Exergy-energy analysis of full repowering of a steam power plant

Abstract

A 320 MW old steam power plant has been chosen for repowering in this paper. Considering the technical conditions and working life of the power plant, the full repowering method has been selected from different repowering methods. The power plant repowering has been analyzed for three different feed water flow rates: a flow rate equal to the flow rate at the condenser exit in the original plant when it works at nominal load, a flow rate at maximum load, and a flow rate when all the extractions are blocked. For each flow rates, two types of gas turbines have been examined: V94.2 and V94.3A. The effect of a duct burner has then been investigated in each of the above six cases. Steam is produced by a double-pressure heat recovery steam generator (HRSG) with reheat which obtains its required heat from the exhaust gases coming from the gas turbines. The results obtained from modeling and analyzing the energy-exergy of the original steam power plant and the repowered power plant indicate that the maximum efficiency of the repowered power plant is 52.04%. This maximum efficiency occurs when utilizing two V94.3A gas turbines without duct burner in the steam flow rate of the nominal load.

Graphical abstract

Keywords

full repowering / exergy analysis / V94.2 and V94.3A gas turbines / double-pressure HRSG / duct burner / Bandarabbas steam power plant / efficiency

1 Introduction

World development in the recent decades has been significantly depended on electrical energy. Approximately 1/3 of energy resources is consumed for producing electric energy [1]. Reduction of fossil fuels as the largest resource of energy and increase of environmental pollution have forced human beings to utilize new methods for optimal use of energy. Since removing fossil fuels from the cycle of energy is almost impossible, new methods have to be found to optimize the use of these fuels. Repowering of the old steam power plants is an appropriate method to reduce thermal wastes and increase the efficiency in power plants. It decreases the expenses of energy production and minimizes environmental pollutions. It also increases the capacity of power plants to a great extent. Therefore, electricity generating companies are paying more attention to this issue [1−3].

Reduction of environmental effects and investment expenses are the most important advantages of repowering a steam power plant [4,5]. Meanwhile, this technique has some restrictions [1]. First, repowered power plant is more complicated than the primary unit and it has a lower potential reliability. Next, due to the use of old units with working life of over 20 years, availability (net annual working hour) of the repowered power plant is decreased. Finally, the repowered power plant may have lower performance flexibility.

Repowering of a steam power plant may increase the net output power by 200% and may enhance the efficiency of the power plant by 30% [6]. In addition, repowering reduces heat rate by 30%−40% and significantly decreases production of NOx [7].

Generally, repowering is categorized into partial repowering and full repowering. Partial repowering is utilized to enhance the function of newer steam power plants. It enhances the efficiency of the power plant after some modifications [8,9]. The most popular methods of partial repowering are the feed water heating method, hot wind box method, and supplemental boiler method [5,10].

Full repowering which is the most common method is mostly applied to repowered steam power plants with a life of over 25 years [5,8,11]. Full repowering includes mounting a number of gas turbines to the heat recovery steam generator (HRSG) and deletion of the main boiler of the power plant in a way that the HRSG is used to feed the steam turbines of the main power plant [3,5−7]. It is the best way of maximizing the efficiency of a power plant [3,6]. This converts the power plant to a combined cycle. Full repowering of a steam power plant requires less investment than constructing a new combined cycle power plant [3,7].

Despite the importance of repowering of steam power plants, there are few published researches. Brander and Chase explained the general techniques for conversion of steam cycles to combined cycles by full recovery and described a few practical cases [2]. Heyena and Kalitventze technically evaluated the integration of gas turbines into steam cycles by one of the partial repowering methods and compared it with the combined cycle [12]. They found out that the partial repowering methods could improve the quality of power production and reduce emissions in steam power plants. Escosa and Romeo studied the effect of partial repowering by both the feed water heating and supplemental boiler methods on exhaust emission [13]. Evaluation of the repowering methods and application of these methods in different regions of the world are also mentioned. Cardu studied the benefits of improvement and reconstruction of the Romanian power plants using various methods including the repowering methods [14]. In Italy, Melli al. examined the improvement of steam cycles using different gas turbine combination methods in power plants with less than 180 MW power in that country and reported the results of application of these methods in some plants [1].

The result of repowering is a combined cycle. The technical quality of these systems is measured by exergy and thermal efficiencies. There are many published papers in which the combined cycles are techno-economically analyzed or optimized. Some of these studies examine the effect of each of the main components of the cycle on the economic and technical characteristics of the system. Bassily studied the influence of parameters on the quality of power generation in a three-pressure combined plant based on the cycle efficiency [15]. Srinivas et al. also analyzed the sensitivity of a two-pressure combined cycle based on the exergy efficiency [16]. Franco studied the reciprocal thermodynamic effects of gas turbine and heat recovery boiler in a small scale combined cycle [17]. Variation of combined cycles is not limited to the pressure levels of heat recovery boilers and the structural properties can be affected by the type of turbine(s) used in the cycle. To study the effects of the operational parameters on a combined cycle with gas unit cooling, Sanjay performed a sensitivity analysis of changes in these parameters on the cycle exergy efficiency [18].

The works presented on repowering are almost focused on the effects of different methods of full and partial repowering on target functions. The present study analyzes the full repowering of an old steam plant, which is recommended as the best repowering method for these plants. In a further step, the effect of two common types of gas turbines in repowering is examined. Then, due to the importance of productivity and the direct relationship between power production and generated steam, the performance of the repowering plant has been studied in three different feed water flow rates: a flow rate of feed water which equals the condenser outlet flow of the original power plant when it works at nominal load, a flow rate of feed water at maximum load, and a flow rate of feed water when all extractions are blocked.

Steam power plants play a key role in Iran. Approximately 30% of the required electricity in the country is generated by these power plants [19]. Bandarabbas steam power plant in southern Iran is an old working steam unit. The efficiency of the power plant has been reduced by 16.7% due to its long working life, technical conditions, and the humid corrosive atmosphere. This fact alone is a convincing reason for its repowering. This paper studies the repowering of Bandarabbas steam power plant with two different gas turbines (V94.2 & V94.3A), compares the modeling results, and chooses the best method for the repowering of this power plant.

2 Specifications of Bandarabbas power plant

Bandarabbas steam power plant with a working life of over 28 years is located at 12 km west of the city of Bandarabbas near the seashore. Consisting of four units of 320 MW with a nominal efficiency of 38.7%, it provides a large part of the required electricity of southern Iran and approximately 1.8% of the total electricity of the whole country. Sea water temperature used by the power plant is 27°C (Fig. 1).

3 Repowering of steam power plant

The long working life and the high importance of Bandarabbas steam power plant are the factors that justify the investment on its repowering. Since this power plant is very important for the country’s electricity network, it cannot be kept offline for a long time, thus the time for its repowering is limited. This is another reason for choosing full repowering of this power plant. In full repowering, the boiler is completely deleted and a HRSG is utilized instead. Preheating of the feed water is also conducted by HRSG. Hence, feed water heaters would be omitted. As the result of this deletion, no steam turbine extraction is required. Extractions would be closed and steam flow becomes a fixed value along all sections of the steam turbine. The heat required by HRSG is supplied by the exhaust gases of one or a group of gas turbines. The result of this combination is a repowered power plant which has similar functions to a combined cycle (Fig. 2).

3.1 HRSG

The HRSG is made in Iran by Mapna Company according to the structure and specifications of Bandarabbas steam power plant. It is a reheating double-pressure HRSG. This boiler consists of three sections with different pressures: preheating and deaerating, low pressure (LP), and high pressure (HP). The preheating section heats the feed water until it gets saturated at the output pressure of the condenser pump. This section is located at the end of the HRSG and uses the remaining heat of the gases in that location. The oxygen in the feed water causes two problems: it causes corrosion of the pumps and the pipes, and along with the CO₂, it acidizes the water which causes corrosion of all metal parts. Hence, the deaerator installed in this section separates the air from the feed water [20]. The LP and HP sections of the HRSG provide the dry steam with the pressure required by the LP and HP sections of the turbine. There is another reheater in the HP section which receives the exiting steam from the HP section of the turbine and sends it to the turbine intermediate pressure (IP) section after reheating. A duct burner is located at the inlet of the HRSG to provide the extra heat required in different conditions.

3.2 Gas turbine

Approximately 95% of the power plants in Iran use V94.2 gas turbines and the rest utilize V94.3A gas turbines. Both of these turbines are made in Iran. Hence, the repowering has been conducted based on these turbines. Natural gas is the primary fuel for these turbines. Table 1 shows the analysis of this gas [20,21].

4 Modeling the original steam power plant

Repowering of a power plant is affected by the thermodynamic properties of the original power plant. As a result, the thermodynamic properties at all parts of the original plant should be defined first. The original steam power plant is modeled and its properties at all points are calculated and applied to the repowered cycle as input data. The most important condition is that, at the steam turbine entrance, the state of steam should be exactly the same, before and after repowering. Table 2 presents the data obtained from modeling the steam power plant by EES software at nominal load. For the dead state, a temperature of 27°C and a pressure of 1 atm are used, with consideration of Bandarabbas climate and elevation.

Table 3 is a comparison of the modeling results with the steam power plant data, which shows that the computed results are in good agreement with the data of the power plant.

5 Modeling the repowered cycle

The repowered cycle has been modeled point to point by EES software. The natural gas as described in Table 1 is taken as the fuel for the gas turbine and duct burner. Since the best efficiency of a gas turbine happens at its design condition, meeting the design condition for the HRSG outlet steam is the basis of modeling the repowered cycle. In other words, the HRSG should be designed in such a way that its dry steam output provides the same temperature, enthalpy and pressure as the steam turbine design condition.

5.1 Steam turbine

All the steam turbine extractions would be closed in the full repowering method. Hence, specifications of the fluid passing through the steam turbine blades would differ from the original condition. This is due to the change of the steam flow rate at different stages. This change would make the steam turbine work under its design condition. In this case, the Stodola empirical equation is used to calculate the ratio of the steam flow rate after closure of the extractions to the steam flow rate of the original plant [22].

(1){\displaystyle \frac{{\dot{m}}_{real}}{{\dot{m}}_{first}}}={\displaystyle \frac{p_{out,real}}{p_{out,first}}}.\sqrt{{\displaystyle \frac{T_{in,first}}{T_{in,real}}}}.\sqrt{{\displaystyle \frac{\mathrm{1}-{(p_{in,real}/p_{out,real})}^{\mathrm{2}}}{\mathrm{1}-{(p_{in,first}/p_{out,first})}^{\mathrm{2}}}}}.

Using the above mentioned ratio, the isentropic efficiency of the steam turbine after changing the mass flow rate is calculated using Eq. (2) [22].

(2){\displaystyle \frac{{\eta }_{ise,st,real}}{{\eta }_{ise,st,first}}}=-\mathrm{1.0176}\times {\left({\displaystyle \frac{{\dot{m}}_{real}}{{\dot{m}}_{fiest}}}\right)}^{\mathrm{4}}+\mathrm{2.4443}\times {\left({\displaystyle \frac{{\dot{m}}_{real}}{{\dot{m}}_{fiest}}}\right)}^{\mathrm{3}}-\mathrm{2.1812}\times {\left({\displaystyle \frac{{\dot{m}}_{real}}{{\dot{m}}_{fiest}}}\right)}^{\mathrm{2}}+\mathrm{1.0535}\times \left({\displaystyle \frac{{\dot{m}}_{real}}{{\dot{m}}_{fiest}}}\right)+\mathrm{0.701.}

The isentropic efficiency is defined by

(3){\eta }_{ise,st,real}={\displaystyle \frac{h_{in,st}-h_{out,st}}{h_{in,st}-h_{out,st,ise}}}.

The power produced by the steam turbine is calculated by

(4){\dot{W}}_{st}={\dot{m}}_{st}(h_{in,st}-h_{out,st}).

5.2 Steam turbine

This section is based on the heat exchange between two fluids: the water coming from the condenser and the exhaust gases of the gas turbine(s) combined with the gases produced in the duct burner.

The thermodynamic properties at different sections of a double pressure reheating HRSG is obtained by

(5){\dot{m}}_{water/steam}(h_{out,water/steam}-h_{in,water/steam})={\dot{m}}_g{C}_{p,g}(T_{out,g}-T_{in,g})(\mathrm{1}-E_l),

where E_l is the percentage of heat loss in each section, which is taken as 5%.

Having properties of the cold fluid at each section of the HRSG, the heat transfer at that section can be found using

(6)\dot{Q}={\dot{m}}_{in,part}(h_{out,part}-h_{in,part}).

The temperature differences of the pinch and the approach are two important parameters in the HRSG. The pinch temperature difference is the difference between the temperature of feed water and the temperature of hot gases at the point inside the evaporator at which evaporation starts (Fig. 3). Approach temperature difference is the difference between the temperatures of feed water at the point at which evaporation starts and at the economizer output (Fig. 3).

The pinch and the approach temperature differences for the HRSG are described by Eqs. (7) and (8)[15].

(7)\mathrm{\Delta }{T}_{pinch,{\left(DEA,\underset{---}{LP},\underset{===}{HP}\right)}_{eva}}=T_{g,out,{\left(DEA,\underset{---}{LP},\underset{===}{HP}\right)}_{eva}}-T_{st,{\left(DEA,\underset{---}{LP},\underset{===}{HP}\right)}_{eva}},

(8)\mathrm{\Delta }{T}_{approach,\left(DEA,\underset{---}{LP},\underset{===}{HP}\right)}=T_{st,{\left(DEA,\underset{---}{LP},\underset{===}{HP}\right)}_{eva}}-T_{st,out,{\left(DEA,\underset{---}{LP},\underset{===}{HP}\right)}_{eco}}.

Two effective parameters in the HRSG are k and y. The former is the ratio of the HRSG input feed water flow rate in the repowered cycle to the steam flow rate passing through the condenser in the original cycle working at the nominal load. The latter is the proportion of the steam flow which passes through the HRSG HP to the HRSG input feed water flow.

(9)k={\displaystyle \frac{{\dot{m}}_{in,HRSG}}{{\dot{m}}_{in,cond,first}}},

(10)y={\displaystyle \frac{{\dot{m}}_{hp}}{{\dot{m}}_{in,HRSG}}}.

5.3 Gas turbine

The compressor, the combustion chamber, and the turbine are separately analyzed in the modeling of the gas turbine(s), and the results are compared with the data obtained from the control room of the two gas stations of Roodeshoor and Sabalan having V94.3A and V94.2 gas turbines respectively (Table 4). The empirical data are taken from two internal reports available in Persian in the above mentioned power stations. The comparison shows that the computational results are accurate.

The data presented in Table 4 is for model validation. In this paper, the net power outputs of V94.2 and V94.3A gas turbines are taken to be 160MW and 293MW respectively. The environmental properties are chosen according to the geographical location of Bandarabbas. The gas turbine consists of three major sections: compressor, combustion chamber and turbine (Fig. 2).

The governing equations for the compressor (Fig. 2) are

(11){\dot{W}}_{comp}={\dot{m}}_{air}{C}_{p,a}(T_{\mathrm{2}}-T_{\mathrm{1}}),

(12)T_{\mathrm{2}}=T_{\mathrm{1}}\left\{\mathrm{1}+{\displaystyle \frac{\mathrm{1}}{{\eta }_{comp}}}[{(r_{pc})}^{({\gamma }_a-\mathrm{1})/{\gamma }_a}-\mathrm{1}]\right\},

r_pcis the compressor pressure ratio.

The governing equations for combustion chamber (Fig. 2) are

(13){\dot{m}}_{air}{C}_{p,a}{T}_{\mathrm{2}}+{\dot{m}}_{fuel}LHV+{\dot{m}}_{fuel}{C}_{p,f}{T}_{fuel}={\dot{m}}_g{C}_{p,g}{T}_{\mathrm{3}}+(\mathrm{1}-{\eta }_{Cch}){\dot{m}}_{fuel}LHV,

(14)FA={\displaystyle \frac{{\dot{m}}_{fuel}}{{\dot{m}}_{air}}},

(15){\dot{m}}_{air}={\displaystyle \frac{{\dot{W}}_{net,GT}}{(\mathrm{1}-FA)C_{P,g}(T_{\mathrm{3}}-T_{\mathrm{4}})-C_{P,g}(T_{\mathrm{2}}-T_{\mathrm{1}})}},

(16){\dot{m}}_g={\dot{m}}_{fuel}+{\dot{m}}_{air}.

For turbine (Fig. 2):

(17){\dot{W}}_t={\dot{m}}_g{C}_{p,g}(T_{\mathrm{3}}-T_{\mathrm{4}}),

(18)T_{\mathrm{4}}=T_{\mathrm{3}}\{\mathrm{1}-{\eta }_t[\mathrm{1}-{(r_{pt})}^{(\mathrm{1}-{\gamma }_g)/{\gamma }_g}]\},

(19)r_{pt}=r_{pc}{r}_{pcch},

r_pcch and r_pt are the compression ratios of combustion chamber and turbine, respectively.

5.4 Duct burner

The duct burner produces the required extra gas flow and heat according to Eq. (20) (Fig. 2).

(20)N{\dot{m}}_g{C}_{p,g,out,GT}{T}_{g,out,GT}+{\eta }_{combustion}{\dot{m}}_{fuel,db}LHV={\dot{m}}_{g,in,HRSG}{C}_{p,g,in,HRSG}{T}_{g,in,HRSG}.

6 Exergy analysis

Exergy is the maximum work a system can do to reach a dead state through a reversible process. Dead state is the state at the surrounding environment. When the system pressure and temperature is balanced with the environment without any kinetic and potential energy and it is chemically neutral, the system has reached a dead state. The exergy efficiency of a repowered power plant is the fraction of the total input exergy that is converted to net work. It is expressed by Eq. (21) [21].

(21){\eta }_{exergy,RC}={\displaystyle \frac{{\displaystyle \sum _{i=\mathrm{1}}^n{\dot{W}}_i}}{{\displaystyle \sum _{i=\mathrm{1}}^n{\dot{E}}_{f_i}}}}.

In this relation, the numerator is the sum of the work produced by both the steam turbine and the gas turbine(s) as shown by

(22){\displaystyle \sum _{i=\mathrm{1}}^n{\dot{W}}_i}={\dot{W}}_{tot,net,gt}+{\dot{W}}_{tot,net,st}.

The denominator is the exergy which enters the total repowered power plant. It equals to the sum of exergy of the fuel that enters the combustion chamber of the gas turbines and the fuel that enters the duct burner. It is defined by

(23){\displaystyle \sum _{i=\mathrm{1}}^n{\dot{E}}_{f_i}}={\dot{E}}_{f,g}+{\dot{E}}_{f,db}={\dot{m}}_{f,g}{e}_f+{\dot{m}}_{f,db}{e}_f.

The fuel exergy ex_f consist of four parts: physical, chemical, kinetic, and potential exergies [4,23,24]:

(24){ex}_f={ex}_f^{PH}+{ex}_f^{CH}+{ex}_f^{KN}+{ex}_f^{PT}.

The kinetic and potential exergies are usually neglected due to relatively low speed and height [24,25]. The others are defined as

(25){ex}_f^{PH}=(h_f-h_{\mathrm{0}})-T_{\mathrm{0}}(s_f-s_{\mathrm{0}}).

The second part of exergy is the chemical exergy, which is related to the difference between the partial pressure of gases and the environment pressure. For a mixture of gases at temperature T_o and total pressure of P_o, all the components are at temperature T_o, but their partial pressures are different from P_o. Therefore, they are not in mechanical equilibrium with the environment. The chemical exergy of a mixture of k gases is obtained using Eq. (26) [25].

(26){\bar{e}}^{CH}=-{RT}_{\mathrm{0}}{\displaystyle \sum _{i=\mathrm{1}}^k{x}_i\ln {\displaystyle \frac{x_i^e}{x_i}}}.

In this equation, x_i and x_i^e are the mole fractions of the component i relative to the gas mixture and the environment respectively.

For air and the exhaust gases (combustion products), the chemical exergy is negligible relative to physical exergy, while for a gaseous fuel, the chemical exergy has a high value. The chemical exergy of a fuel which is a mixture of k gases is calculated by Eq. (27) [26].

(27){\bar{e}}^{CH}={\displaystyle \sum _{i=\mathrm{1}}^k{x}_i{\bar{e}}_i^{CH}+{RT}_{\mathrm{0}}{\displaystyle \sum _{i=\mathrm{1}}^k{x}_i\ln x_i}}.

In this equation, the term {\bar{e}}_i^{CH} is the standard chemical exergy of the component i [25].

The specific exergy loss coefficient is defined by Eq. (28). It relates the fuel lower heating value to its chemical exergy. Its value for natural gas is approximately 1.0308 [21,23].

(28)\mathit{\xi }={\displaystyle \frac{{ex}_f^{CH}}{{LHV}_f}}.

An empirical relation derived by Ameri [21] and Moran [27] to calculate specific exergy lose coefficient for hydrocarbons with chemical formula C_xH_y is

(29)\xi =\mathrm{1.033}+\mathrm{0.0169}\left({\displaystyle \frac{y}{x}}\right)-{\displaystyle \frac{\mathrm{0.0698}}{x}}.

7 Results and analysis

Due to the results obtained from modeling, the minimum numbers of gas turbines which can provide appropriate steam conditions for steam turbines and in the framework of HRSG restrictions are 1.75 for V94.3A and 3.45 for V94.2. Hence, two gas turbines of V94.3A and four gas turbines of V94.2 are required for repowering. In this paper, the repowered cycle is modeled once for utilizing two and three V94.3A gas turbines and another time for utilizing four and five V94.2 gas turbines. Three different steam flow rates were considered in this paper:

1)The feed water flow rate passing through the heat recovery boiler equals the water flow rate passing through the condenser in the original power plant (before repowering), when it works at nominal load;

2)The boiler feed water flow rate passing through the heat recovery boiler equals the water flow rate passing through the condenser in the original power plant (before repowering), when it works at maximum load. In this case, the water flow rate is approximately 20% more than its amount at nominal load;

3)The feed water flow rate passing through the heat recovery boiler equals the water flow passing through the condenser in the original power plant (before repowering), when all of the extractions of the steam turbine(s) are closed.

The effect of the duct burner is also studied for each case. Table 5 shows the results of this study.

As seen in Table 5, the maximum efficiency of the repowered cycle is 52.04%. It occurs when two V94.3A gas turbines are utilized, no duct burner is used and feed water flow equals the steam flow at nominal load of the main power plant. The minimum repowering efficiency is 40.21% when five V94.2 gas turbines are utilized with duct burner and feed water flow equals the steam flow at nominal load.

The duct burner is used for regulation of the power output which may change with the change of climate. It is used to compensate the power reduction due to reduction of ambient air temperature or other reasons. As shown in Table 5, the use of duct burner increases the total power output, but decreases the plant efficiency.

Figure 4 demonstrates that the highest net power produced by the repowered cycle is 1262 MW. It occurs when three V94.3A gas turbines are utilized and extractions are deleted. This case has smaller exhaust loss when no duct burner is used. The minimum exhaust loss for V94.3A turbines happens when two gas turbines are utilized and the plant works at overload condition and a duct burner is used. In addition, when the V94.2 gas turbines are utilized, the highest net output power is 1169 MW when five gas turbines are used and extractions are deleted and a duct burner is used. When the V94.2 turbines are used, the lowest exhaust loss occurs when four gas turbines are used and the plant works without duct burner and the extractions are deleted.

Since power plant efficiency is the main parameter in repowering, the conditions in which the highest efficiency occurs will be used in the rest of the analyses. As a result, two V94.3A gas turbines are used, the steam flow at nominal load is considered and no duct burner is utilized.

Table 5 indicates that all the restrictions required for repowering have been observed. These restrictions are mostly related to the HRSG and are applied to pinch and approach temperature differences at different sections of the HRSG and temperature of exhaust gases. The restrictions are defined below:

1)Pinch temperature difference of each evaporator cannot be less than 5K [28−30];

2)Approach temperature difference of each evaporator cannot be less than 5K [29,30];

3)The temperature of gases which exits the HRSG should not be less than 120°C to prevent dew formation at the end of the stack [6,30].

The HRSG modeling performed by Kumar et al. illustrates the trend of feed water and gas temperature variation along HRSG. The pinch and approach temperature differences at different sections of HRSG can be found from this analysis [31]. The results of the point to point modeling of the HRSG are presented in Fig. 5, which shows the temperature variation of feed water and hot gases inside the HRSG.

The heat absorption by feed water in each section of the HRSG has a significant effect on the HRSG thermodynamic specifications and the power plant efficiency. The values of heat absorption at different sections are compared in Fig. 6. It can be seen that a large part of the heat is lost in the stack. To reduce this loss, the HRSG pressure should be increased [26]. Heat absorption depends on some parameters like the feed water working pressure, its mass flow rate and its inlet and outlet temperatures (Eq. (6)). The comparison of the HP and LP evaporators in Fig. 6 shows that increase of pressure enhances heat absorption in a section. In addition, feed water enthalpy does not change a lot in passing through the deaerator.

The increase of power and efficiency of a power plant is always an important issue in repowering. Many factors are effective in this regard, including the ratio of the HRSG inlet feed water flow to condenser inlet steam flow at nominal load (k), the ratio of the steam flow at the HP section to total input feed water flow (y), the flow rate and temperature of the HRSG input gases, the temperature of the HRSG output gases, and the flow rate of duct burner input fuel.

The heat required for the repowered plant is provided by gas turbines and duct burner. The ideal amount of this heat depends on the feed water flow rate which enters the HRSG. For a constant steam flow, if the provided heat is more than the required value, the thermal efficiency decreases, since a part of the heat exits without being used. If the provided heat is less than the required value, it enters the HRSG restrictions, something which should not occur. As seen in Fig. 6, for a constant feed water flow, the increase of duct burner heat decreases the cycle efficiency, but it does not affect the power plant output power. The reason for this is that the output power depends on the value of inlet feed water flow and the proportion of flow division between the HP and the LP sections, and all these values are constant (Fig. 7).

The increase of feed water flow entering the HRSG (increase of k) leads to an increase of the heat absorbed by steam. As a result, heat loss is reduced and the exhaust gas temperature decreases (Fig. 8).

Similarly, if the steam flow rate passing through the HP section is increased (increase of y), the heat transfer from hot gases to steam will be enhanced. Hence, the heat loss and exhaust gas temperature are reduced, as depicted in Fig. 8. The direct result of the reduction of heat loss is the increase of exergy efficiency. As a result, by increasing k and y, the boiler efficiency and the plant thermal efficiency are increased, as shown in Fig. 9.

In fact the increase of the feed water flow entering the HRSG (k) and the steam flow passing through HP section (y) has two effects:

1)The steam flow into the turbines (especially HP turbine) increases which leads to the increase in power production and efficiency of the plant.

2)More heat is absorbed from the gases in the HRSG and hence the heat loss in the boiler decreases and the efficiency of the plant is increased.

The increase of HRSG feed water flow increases the steam flow passing through the steam turbine. According to Eq. (4), the turbine output power is directly proportional to the steam flow passing through it. Hence, the steam turbine power and the net power of the whole power plant are increased as the result of increase in flow of HRSG feed water (Fig. 10).

Another effective parameter in the power plant net power output is the steam flow passing through the HP section of the HRSG, which is represented by y. This flow is sent to the HP section of the turbine and then goes through the IP section after reheating. According to Eq. (4), the power plant net power output is directly proportional to the HP steam flow rate. Therefore, the power plant output power increases by increase of y. This fact is also shown in Fig. 10.

A major part of energy loss is related to the heat released by exhaust gases. A part of this heat is irreversible and is a power plant requirement. For example, a temperature of 120°C is required to prevent dew formation and to provide the buoyancy force required for the flow of exhaust gases. However, higher temperatures increase heat loss and decrease system efficiency and power output. These facts are illustrated in Fig. 11.

8 Conclusions

Due to Iran economic conditions after cutting the subsidies, optimal utilization of fossil fuels is the major priority in saving energy. The current paper attempted to enhance the performance of Bandarabbas steam power plant which played a key role in Iran electricity production. To do so, the full repowering method was selected on account of the long working life of the power plant. Because of the structure of Bandarabbas steam power plant and availability of the HRSGs in Iran, double-pressure HRSG with reheat was selected for repowering. The repowered plant was investigated for three different values of steam flow rate and for two gas turbines of V94.2 and V94.3A. The effect of duct burner was also investigated in each case. The results indicated that exergy efficiency due to repowering was increased by 34.5% when the V94.3A gas turbines were used and it was increased by 12.8% when the V94.2 gas turbines were used. The lowest exhaust heat loss happened when two V94.3A gas turbines were utilized with a duct burner at the steam flow rate of overload condition (135.5 MW). The highest rate of exhaust heat loss occurred when five V94.2 gas turbines were utilized with a duct burner and at the steam flow of nominal load (767.5 MW).

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