turbine World 2004e2005 handbook [40]. Fig. 5 compares the simulation result and the data in Ref. [30] and Ref [41] of the gas turbine power output capacity (with T3 at 1327 C constantly) affected by T1 to validate the models. Gas turbine performance highly depends on the compressor inlet temperature [41]. The simulation result indicates that the gas turbine can produce more power with lower compressor inlet temperature. It is calculated that the difference of the gas turbine cycle power output capacity between 5 C and 55 C is 106.7 MW. 4.2. Off-design characteristic analysis of the topping cycle Fig. 6 presents the running lines of the two strategies with IGV control. The characteristic map of the compressor with inlet angle ranging from 88 (100%) to 49 (60%) is simulated by the stagestacking method as mentioned before. Depending on the matching relationship between the compressor and the turbine, the working points can be specified in the models by the Flügel formula. In addition, each characteristic curve in the figure possesses its own IGV opening with constant rotational speed. The running lines of strategies with HEAT method are shown in Fig. 22 for clearly discussion.
Fig. 6. Operation characteristic curve of compressor with IGV control.
Fig. 7 presents the variation of the gas turbine cycle thermal efficiency and T4 with various compressor inlet temperature at full load capacity, at which T3 is maintained at its design value during simulation. The curves indicate that T1 negatively affects the cycle
Table 2 The key parameters of the main operating conditions in the four strategiesa. Description
Design condition
Inflexion of HEAT-T3
Inflexion of HEAT-T3-T4
Inflexion of IGV-T3
Inflexion of IGV-T3-T4
Point T1 ( C) IGV Compressor pressure ratio Inlet air mass flow (kg/s) Gas turbine power output (MW) Gas turbine cycle efficiency (%) T3 ( C) T4 ( C) Main steam mass flow (kg/s) Steam turbine Power output (MW) Steam turbine cycle efficiency (%) Combined cycle power outputa (MW) Combined cycle Efficiencya (%)
A 15 88 (100%) 15.4 642.0 262.29 37.48 1326.8 614.9 80 147.1 33.90 401.25 57.34
B 57 88 (100%) 13.2 549.9 194.73 34.34 1327.1 650.8 72.1 137.0 34.76 325.14 57.33
C 65 88 (100%) 12.7 532.9 179.02 33.46 1313.4 651.3 69.6 132.7 34.77 305.54 57.1
D 15 71.3 (83.3%) 12.8 535.8 212.25 35.24 1327.0 652.0 70.2 134.0 34.62 339.34 56.34
E 15 49 (60%) 8.8 390.0 105.90 27.36 1176.6 651.5 51.7 95.0 34.03 196.9 50.88
a
Mechanical loss and generator loss are considered during the combined cycle power output calculation.
Y. Yang et al. / Energy 178 (2019) 386e399
393
Fig. 9. Variation in turbine inlet temperature with combined cycle power output.
Fig. 7. The effect of inlet air temperature on the gas turbine cycle.
efficiency which can be explained by the fact that PR decreases with the increasing T1, leading to the decrement of turbine expansion ratio and the increment of T4. This conclusion is in agreement with previous statements [42]. Considering HRSG can recover the energy of the turbine exhaust flue gas, the bottoming cycle power ability is gradually improved by the increasing T4. The variation trends of the gas turbine combined cycles with the four operation strategies in terms of T2, T3, compressor PR, T4 and turbine exhaust mass flow are shown in Figs. 8e12, respectively. A phenomenon of “parallel line” can be observed in these figures under relative low load, because the same regulation of “fuel-only control” is employed in this range. As expected, T2 is affected by T1 with a rising trend for combined cycles with HEAT control under decreasing load conditions. T3 is firstly maintained constant with decreasing load for all strategies. When T4 reaches its limit, the combined cycle load with IGV control decreases to 84.6% (point D), while the combined cycle load with HEAT control is declined to 81% (point B). Then T3 is constantly decreased to 1176.6 C for IGV-T3-T4 and to 1313.4 C for HEAT-T3-T4 under cycle load 49.1% (point E)
Fig. 10. Variation in compressor pressure ratio with combined cycle power output.
Fig. 11. Variation in turbine exhaust temperature with combined cycle power output. Fig. 8. Variation in compressor outlet temperature with combined cycle power output.
394
Y. Yang et al. / Energy 178 (2019) 386e399
Fig. 12. Variation in turbine exhaust mass flow with combined cycle power output.
much worse. Particularly, the gap between IGV-T3-T4 and HEAT-T3T4 is firstly enlarged and then disappeared. At combined cycle load of 84.6% (point D), the topping cycle efficiency of IGV-T3/IGV-T3-T4 is 35.24% and that of HEAT-T3/HEAT-T3-T4 is 35.19%. The Inferiority of the performance of IGV-T3-T4/HEAT-T3-T4, comparing with that of IGV-T3/HEAT-T3 at relative low load (45%e70%), is mainly due to the decrement of PR which plays a significant role in gas turbine cycle performance. From a thermodynamic perspective, the results in Fig. 13 indicate that the cycles with IGV control result in relative higher topping cycle efficiency comparing with the cycles with HEAT control with load of 70%e100%. Furthermore, the operation strategy of HEAT-T3 is treated as an advisable choice only next to IGV-T3 to improve the performance of the gas turbine cycle in low load condition (45%e70%). Moreover, the exergy analysis on the topping cycle is presented in Figs. 14 and 15, respectively. The irreversibility loss in the combustion chamber is the largest part among the gas turbine cycle which is mainly affected by T2 [26]. Thanks to the increasing T2 resulting from the HEAT process, the combustion chamber exergy efficiency is improved and the fuel consumption rate is decreased. As a result, the topping cycle exergy efficiency with HEAT strategy is higher than that with IGV control under relative high load conditions (70%e100%).
and 76.1% (point C), respectively. The compressor PR for the strategy of HEAT-T3/IGV-T3 related to the T4 peak value is 13.2 (point B)/12.8 (point D), respectively. The PR of IGV-T3 is performed the steepest slope (A-D) when the cycle load decreases from 100% (point A) to 84.6% (point D), as presented in Fig. 10. This can be explained by the fact that the PR positively affects the temperature difference between T3 and T4. Furthermore, the relationship between the turbine inlet pressure, temperature and mass flow rate is described as the Flügel formula. Under the assumption of constant T3 (A-B&A-D), the curve trends of the mass flow rates would be similar with that of PRs. When the load reduces from 81% to 76.1% (from 84.6% to 49.1%) for cycle with HEAT-T3-T4 (IGV-T3-T4), the PR would continually drops to 12.7 (8.8) but the slope would be more gentle influenced by the drop of T3 (BeC & D-E). Worthy of note is that, under the fuel only control condition, T4 drops constantly and the turbine exhaust flow increases slightly under relative low load for all strategies, as shown in Fig. 12. The variation of the topping cycle efficiency curves are summarized in Fig. 13. It is observed that the curve trends of the four strategies are similar, and IGV-T3 presents the highest topping cycle efficiency, while the performance of HEAT-T3-T4/IGV-T3-T4 are
In this study, the inlet temperature of the HRSG is directly assumed to be T4 and the HRSG exhaust flue gas is connected to the heat exchanger for the systems with HEAT-T3/HEAT-T3-T4. The heat loss is neglected in these processes. The variation curves of the bottoming cycle efficiencies are presented in Fig. 16. For the combined cycle load range of 45%e 100%, the variation intervals of the bottoming cycle efficiencies of unit with IGV-T3 (IGV-T3-T4) and HEAT-T3 (HEAT-T3-T4) are 27.6%e34.6% (31.4%e34.6%) and 26.4%e34.8% (28.3%e34.8%), respectively. It is well known that the bottoming cycle efficiency is highly dependent on T4 under part load conditions. As shown in Fig. 13, IGV-T3-T4 (HEAT-T3-T4) obtains higher T4 at the expense of reducing the topping cycle performance under low load condition, comparing with IGV-T3 (HEAT-T3), which results in higher bottoming cycle performance.
Fig. 13. Variation in topping cycle efficiency with combined cycle power output.
Fig. 14. Variation in combustion chamber exergy efficiency with combined cycle power output.
4.3. Off-design characteristic analysis of the bottoming cycle
Y. Yang et al. / Energy 178 (2019) 386e399
Fig. 15. Variation in topping cycle exergy efficiency with combined cycle power output.
395
Fig. 17. Variation in HRSG exhaust temperature with combined cycle power output.
Fig. 18. Variation in bottoming cycle exergy efficiency with combined cycle power output. Fig. 16. Variation in bottoming cycle efficiency with combined cycle power output.
4.4. Combined cycle design/off-design characteristics With HEAT process, the waste energy in the HRSG exhaust gas has been partly recovered. Additionally, the compressor inlet air is heated to reduce its mass flow rate because the inlet area is constant for a given compressor. The HEAT regulation presents a similar operating characteristics with IGV control but with narrower range of adjustment. The performance would obviously changes when the HEAT adjustment ends and switches to fuel only control. The outlet temperature of the HRSG (T5) is shown in Fig. 17 with range among 91 C-107 C. The exergy efficiencies of the bottoming cycles are presented in Fig. 18. As aformentioned, attemperators are utilized when T4 is higher than the design value for bottoming cycle safety operation. So the bottoming cycle exergy efficiency decreases with increasing T4, and this parameter recovers with decreasing T4 if it is still higher than the design value (615 C). After the operation strategy changes back to the fuel only control and T4 is lower than 615 C, the attemperators are closed and the bottoming cycle performance decreases with the decreasing T4 as presented.
The comparison of the combined cycle efficiencies for the two units with different operation strategies are demonstrated in Fig. 19. The minimum values of the combined cycle thermal efficiencies of systems with IGV-T3/IGV-T3-T4/HEAT-T3/HEAT-T3-T4 are 46.9%/49.7%/49.0%/49.7% within the load range of 45%e100% respectively. Evidently, the combined cycle efficiency curves show that the unit with HEAT-T3-T4 exhibits the best part load performance comparing with other strategies when the combined cycle load is above 48%. The difference between cycle performance of HEAT-T3T4 (IGV-T3-T4) and HEAT-T3 (IGV-T3) is 0.7% pt. (2.5% pt.) under low load. The difference between that of HEAT-T3 and IGV-T3 becomes 1.1% pt. at 85% cycle load and 1.9% at 50% cycle load. This is because that the irreversibility loss during combustion is reduced by the HEAT process and the fuel consumption rate is decreased. While the gap between HEAT-T3-T4 and IGV-T3-T4 is firstly increased to 1.7% pt. at 76.1% combined cycle load (point B) and then gradually diminished at 48% combined cycle load because of the narrow
396
Y. Yang et al. / Energy 178 (2019) 386e399
Fig. 19. Variation in combined cycle efficiency with combined cycle power output.
range of the HEAT process. During HEAT process, T1 increases, leading to the decrement of inlet air density. As the inlet area held constant for a given compressor, the mass flow rate would be expected gradually decrease. The compressor outlet air will receive a relative higher temperature and lower compressor pressure ratio which will decreases the fuel consumption rate with a constant or a suitable T3 for the requirement of the gas turbine. Considering the constant turbine exhaust pressure assumption, the decrement of turbine expand rate leads to the decreasing turbine power output and increasing T4. The strategy with HEAT decreases the irreversibility loss in topping cycle and reduces fuel consumption rate by increasing T2, resulting in a good load regulating method. The IGV operation range of compressor with IGV-T3-T4 is 60%e 100%, related to the combined cycle load range from 49.1% (point E) to100% (point A). But for the combined cycle unit with HEAT-T3-T4 this total load range regulated by the HEAT process could only varied from 100% to 76.1% (point C). The heat transfer temperature difference limits the range of the strategy. Thus fuel only control is applied at 76.1% combined cycle load and the combined cycle efficiency reduces rapidly. And, it should be mentioned that, we just compared the two pairs of operating strategy IGV-T3-T4 (and IGVT3) and HEAT-T3-T4 (and HEAT-T3). If the HEAT method is combined with IGV, there will be a foreseeable best operating strategy, which may be researched in the future. On the other hand, for combined cycle units that always participate in the pitching peak with power output larger than 48%, HEAT-T3-T4 strategy is highly recommended due to its high part load efficiency. The combined cycle exergy efficiency is shown in Fig. 20. The exergy efficiency curves show similar trend as for the combined cycle efficiency. As previously mentioned, the advantage of the HEAT strategy over the IGV method is mainly due to the improvement of the combustion chamber performance. Higher T2 would decreases the exergy loss during combustion and saves the fuel consumption rate with the same performance requirement, resulting in a better combine cycle efficiency and combine cycle exergy efficiency. From the view of the Second Law of Thermodynamic, the higher temperature of the outlet air from the compressor will greatly reduces the exergy destruction during the combustion process, which results in a higher cycle exergy efficiency. In Fig. 21, the specific works of the four strategies are presented. It is shown that the specific works of cycles with HEAT-T3/HEAT-T3-
Fig. 20. Variation in combined cycle exergy efficiency with combined cycle power output.
Fig. 21. Variation in the specific work with combined cycle power output.
T4 is decreased from 611.4 kJ/kg to 579.1 kJ/kg at load range from 100% (point A) to 81% (point B) due to HEAT process. The compressor can absorbs more than half of the total work produced by the gas turbine, and the specific work of the compressor would be gradually enhanced by the increasing T1. It is also well known that relative low operating T3 results in low specific work. For the combined cycle with HEAT-T3-T4, the specific work is decreased into 561.84 kJ/kg at cycle load 76.1% (point C) affected by the increasing T1 and decreasing T3. The specific power output of the cycle with IGV control is kept almost stable in relative high load. As for IGV-T3/IGV-T3-T4 control, the specific power output of the combined cycle slightly increases from 611.41 kJ/kg to 619.17 kJ/kg when the cycle load is above 84.6% (point D). As for the cycle with IGV-T3-T4 at 49.1% (point E)-84.6% (point D) load, the variation interval of the specific work is 494.94 kJ/kg-619.17 kJ/kg resulting from the changing T3. “Four parallel lines” could be observed in the figure because the same rule (fuel only control) is applied under relative low load. To reduce the load of a combined cycle, normally either T3 or the flue gas mass flow of the working medium would be reduced. But as for the unit with HEAT-T3/HEAT-T3-T4, the specific work is reduced
Y. Yang et al. / Energy 178 (2019) 386e399
397
by increasing the compressor consumption specific work, therefore the cycle performance keeps preferably within corresponding load range. 4.5. Parametric analysis of the pressure loss of the compressor inlet heat exchanger After discussing the combined cycle performances of the novel HEAT process, the effect of the pressure loss of the compressor inlet heat exchanger is analyzed in this part. Previous discussion on Figs. 6e21 are based on the simulated results assuming that the pressure loss of the compressor inlet heat exchanger is ignorable. However, this hypothesis is good for discussion but not very close to reality. In this part, the simulation results of a retrofitted combined cycle using a heat exchanger with 2% or 5% inlet pressure loss at each side are simulated and compared as shown in Figs. 22e25. During the simulation process, the pressure drop ratios remain unchanged for all systems under part load conditions. The performance effects of the pressure loss on the compressor running lines with two different strategies (HEAT-T3&HEAT-T3-T4) are illustrated in Fig. 22. The characteristic map of the compressor inlet temperature ranging from 15 C to 65 C with constant rotational speed is simulated by the off-design compressor model [32] with variable inlet air pressure and density. The flow characteristic and the volume flow rate of the compressor are not affected by the
Fig. 23. Variation in turbine expansion ratio with combined cycle power output.
changing inlet pressure, but the compressor inlet mass flow rate decreases resulting from the reducing inlet air density. As can be seen from the figure, the running lines are parallel and the pressure loss increases PR. But the inlet pressure is affected so that the
Fig. 22. Operation characteristic curves of compressor with HEAT control.
398
Y. Yang et al. / Energy 178 (2019) 386e399
Fig. 24. Variation in combined cycle efficiency with combined cycle power output for strategies with HEAT.
Fig. 25. Variation in combined cycle exergy efficiency with combined cycle power output for strategies with HEAT.
curves of the turbine expansion ratio coincide as compared in Fig. 23. The system with larger pressure loss related to higher compressor PR but similar compressor outlet pressure/turbine inlet pressure for a given compressor inlet mass flow condition. In addition, the inflexion of HEAT-T3 (point B) is declined from 57 C to 55 C and then to 49 C caused by the increasing pressure loss from 0% to 2% and then to 5%. This is because that the turbine expansion ratio decreases under the same cycle load with constant T3, so that T4 raises for a given operation condition. But the turbine exhaust temperature limit is steady all the time, thus the inflexion value changes. The pressure loss shows a negative effect on the topping cycle performance and increases T4, which will enhance the bottoming cycle efficiency slightly but still cannot compensated the performance decline. Considering the negative effect of the increasing backpressure on the cycle performance [9], the HRSG exhaust pressure is set as constant by adding a fan behind the heat exchanger hot side. The extra power consumption by the fan is simulated by the commercial software Aspen Plus V 11.1 [22] and deducted from the combined cycle power output. As depicted in Figs. 24 and 25, the combined cycle efficiency and exergy efficiency of cycles with two different strategies (HEAT-T3 & HEAT-T3-T4) and three different compressor inlet pressure losses (0%, 2%, 5%) are compared. The simulation results convey that the compressor inlet heat exchanger pressure loss of 2% could decreases the combined cycle efficiency for about 0.5% pt. and that of 5% decreases the efficiency for about 1.2% pt. This negative affection remains identical when T3 or T4 is stable and alleviates if fuel only control is utilized for cycle load control. It should be mentioned that the total power output ability is also decreased by the inlet pressure loss, which will affects the load
range of a certain combined cycle unit. Each 1% pressure loss decreases about 5.7 MW on the design combined cycle power output ability. As a result, the combined cycle performance prominently affected by the compressor inlet heat exchanger characteristics. The pressure loss shows a negative effects both on the combined cycle power output ability and cycle efficiency. Certainly, the heat exchanger for commercial utilization will not expected to have such large pressure drop equal to or higher than 2% [43], therefore the novel strategy will still expected to perform a satisfactory cycle efficiency with reasonable heat exchanger pressure drop assumption. 5. Conclusion Basis on the design parameters of the given gas turbine combined cycle, a special method named HEAT (heating the compressor inlet air by part of the HRSG exhaust gas) is purposed in this paper. Topping, bottoming, and combined cycle performances of the two pairs of operation strategies (HEAT-T3/IGV-T3 & HEAT-T3-T4/IGVT3-T4) are compared with both energy analysis and exergy analysis to highlight the advantage of the novel method on part load performance improvement. The cycle with HEAT-T3-T4 is suggested for its significant part load performance. A maximum efficiency increment of 1.7% pt. could be received comparing the cycle performance with HEAT-T3-T4 and the cycle with IGV-T3-T4 under part load conditions. Even taking the cycle with IGV-T3 for comparation, a maximum performance improvement of 1.1% pt. could still be obtained for that with HEAT-T3. The simulation result indicates that the combined cycle efficiency of the novel cycle with HEAT-T3-T4 is almost unchanged with the increasing T1 and constant T3, because the specific work is gradually reduced (from
Y. Yang et al. / Energy 178 (2019) 386e399
611.41 kJ/kg to 579.1 kJ/kg) for load adjustment. This specific work decrement is due to the increasing specific work consumed by the compressor rather than lowering T3 which would leads to a remarkably combined cycle efficiency drop. The efficiency improvement by HEAT control is mainly because the waste energy from the HRSG exhaust gas is absorbed by the compressor inlet air to increase the compressor outlet air temperature. Higher temperature of the compressor outlet air can effectively reduces the extra exergy destruction during the combustion process, so that the cycle performance is improved and the fuel consumption rate is saved. For restriction of the temperature difference in the compressor inlet heat exchanger, the adjustment range of the novel regulation is much narrower than that of the IGV control, resulting in the diminishment of the advantage on combined cycle performance when the cycle load is lower than 48%. Moreover, the affection of the pressure loss is discussed within range of 0%e5%. Each 1% pressure loss caused by the compressor inlet heat exchanger is found to decrease 1.4% on the design cycle power output ability, and 0.25% pt. on the combined cycle efficiency. Besides, the HEAT method can be combined with IGV method, which is a foreseeable best operating strategy that can be researched in the future. In all, the novel method HEAT is suggested as an innovative solution for gas turbine combined cycle load management. Acknowledgements This study was supported by National Natural Science Foundation of China of China (Grant No.51436006); National Natural Science Foundation of China of China (Grant No.51306049); Fundamental Research Funds for the Central Universities (No.2017MS15). References [1] Bakhshmand Sina Kazemi, Khoshbakhti Saray Rahim, Bahlouli Keyvan, et al. Exergoeconomic analysis and optimization of a triple-pressure combined cycle plant using evolutionary algorithm. Energy 2015;93:555e67. [2] Zhang Guoqiang, Zheng Jiongzhi, Xie Angjun, et al. Thermodynamic analysis of combined cycle under design/off-design conditions for its efficient design and operation. Energy Convers Manag 2016;126:76e88. [3] Matta RK, Mercer GD, Tuthill RS. Power systems for the 21st century-" H" gas turbine combined-cycles. NY: GE Power systems schenectady; 2000. [4] Haglind F. Variable geometry gas turbines for improving the part-load performance of marine combined cyclesecombined cycle performance. Appl Therm Eng 2011;31(4):467e76. [5] Aguilar Francisco Jimenez-Espadafor, Rodriguez Quintero R, Carvajal Trujillo E, et al. Analysis of regulation methods of a combined heat and power plant based on gas turbines. Energy 2014;72:574e89. [6] Barelli Linda, Ottaviano Andrea. Supercharged gas turbine combined cycle: an improvement in plant flexibility and efficiency. Energy 2015;81:615e26. [7] Zhang Guoqiang, Bai Ziwei, Yang Yongping. Analysis of recuperated combined cycle with small temperature rise under design/off-design conditions. Energy Procedia 2017;105:1507e12. [8] Liu Taohong, Zhang Guoqiang, Li Yongyi, et al. Performance analysis of partially recuperative gas turbine combined cycle under off-design conditions. Energy Convers Manag 2018;162:55e65. [9] Li Yongyi, Zhang Guoqiang, Bai Ziwei, et al. Backpressure adjustable gas turbine combined cycle: a method to improve part-load efficiency. Energy Convers Manag 2018;174:739e54. [10] De Lucia Maurizio, Lanfranchi Carlo, Boggio Vanni. Benefits of compressor inlet air cooling for gas turbine cogeneration plants. In: ASME 1995 International gas turbine and aeroengine congress and exposition. American Society of Mechanical Engineers; 1995. [11] Farzaneh-Gord Mahmood, Deymi-Dashtebayaz Mahdi. Effect of various inlet air cooling methods on gas turbine performance. Energy 2011;36(2): 1196e205. [12] Basrawi Firdaus, Yamada Takanobu, Nakanishi Kimio, et al. Effect of ambient temperature on the performance of micro gas turbine with cogeneration system in cold region. Appl Therm Eng 2011;31(6e7):1058e67. [13] De Sa Ashley, Zubaidy Sarim Al. Gas turbine performance at varying ambient temperature. Appl Therm Eng 2011;31(14e15):2735e9.
399
[14] Lee Jae Hong, Tong Seop Kim, Eui-hwan Kim. Prediction of power generation capacity of a gas turbine combined cycle cogeneration plant. Energy 2017;124:187e97. [15] Gareta R, Romeo LM, Gil A. Economic optimization of gas turbine inlet aircooling systems in combined cycle applications. In: 15th International conference on efficiency, costs, optimization, simulation and environmental impact of energy systems; 2002. [16] Ali Sohani, Farasati Yashar, Sayyaadi Hoseyn. A systematic approach to find the best road map for enhancement of a power plant with dew point inlet air pre-cooling of the air compressor. Energy Convers Manag 2017;150:463e84. [17] Thamir K Ibrahim, Rahman MM, Abdalla Ahmed N. Improvement of gas turbine performance based on inlet air cooling systems: a technical review. Int J Phys Sci 2011;6(4):620e7. [18] Wang FJ, Chiou JS. Integration of steam injection and inlet air cooling for a gas turbine generation system. Energy Convers Manag 2004;45(1):15e26. [19] Zhang Guoqiang, Zheng Jiongzhi, Yang Yongping, et al. A novel LNG cryogenic energy utilization method for inlet air cooling to improve the performance of combined cycle. Appl Energy 2016;179:638e49. [20] 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 Manag 2007;48(4):1055e64. [21] Kim JH, Song TW, Kim TS, et al. Model development and simulation of transient behavior of heavy duty gas turbines. J Eng Gas Turbines Power 2001;123(3):589e94. [22] Liu Zuming, Karimi Iftekhar A. New operating strategy for a combined cycle gas turbine power plant. Energy Convers Manag 2018;171:1675e84. [23] Wang Haozhi. The method of the S109FA unit‘s operation about deeply unloading in the abnormal working conditions Master’degree thesis. Guangzhou (China): South China University of Technology; 2010. [24] Zhang Shizheng. Polynomial expressions of thermodynamic properties of the products of combustion of fuel with air. J Eng Thermophys 1980;1:10e6 [in Chinese]. [25] Zhang Guoqiang, Zheng Jiongzhi, Yang Yongping, et al. Thermodynamic performance simulation and concise formulas for triple-pressure reheat HRSG of gasesteam combined cycle under off-design condition. Energy Convers Manag 2016;122:372e85. [26] Song TW, Sohn JL, Kim JH, et al. Exergy-based performance analysis of the heavy-duty gas turbine in part-load operating conditions[J]. Exergy. Int J 2002;2(2):105e12. [27] Zhu Mingshan. The exergy analysis of the energy system[M]. Tsinghua University; 1985 [in Chinese]. [28] Stone Aubrey. Effects of stage characteristics and matching on axial-flowcompressor performance. ASME; 1957. [29] Mehrpanahi A, Payganeh GH. Multi-objective optimization of IGV position in a heavy-duty gas turbine on part-load performance. Appl Therm Eng 2017;125: 1478e89. [30] Brooks Frank J. GE gas turbine performance characteristics. Schenectady, NY: GE Power Systems; 2000. [31] Song TW, Kim TS, Kim JH, et al. Performance prediction of axial flow compressors using stage characteristics and simultaneous calculation of interstage parameters. Proc IME J Power Energy 2001;215(1):89e98. [32] Jiao Shujian, Sun Shoulin, Zhang Yanchun. Gas turbine and gas-steam combined cycle system. first ed. Beijing: China Electric Power Press; 2007. p. 41e191 [in Chinese]. [33] Jonsson Maria, Bolland Olav, Bücker Dominikus, et al. Gas turbine cooling model for evaluation of novel cycles. Proceedings of ECOS 2005:20e2. [34] Palmer Carl A, Erbes Michael R, Pechtl Peter A. Gatecycle performance analysis of the LM 2500 gas turbine utilizing low heating value fuels, vol. 8. NEW YORK, NY(USA): ASME; 1993. p. 69e76. [35] Zhang Lili. Modeling and performance analysis for coal gas fired gas turbine. Institute of Engineering Thermophysics Chinese Academy of Sciences; 2010 [in Chinese]. [36] Gate Cycle. Gate cycle for windows version 5.51. r. General Electric Company; 2003. [37] Ganapathy V. Simplify heat recovery steam generator evaluation. Hydrocarb Process 1990;69(3). [38] Cui Ning. Study and application on real-time dynamic simulation model for heavy-duty gas turbine combined cycle power unit. 2008. PhD thesis North China Electric Power University, Baoding (China). [in Chinese]. [39] Wu Xiaobing. Parameter optimization and performance analysis on largescale combined-cycle cogeneration systems. Beijing (China): Tsinghua University; 2005. [40] Handbook. Gas turbine world 2004-2005. USA: Gas Turbine World; 2005. [41] Antonio Peretto, Bhargava Rakesh. A parametric evaluation of fogging technology for gas turbine performance enhancement. Bologna, Italy: Alma Mater Studiorum Universita’ Degli Studi Di Bologna; 2000. lez-Díaz Abigail, Alcara z-Caldero n Agustin M, Gonza lez-Díaz Maria [42] Gonza Ortencia, et al. Effect of the ambient conditions on gas turbine combined cycle power plants with post-combustion CO2 capture. Energy 2017;134:221e33. [43] Kim TS, Ro ST. Power augmentation of combined cycle power plants using cold energy of liquefied natural gas. Energy 2000;9(25):841e56.