Research Article Effective Parameters on ...

Hindawi Publishing Corporation Advances in Mechanical Engineering Volume 2014, Article ID 781503, 13 pages http://dx.doi.org/10.1155/2014/781503

Research Article Effective Parameters on Performance of Multipressure Combined Cycle Power Plants Thamir K. Ibrahim1,2 and M. M. Rahman1,3 1

Faculty of Mechanical Engineering, Universiti Malaysia Pahang, 26600 Pekan, Pahang, Malaysia Department of Mechanical Engineering, Faculty of Engineering, University of Tikrit, Tikrit, Iraq 3 Automotive Engineering Centre, Universiti Malaysia Pahang, 26600 Pekan, Pahang, Malaysia 2

Correspondence should be addressed to M. M. Rahman; [email protected] Received 3 August 2013; Accepted 9 March 2014; Published 4 May 2014 Academic Editor: B. V. S. S. S. Prasad Copyright © 2014 T. K. Ibrahim and M. M. Rahman. This is an open access article distributed under the Creative Commons Attribution License, which permits unrestricted use, distribution, and reproduction in any medium, provided the original work is properly cited. A parametric analysis is performed for numerous configurations of a combined cycle gas turbine (CCGT) power plant, including single-pressure, double-pressure, triple-pressure, triple-pressure with reheat, and supplementary triple-pressure with reheat. The compression ratio of the gas turbine and the steam pressure of the steam turbine are taken as design parameters. The thermodynamic model was developed based on an existing MARAFIQ CCGT power plant and performance model code developed using the THERMOFLEX software. The results show that the highest overall power and thermal efficiency occur for the supplementary triple pressure with reheat CCGT configuration. The overall efficiency increases with an increase of the compression ratio to 18– 20, depending on the configuration of the CCGT, and then decreases with any further increase of compression ratio. The triplepressure with reheat CCGT configuration has the highest overall thermal efficiency. The specific fuel consumption decreases with an increase of the compression ratio to 18–20, and the triple-pressure with reheat CCGT has the lowest specific fuel consumption. The simulation model gives good results compared with the MARAFIQ CCGT power plant. Consequently, it can be stated that the compression ratio and steam pressure strongly influence the overall power and thermal efficiency of CCGTs.

1. Introduction The combination of the Brayton cycle for gas turbines (GTs) and the Rankine cycle for steam turbines (STs) is suitable for attaining efficient CCGT power plants. The Brayton cycle generates high temperatures and discards heat at a temperature that can be used easily by the Rankine cycle plant as a source of energy. Air and steam are the most frequently operating fluids for CCGT power plants [1]. Kaushik et al. [2] studied the performance of a CCGT power plant through simulation and modeling. The activities of the GT at part load were studied and discussed. The CCGT plants represent the result of a sensitivity study on ambient temperature and GT performance [3, 4]. The optimum values of entire thermal efficiency and output of power with the values of the decision variable are demonstrated for CCGT plants [5]. The emulator of the CCGT cogeneration plant was created by Khaliq and Kaushik [6]. The simulator is based on

the computational model working on the principle of power plant modeling. This simulator comprises two portions; one of the portions deals with the simulation of the flow of fluid within the power plant, whilst the second simulates the regulating system of the plant [7]. Many investigations have been done on heat recovery steam generator (HRSG) but a comparison of the different types of HRSGs was not found in any search. Srinivas [8] introduced optimization modeling of the double-pressure reheated HRSG to attain the greatest presentation of the CCGT and then compared and validated the model against the available data. In addition, Sarabchi and Polley [9] structured the thermodynamic optimization of a single-pressure united rotation. The functioning features of a triple-pressure reheat HRSG utilizing Gate cycle software were evaluated by Shin et al. [10]. They revealed that different performances of the HRSG are due to multiple configurations of the HRSG with deviations dependent upon the ambient temperature. Bassily [11] set price optimization and

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evaluation as significant methods for enhancing the entire thermal efficiency of a triple-pressure steam-reheat CCGT. Bassily [12] arithmetically optimized and evaluated doubleand triple-pressure CCGT plants. The arithmetical cost optimization, as well as the modeling of the triple-pressure reheat steam commercial CCGT plants, was investigated by Bassily [13]. There is an extraordinary amount of research on optimizing and modeling CCGT plants for a number of configurations of HRSGs [14, 15]. Mohagheghi and Shayegan [16] developed computer code to examine the competence for a variety of types of HRSGs: single-pressure, dual-pressure, dual-pressure with reheat, triple-pressure, and triple-pressure with reheat. From the thermodynamics optimization of the HRSG, they obtained a high rate of generating power in the steam cycle. The modeling of HRSG was formerly based on successive equation solving and chronological methodology, but this was involved time-consuming and complicated methods [17]. Thus, a computer program was created that is able to produce the equations so that the functioning of the HRSG can be estimated [16]. Arrieta and Lora [18] considered a multipleshaft configuration and gathered two Siemens AG 501F GTs, attached to triple-pressure reheated with accompanying steam turbine and supplementary firing HRSGs. With the utilization of Gate cycle software, the thermodynamic simulation obtained satisfactory results. The outcomes revealed that the overall power output and thermal efficiency have an impact of additional firing, and, additionally, they showed variations in the ambient temperature. Considering the operational manner of CCGT power plants, it appears that about one-third of the total power is generated by the ST cycle and two-thirds by the GT cycle. Thus, the orientation value of the power production of a gas turbine is 200 MW when the ST cycle generates about 100 MW [19, 20]. For most installed plants, the outlet temperature of the exhaust gases of the GT is around 850 K, while the highest HRSG pressure is between 120 and 165 bars [21]. Consequently, a comparative study of effective parameters (compression ratio and steam pressure) on the overall power output, steam mass flow rate, overall thermal efficiency, and heat rate requires managing the parameters of the system. Thus, the aim of the present study is to develop a strategy to enhance the overall performance of multiconfigurations of CCGTs, utilizing the effect of the compression ratio and steam pressure and comparing the results with data from real combined cycle plants.

2. Description of the Triple-Pressure Reheat Combined Cycle Generally, heat recovery steam generators have a different number of pressure levels and supplementary firing (duct burner). The HRSGs can be single-pressure combinecycle (SPCC), dual-pressure combine-cycle (DPCC), triplepressure combine-cycle (TPCC), triple-pressure with reheat combine-cycle (TPRCC), and supplementary triple-pressure with reheat combine-cycle (TPRBCC), as shown in Figure 1.

The multiple pressure levels are used to increase the energy recovered from the exhaust gas of the GT and decrease the energy loss from the CCGT. Five different configurations of CCGT systems are chosen in this study. Each of the CCGT configurations has the same GT specification. The GT cycle specifications are similar to a real CCGT power plant, such as the MARAFIQ CCGT power plant in Saudi Arabia. In addition to the three GT and auxiliary systems, the SPCC consists of three single-pressure HRSGs and a steam turbine. The DPCC configuration consists of three dual-pressure HRSGs and a condensing steam turbine. The TPCC configuration consists of three triple-pressure HRSGs and a condensing steam turbine. The TPRCC configuration consists of three triple-pressure reheat HRSGs and a condensing steam turbine. The TPRBCC configuration consists of three supplementary triple-pressure reheat HRSGs and a condensing steam turbine. A triple-pressure reheat HRSG with supplementary firing unit and a condensing steam turbine can be described as a complex triple-pressure CCGT plant. As shown in Figure 2, the configuration of three GTs and three HRSGs connected with one ST was associated with the unit in this model. Energy and mass balances are presented because a more significant and complex configuration is identified with a supplementary triple-pressure reheat CCGT plant. A schematic diagram of the triple-pressure reheat combined cycle with a supplementary firing unit (TPRBCC) power plant in the midst of a simple GT cycle is depicted by Figure 3. This would produce a turbine inlet temperature of 1600 K. The majority of the air at 1 is compressed to a higher pressure at 2. At this point, the air enters the combustion chamber (CC) and undergoes combustion by utilizing additional fuel. This results in combustion gas at 3. Expansion of the gas at 3 subsequently takes place at the chimney or HRSG at 4. Heat is transferred to the steam once the gas at 4 enters the HRSG before exiting the stack temperature at 5. In the HRSG, expansion of the steam at the outlet of the high-pressure superheater at 6 occurs in the high-pressure steam turbine (HPST), resulting in a lower-pressure and temperature at 7. In the reheat section, steam at 7 is reheated to a higher temperature at 8. There, further expansion of the steam in the intermediate-pressure steam turbine (IPST) to a lower-pressure at 10 occurs. Followed by this is the expansion of the superheated steam at the outlet of the intermediate-pressure section of the HRSG at 9, in the IPST to a lower-pressure and temperature at 10, where it enters into the low-pressure steam turbine. At the outlet of the low-pressure section of the HRSG at 11, the superheated steam and steam at 10 undergoes further expansion in the low-pressure steam turbine (LPST) to lower-pressure and temperature at 12. From the low-pressure steam turbine at 2 bar, steam is extracted before being fed to the open feedwater heater (deaerator) at 13. At 12, and the steam with lowpressure and low-temperature undergoes condensation in the condenser to transform it into saturated water at 14. The resulting water at 15 is the outcome of saturated water flowing out of the condenser at 14 and subsequently mixing with steam at 13 inside the deaerating condenser. At 15, the



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Intermediate- High-pressure Reheat (RH) pressure (IP) (HP) Heat transfer

(e)

Figure 1: A typical temperature versus heat transfer diagram of HRSG combined cycle: (a) single-pressure, (b) dual-pressure, (c) triplepressure, (d) triple-pressure reheat and, (e) supplementary triple-pressure reheat.

saturated water exiting the deaerating condenser is pumped to a higher pressure at 𝑇𝑤1LP . Saturated water at 𝑇𝑤2LP is the result of the heating to which water at 𝑇𝑤1LP is exposed in the low-pressure economizer section of the HRSG. The water

then enters the low-pressure steam drum (D1). In the lowpressure superheater of the HRSG, the saturated steam at the outlet of drum D1 at 𝑇𝑠LP is superheated. Steam at 𝑇𝑠𝑠LP is produced as a result. The low-pressure steam turbine becomes


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Figure 2: A schematic diagram of the MARAFIQ CCGT power plant.

the recipient of the steam at 𝑇𝑠𝑠LP where it undergoes expansion until it enters the condenser pressure at 14. At the outlet of drum D1, the saturated water is pumped to the pressure of drum D2 at 𝑇𝑤1IP . In the intermediate-pressure economizer

section of the HRSG, water at 𝑇𝑤1IP undergoes heating until it claims a saturated condition. Before being partly evaporated in the HRSG’s intermediate pressure evaporation section, the saturated water at 𝑇𝑤2IP undergoes heating.



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Figure 3: A schematic diagram of the supplementary firing triple-pressure steam-reheat combined cycle power plant.

At the top of drum D2 at 𝑇𝑠IP , the saturated vapor is superheated to a higher temperature at 𝑇𝑠𝑠IP in the intermediatepressure superheater section of the HRSG. The intermediatepressure steam turbine becomes the recipient of the steam, which experiences expansion until it reaches the condenser at 14. At 𝑇𝑤1HP , the saturated water present at the outlet of drum D2 is pumped to the pressure of drum D3. In the highpressure economizer section of the HRSG, water at 𝑇𝑤1HP undergoes heating to acquire the saturated water condition. In the high-pressure section of the HRSG, the saturated water is heated and partly evaporated. The high-pressure superheater section of the HRSG is the place where the saturated vapor at the top of drum D3 at 𝑇𝑠HP , is superheated to a higher temperature at 𝑇𝑠𝑠HP . The steam expands into the reheat section at 7 after the superheated steam at 𝑇𝑠𝑠HP enters

the HPST. In the reheat section of the HRSG, the steam at 7 is superheated to a higher temperature at 8, which has some impact on the duct burner. Before being pumped to 16, all the steam at 12 will go through the process of condensation in the condenser to water at 14. The temperature transfer diagram for the CCGT power plant with a simple GT cycle is illustrated in Figure 1(e). The temperature-entropy diagram for the CCGT is shown in Figure 4.

3. Thermodynamic Model and Analysis Because of higher overall thermal efficiencies than individual steam or gas turbine cycles, the CCGT plants preferred as an attractive development in power generation [22, 23]. Hence, due to escalating fuel prices and decreasing fossil


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(10) The terminal temperature difference (TTD) in the HP, IP, and LP superheaters (temperature difference between flue gas and superheated steam) are taken as 20∘ C. (11) The degree of superheat (DSH) in the LP and IP superheater (temperature difference between superheated steam and the saturated steam) is taken as 60∘ C. (12) The temperature difference between steam and outlet cooling water in the condenser is taken as 15∘ C.

Tw1LP

(13) Isentropic efficiency of steam turbine is taken as 90%. (14) Pressure drops in the combustion chamber, HRSG, and condenser are neglected. Entropy (kJ/kg·K)

Figure 4: Temperature-entropy diagram for supplementary firing triple-pressure reheat HRSG combined cycle.

(15) Heat losses in the combustion chamber, HRSG, turbines, and condenser are neglected. (16) All the processes are steady state and steady flow. (17) The fuel injected into the combustion chamber was natural gas.

fuel resources, the optimal design of CCGT plants is vital [24]. In order to ascertain the mass flow rate of steam generated at the drums, the thermodynamic properties of each state, the electrical output of the system, and the thermal energy of a process, a thermodynamics analysis is applied to each case study. Assumptions have been made of the processes being steady state and steady flow. Moreover, it is assumed that the impact of potential and kinetic energy on the system is negligible. The variation of enthalpy and temperature for different substances has been accounted for by the formulations of ideal gases. Steam tables have been used for water [25]. During the analysis of the combined cycle, the following assumptions have been made depending on the real data for the MARAFQ CCGT power plant [26, 27]. (1) Atmospheric conditions are taken as temperature 288 K, pressure 1.01325 bar, and relative humidity 60%. (2) Turbine inlet temperature (TIT) in gas turbine cycle is 1600 K. (3) Isentropic efficiencies of compressor and gas turbine are 83.5%. (4) Heat loss from the combustion chamber is considered to be 3% of the lower fuel heating value [28]. Moreover, all other components are considered adiabatic. (5) Maximum temperature of steam cycle is 833 K. (6) The condenser pressure is taken as 0.12 bar.

3.1. Gas Turbine Model. The network of the gas turbine (𝑊Gnet ) is calculated by 𝑊Gnet = 𝐶𝑝𝑔 × TIT × 𝜂𝑡 (1 −

1 (𝛾 −1)/𝛾𝑔 𝑟𝑝 𝑔

) − 𝐶𝑝𝑎

(𝛾 −1)/𝛾𝑎

× 𝑇1 (

𝑟𝑝 𝑎

𝜂𝑚 𝜂𝑐

(1)

),

where 𝐶𝑝𝑔 is the specific heat of flue gas, 𝐶𝑝𝑎 is the specific heat of air, which can be fitted by (2) for the range of 200 K < T < 800 K, and 𝜂𝑚 is the mechanical efficiency of the compressor and turbine [29]: 𝐶𝑝𝑎 = 1.0189 × 103 − 0.13784𝑇𝑎 + 1.9843 × 10−4 𝑇𝑎2 + 4.2399 × 10−7 𝑇𝑎3 − 3.7632 × 10−10 𝑇𝑎4 ,

(2)

where 𝑇𝑎 = (𝑇2 + 𝑇1 )/2 in Kelvin. The specific heat of flue gas is given by [30] 𝐶𝑝𝑔 = 1.8083 − 2.3127 × 10−3 𝑇 + 4.045 × 10−6 𝑇2 − 1.7363 × 10−9 𝑇3 .

(3)

The output power from the turbine (P) is expressed as 𝑃 = 𝑚̇ 4 × 𝑊Gnet ,

(4)

(7) The rotational speed is assumed to be constant for the steam and gas turbine cycle at 3600 rpm.

where 𝑚̇ 4 is the mass flow rate of the exhaust gases through the gas turbine, which is expressed as in

(8) The pressure of high-pressure (HP) steam is taken as 100 bar.

𝑚̇ 4 = 𝑚̇ 𝑎 + 𝑚̇ 𝑓 .

(9) The pinch points in HP, intermediate-pressure (IP), and low-pressure (LP) evaporators (minimum temperature difference between the flue gas and the saturated steam) are taken as 15∘ C.

(5)

The specific fuel consumption (SFC) is determined by [31–33] SFC =


3600𝑓 . 𝑊Gnet

(6)


7

The heat supplied is also expressed as

3.3. Condenser Model. The heat rejected from the condenser is expressed by

𝑄add = 𝐶𝑝𝑔𝑚 × [TIT − 𝑇1 × (1 + [

(𝛾 −1)/𝛾𝑎

𝑟𝑝 𝑎

𝜂𝑐

−1

)] . ]

(7)

The gas turbine efficiency (𝜂th ) can be determined by [29–34] 𝜂th =

𝑊Gnet . 𝑄add

(8)

𝑄cond = 𝑚̇ 𝑤 (ℎ12 − ℎ14 ) .

3.4. Pump Model. The condensate water from the condenser is extracted by the pump and is raised to the economizer pressure. The corresponding work for 3 levels is given by 𝑊𝑝 = 𝑚̇ 𝑤 × V𝑓16 (𝑝𝑠ℎLP − 𝑝𝑐 ) + (𝑚̇ 𝑤 − 𝑚̇ 𝑠LP )

3.2. Steam Turbine Cycle Model. It is assumed that the ST efficiency and the pump efficiency are represented by 𝜂𝑠𝑡 and 𝜂𝑝 , respectively. The ideal and actual processes are represented on the temperature-entropy diagram by solid and dashed lines, respectively, as shown in Figure 4 [29]. Duct Burner. The duct burner is used to burn additional fuel in the supplementary firing, which leads to an increase in the temperature of the exhaust gas that passes through the HRSG. In a duct burner, 𝑚̇ 4 × 𝐶𝑝𝑔4 × 𝑇4 + 𝑚̇ 𝑓𝑑𝑏 × LHV = (𝑚̇ 4 + 𝑚̇ 𝑓𝑑𝑏 ) × 𝐶𝑝𝑔4𝑏 × 𝑇4𝑏 + (1 − 𝜂𝑑𝑏 ) × 𝑚̇ 𝑓𝑑𝑏 × LHV, (9)

× V𝑓17 (𝑝𝑠ℎIP − 𝑝𝑠ℎLP ) + (𝑚̇ 𝑤 − 𝑚̇ 𝑠IP − 𝑚̇ 𝑠LP )

Heat Recovery Steam Generator Model. The analysis of the HRSG unit is based on the pinch and approach point. From Figure 1(e), it is clear that the energy balance for the highpressure steam evaporator side of the HRSG gives [35, 36]

𝑇𝑤2HP = 𝑇𝑠HP − 𝑇𝑎𝑝 .

Therefore, the network for the steam turbine power plant is expressed as 𝑊snet = 𝑊st − 𝑊𝑝 .

𝐶𝑝𝑔10 × 𝑇𝑔10 𝐶𝑝𝑔11

−

𝑚̇ 𝑠LP × (ℎ𝑠LP − ℎ𝑤2LP ) . 𝑚̇ 𝑔 × 𝐶𝑝𝑔11 × ℎ1𝑓

The performance of a combined cycle gas turbine power plant, including the thermal efficiencies for the gas turbine cycle, steam turbine cycle, and overall efficiency, is calculated by the following equations [21]. The efficiency for the steam turbine power plant is defined as

(11)

(12)

Steam Turbine Model. By performing the energy balance for a steam turbine, as shown in Figure 4, the following relation is obtained [36]: 𝑊ST = 𝑚̇ 𝑠HP × ℎ6 − 𝑚̇ 𝑠RH × ℎ7 + 𝑚̇ 𝑠RH × ℎ8 + 𝑚̇ 𝑠IP × ℎ9 + 𝑚̇ 𝑠LP × ℎ7 − 𝑚̇ 𝑤 × ℎ12 ,

(13)

where 𝑚̇ 𝑤 is the water mass flow rate and is determined by 𝑚̇ 𝑤 = 𝑚̇ 𝑠HP + 𝑚̇ 𝑠IP + 𝑚̇ 𝑠LP .

𝜂st =

𝑊snet . 𝑄av

(18)

The overall thermal efficiency of the combined cycle gas turbine power plant is expressed as [24, 28, 37–42] 𝜂all =

3 × 𝑊Gnet + 𝑊snet . 3 × 𝑄add

(19)

4. Results and Discussion

The heat available with exhaust gases from gas turbine is given by 𝑄av = 𝑚𝑔 × (𝐶𝑝𝑔1 𝑇𝑔1 − 𝐶𝑝𝑔11 𝑇𝑔11 ) × ℎ1𝑓 .

(17)

(10)

The temperature of the exhaust gases at the exit from the lowpressure economizer of the HRSG could be found by considering the energy balance of the low-pressure economizer (Figure 1(e)): 𝑇𝑔11 =

(16)

× V𝑓18 (𝑝𝑠ℎHP − 𝑝𝑠ℎIP ) .

where 𝜂𝑑𝑏 is the duct burner efficiency and is taken as 93% [17], 𝑚̇ 𝑔 = 𝑚̇ 4 + 𝑚̇ 𝑓𝑑𝑏 , and 𝑇𝑔1 = 𝑇4𝑏 ; the result from (9) is the 𝑇𝑔1 .

𝑇𝑔4 = 𝑇𝑠HP + 𝑇𝑝𝑝 ,

(15)

(14)

The effect of the compression ratio and steam pressure on the overall thermal efficiency and power output in the CCGT is demonstrated and validated against the real power plant of MARAFIQ in Saudi Arabia. The MARAFIQ CCGT power plant comprises three GT units and three HRSG units connected by a steam turbine unit. The effect of the steam cycle high-pressure on the overall thermal efficiency of the CCGT of the different configurations of the HRSG is shown in Figure 5. It was observed that the overall efficiency of the configuration increased with increasing steam pressure of the HRSG. All the configurations had the overall efficiency enhanced. With the exception of the SPCC configuration for HRSG, the overall efficiency of the cycle increased for the DPCC, TPCC, TPRCC, and TPRBCC, which makes them suitable for power plants. The comparison between the simulated overall power outputs of the CCGT configurations versus the real results from the MARAFIQ CCGT power plant is shown in Figure 6. It is also evident that the increase in the steam pressure of the HRSG leads to an increase in the overall power output of the cycle. In addition, compared with other configurations, the TPRBCC configuration has


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Figure 5: Effect of steam pressure of HRSG on overall thermal efficiency of the CCGT power plants configurations.

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TPRCC TPRBCC

Figure 7: Effect of the compression ratio on exhaust temperature of the HRSG configurations.

the highest power. In comparison with the MARAFIQ CCGT power plant, the power output from the TPRBCC power plant is much higher. This is due to the assumption that the simulation model is based on complete combustion [21]. Furthermore, the operating method and type of fuel used contribute to increased losses for the real power plant. The effect of compression ratio on the exhaust temperature of the HRSG configurations is shown in Figure 7. It is evident that, with an increase in the compression ratio for all the configurations of CCGT, there is an increase in the compression ratio which may also cause an increase in the exhaust gas temperature of the HRSG [14]. There is a much lower temperature of the exhaust gases from the HRSG in the TPRBCC configuration, compared with the higher exhaust gas temperature in the SPCC configuration. With the SPCC configuration, less energy is recovered from the exhaust gases that have passed through the HRSG. Figure 8 shows the effect of the compression ratio on the steam mass flow rate of the CCGT power plant configurations. When the compression ratio increases from 10 to 30, there is a decrease in the steam generated in the ST cycle by about 140 kg/s. The reduction in the exhaust gas temperature after the increase in the compression ratio is the main reason behind this. The number of pinch points determines the mass generated steam, which are one for SPCC, two for DPCC, and three for TPCC, TPRCC, and TPRBCC. Therefore, the amount of steam generated increases with the number of pressure levels. Bassily [12] discussed similar trends in the results. In every case, other factors, such as pressure, determined the effect of the mass generated steam.


Advances in Mechanical Engineering 350

Steam mass flow rate (kg/s)

300

250

200

150

8

12

16 20 24 Compression ratio

28

32

TPRCC TPRBCC

SPCC DPCC TPCC

Figure 8: Effect of the compression ratio on the steam mass flow rate of the CCGT power plants configurations.

500

Power output (MW)

The phasing out of steam at the various configurations, temperatures, and pressures determines the steam generated [13]. The comparison between the simulated power outputs of the GT and ST configurations versus the real power plants of MARAFIQ CCGT in Saudi Arabia is shown in Figure 9. It was observed that, when the compression ratio reached the value of 12.6, the GT power output decreased with further increases in the compression ratio. The increase in the compression ratio led to a decrease in the steam turbine power output. This is because of the opposite effect of the compression ratio on the steam generated in the ST power plant [1, 17, 19, 40]. The increase in the compression ratio leads to a decrease in the steam generated in the HRSG (Figure 8). This is due to reduced temperature of exhaust gases from the GT and the increased temperature of the exhaust gases from the HRSG. The amount of energy recovered is reduced from the exhaust gases in the HRSG [28, 37]. The simulation results are in agreement with the real power plants of MARAFIQ CCGT. The effect of the compression ratio on performance of the CCGT power plants for different configurations of the HRSG is shown in Figure 10. The effect of compression ratio on the overall thermal efficiency of the CCGT plants configurations is shown in Figure 10(a). There is an increase in the overall thermal efficiency with an increase in the compression ratio. However, the thermal efficiency starts to decrease with increasing compression ratio after the compression ratio has reached a value of 18–20. This is due to the reduction in the power of the GT cycle with increasing compression ratio [42]. In addition, the increase in the compression ratio results in the decrease of the steam generated in the HRSG. Secondly, a lower overall thermal efficiency is obtained in the SPCC, whereas a higher overall efficiency is obtained in the TPRCC configuration. Figure 10(b) shows the effect of the compression ratio on the heat rate of the CCGT power plant configurations. It was also observed that increasing compression ratio was matched with a decreasing heat rate. However, after the compression ratio reached 18–20, an increase in the heat rate was observed with further increases in the compression ratio. The effect of the compression ratio on the specific fuel consumption of the CCGT power plant configurations is shown in Figure 10(c). It is evident that higher specific fuel consumption occurred with the SPCC configuration, whereas lower specific fuel consumption occurred with the TPRCC at a compression ratio of 18. The simulated overall power output of the CCGT configurations was compared with real results from the MARAFIQ CCGT in Saudi Arabia, together with the effect of compression ratio and presented in Figure 11. The increase in the compression ratio of the GT cycle for all configurations resulted in the reduction of the overall power output. Highest overall output was obtained with the TPRBCC, and lowest overall output was obtained with the SPCC. The TPRBCC configuration power plant has a much higher overall power output compared with the MARAFIQ CCGT, because the real CCGT plant has many losses due to leakages, incomplete combustion, and control problems.

9

400

300

200

8

12


Gas turbine SPCC DPCC TPCC TPRCC TPRBCC MARAFIQ CCGT

28

32

Steam turbine SPCC DPCC TPCC TPRCC TPRBCC MARAFIQ CCGT

Figure 9: Comparison between simulated power outputs of the GT and ST configuration versus practical results from MARAFIQ CCGT power plant with effect of the compression ratio.


10


Overall thermal efficiency

0.53

0.52

0.51

0.50

8

12


28

32

TPRCC TPRBCC

SPCC DPCC TPCC

(a) Overall thermal efficiency

0.049 Specific fuel consumption (kg/kW·h)

Heat rate (kJ/kW·h)

7200

7100

7000

6900

6800

8

12


28

32

TPRCC TPRBCC

SPCC DPCC TPCC

(b) Heat rate

0.048

0.047

0.046

8

12


SPCC DPCC TPCC

28

32

TPRCC TPRBCC

(c) Specific fuel consumption

Figure 10: Effect of the compression ratio on performance of different configurations of the CCGT power plants.

4.1. Uncertainty Evaluation. According to Holman [43], experimental errors found in the MARAFQ CCGTs output power have been dealt with by carrying out a comprehensive error investigation. Table 1 shown below lists down

error readings pertaining to both the CCGT power plants recorded from various instruments, whereas Table 1 lists down the readings having the maximum possible error in them.



11 Table 1: Uncertainties of instruments and properties.

Name of instrument

Range of instrument

Variable measured

1

Thermocouple

0–120∘ C

2

Pressure gage

0–20 bar

3

Pressure gage

0–140 bar

Ambient temperature Compressed air pressure Superheated steam pressure

Item number

Least division in measuring instrument

Min. and max. values measured in experiment

Uncertainty error (%)

0.2∘ C

14.44–44.9∘ C

0.0991

0.1 bar

14.6–15.96 bar

0.6544

0.2 bar

88.1–121.2 bar

0.2024

5. Conclusions

Overall power output (MW)

900

800

700

600

8

12


28

32

TPRCC TPRBCC MARAFIQ CCGT

SPCC DPCC TPCC

Figure 11: Comparison between simulated overall power outputs versus practical results from MARAFIQ CCGT power plant with effect of the compression ratio.

Equation (20) represents the assessment of uncertainty of the CCGT: 𝑈𝑃CCGT 𝑃CCGT = √(

2

2

𝛿𝑃 2 𝛿𝑃 𝛿𝑃 ) ∗ 𝑤𝑇21 + ( ) ∗ 𝑤𝑟2𝑝 + ( ) ∗ 𝑤𝑆2ℎ𝑝 , 𝛿𝑇1 𝛿𝑟𝑝 𝛿𝑆ℎ𝑝 (20)

where 𝑤𝑆ℎ𝑝 = Least division of the pressure of the superheat steam; 𝑤𝑟𝑝 = Least division compression ratio; 𝑤𝑇1 = Least division temperature.

The performance of CCGT power plants is checked by parametric analysis. A new methodology of thermodynamic model was launched to check the performance of a working CCGT plant, called the MARAFIQ CCGT plant. Different parameters are used for different purposes, but steam pressure is the major parameter used to get the maximum energy from the exhaust. To obtain maximum thermal efficiency as well as power generation, multiconfigured parameters are designed. Parametrical analysis is done to learn about the performance of CCGT plants, using the parameters of compression ratio and steam pressure. It is observed that, as the steam pressure of the HRSG is increased, the overall thermal efficiency, as well as power generation, of the CCGT plant is increased for the SPCC, TPCC, DPCC, TPRCC, and TPRBCC configurations. The thermal efficiency was increased with an increase of the compression ratio until 19; after that, the overall thermal efficiency decreases. The performance of the MARAFIQ CCGT plant, together with a running performance for CCGT power plants, is checked by the parameters discussed above. It is concluded that the compression ratio has the most effect on the CCGT configuration, overall thermal efficiency, and power output. Furthermore, the TPRCC configuration had the highest thermal efficiency, whereas the TPRBCC configuration had the highest overall power output.

Nomenclature 𝐶𝑝𝑎 : 𝐶𝑝𝑓 : 𝐶𝑝𝑔 : 𝑓: ℎ: ℎ1𝑓 : 𝑚̇ 𝑎 : 𝑚̇ 𝑓 : 𝑚̇ 𝑔 : 𝑚̇ 𝑤 : p:

The specific heat of the air (kJ/kg⋅K) The specific heat of the fuel (kJ/kg⋅K) The specific heat of flue gas (kJ/kg⋅K) The fuel-air ratio Enthalpy (kJ/kg) The heat loss factor in the heat recovery steam generator The air mass flow rate (kg/s) The fuel mass flow rate (kg/s) The mass flow rate of the exhaust gases through the gas turbine (kg/s) The water mass flow rate (kg/s) Pressure (bar)


12


𝑄add : 𝑄av :

The heat supplied (kJ/kg) The heat available with exhaust gases from gas turbine cycle (kJ/kg) The heat rejected from the condenser 𝑄cond : (kJ/kg) 𝑃: The net power output of the turbine (MW) Compression ratio 𝑟𝑝 : 𝑇: Temperature (K) 𝑇1 : Compressor inlet air temperature (K) The average temperature (K) 𝑇𝑎 : 𝑇s : The saturation steam temperature (K) 𝑇𝑤1 : The temperature of water entering the economizer (K) The temperature of water entering the 𝑇𝑤2 : evaporator (K) V𝑓 : Specific volume of the water (m3 /kg) 𝑊Gnet : The network of the gas turbine (kJ/kg) 𝑊𝑝 : The work of the pump (kJ/kg) 𝑊snet : The work net of the steam turbine cycle (kJ/kg) The work of the steam turbine (kJ/kg) 𝑊st : 𝑈𝑃CCGT /𝑃CCGT : Uncertainty of the CCGT 𝑤𝑆ℎ𝑝 : Least division of the pressure of the superheat steam Least division compression ratio 𝑤𝑟𝑝 : 𝑤𝑇1 : Least division temperature. Greek Symbols 𝜂𝑡 : 𝛾𝑎 : 𝛾𝑔 : 𝜂𝐶: 𝜂𝑑𝑏 : 𝜂𝑚 :

Turbine efficiency Specific heat ratio of air Specific heat ratio of gases Isentropic compressor efficiency The supplementary firing efficiency The mechanical efficiency of the compressor and turbine 𝜂𝑝 : The water pump efficiency 𝜂st : The steam turbine efficiency 𝜂th : The thermal efficiency of the gas turbine.

Abbreviation CCGT: DPCC: GT: HP: HPST: HRSG: IP: IPST: LP:

Combined cycle gas turbine Dual-pressure combined cycle Gas turbine High-pressure High-pressure steam turbine Heat recovery steam generator Intermediate pressure Intermediate-pressure steam turbine Low-pressure

LPST: SPCC: ST: TIT: TPCC: TPRBCC: TPRCC:

Low-pressure steam turbine Single-pressure combined cycle Steam turbine Turbine inlet temperature Triple-pressure combined cycle Supplementary triple-pressure with reheat combined cycle Triple-pressure with reheat combined cycle.

Conflict of Interests The authors declare that there is no conflict of interests regarding the publication of this paper.

Acknowledgments The authors would like to thank University Malaysia Pahang for providing laboratory facilities and financial support under Doctoral Scholarship scheme (no. GRS100332).

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