DEVELOPMENT OF A TOOL FOR SIMULATING PERFORMANCE OF SUB SYSTEMS OF A COMBINED CYCLE POWER PLANT

Transcription

1 DEVELOPMENT OF A TOOL FOR SIMULATING PERFORMANCE OF SUB SYSTEMS OF A COMBINED CYCLE POWER PLANT PRABODHA JAYASINGHE Master of Science Thesis Stockholm, Sweden Year

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3 DEVELOPMENT OF A TOOL FOR SIMULATING PERFORMANCE OF SUB SYSTEMS OF A COMBINED CYCLE POWER PLANT Prabodha Jayasinghe MSc Thesis 2012 Department of Energy Technology Division of Heat and Power Technology Royal Institute of Technology Stockholm, Sweden iii

4 Master of Science Thesis GI MSC EKV889 Development of a tool for simulating performance of sub systems of a combined cycle power plant Approved Examiner Prabodha Jayasinghe Supervisor Dr. Anders Nordstrand Dr. Anders Nordstrand Commissioner Professor R.A. Attalage Dr. Anders Nordstrand Contact person Dr. Anders Nordstrand Abstract In Sri Lanka, around 50% of the electrical energy generation is done using thermal energy, and hence maintaining generation efficiencies of thermal power plants at an acceptable level is very important from a socio-economic perspective for the economic development of the country. Efficiency monitoring also plays a vital role as it lays the foundation for maintaining and improving of generation efficiency. Heat rate, which is the reciprocal of the efficiency, is used to measure the performance of thermal power plants. In combined cycle power plants, heat rate depends on ambient conditions and efficiencies of subsystems such as the gas turbine, Heat Recovery Steam Generator (HRSG), steam turbine, condenser, cooling tower etc. The heat rate provides only a macroscopic picture of the power plant, and hence it is required to analyse the efficiency of each subsystem in order to get a microscopic picture. Computer based modelling and simulation is an efficient method which can be used to analyse each subsystem of a combined cycle power plant. Objective of this research study is to develop a computer based tool which simulates the performance of each of the subsystems of a combined cycle power plant of rated capacity 163MWe in Sri Lanka. At the time this research was commenced, analysis on the power plant was focused only on the heat rate, but performances of subsystem were not investigated. In this research analysis of the plant is divided into several main systems, in order to study them macroscopically. Then, these main systems are further divided into subsystems in order to have a microscopic perspective. Engineering equation solver (EES) was used to develop the tool for modelling and simulation, and the final computer model was linked with Microsoft excel package for data communication. Final computer model is employed using both present and past operating data in order to compare present and past performance of the power plant. iv

5 In combined cycle power plants steam is injected into the gas turbine to reduce the NOx generation and this steam flow is known as NOx flow. According to the result it was found that turbine efficiency drops by 0.1% and power output increases by 1MW when NOx flow increases from 4.8 to 6.2kg/s. Further it concluded that gas turbine efficiency drops by 0.1% when ambient temperature increases by 3 oc; and gas turbine power output decreases by 2MW when ambient temperature increases from 27 to 31 degrees. Regarding the steam cycle efficiency it was found that steam turbine power output drops by 0.5MW when ambient temperature increases from 27 to 31 o C; and steam cycle efficiency increases by 1% when NOx flow increases from 4.8 to 6.2kg/s. Further, the steam turbine power output decreases by 0.25MWe When NOx flow increases from 4.8 to 6.2kg/s Heat rate, which is the most important performance index of the power plant, increases by 10units (kj/kwh) when ambient temperature increases by 3 o C. Heat rate also increases with increase NOx flow which was 6.2kg/s in year 2007 and 4.2kg/s in year Hence, heat rate of the power plant has improved (decreased) by 10units (kj/kwh) from year 2007 to year Other than above, following conclusions were also made during the study using model developed with the perspective of study. 1) HRSG efficiency has not change during past 4 years 2) Significant waste heat recovery potential exists in the gas turbine ventilation system in the form of thermal energy v

6 Acknowledgments It is with great pleasure that I would like to first extend my sincere gratitude to Professor Rahula Attalage for his valuable guidance, advice and supervision throughout this project. The time he spared to examine critical results, is also fondly remembered here. I also would like to extend my courteous gratitude to Dr. Anders Nordstrand for his assistance during the thesis project as well as Dr. Primal for his patience and steadfast encouragement to complete this study. Furthermore, I must mention Ms. Shara Osman and Ms. Chamika Kalhari as they were extremely supportive of this project with their assistance in many aspects from the beginning to its very completion vi

7 List of Abbreviations HRSG FOSV IGV LLP LP HP HPBFP LPBFP CEP IBD CBD HPSV HPCV CST SWAS CW CT CWST Heat recovery steam generator Fuel oil stop valve Inlet guide vanes Low Low pressure Low pressure High pressure High pressure boiler feed pump Law pressure boiler feed pump Condensate extraction pump Intermittent blow down Continues blow down High Pressure Steam Valve High Pressure Control Valve Condensate Storage Tank. Steam water analysis system Cooling water Cooling tower Cooling water storage tank vii

8 Development of a tool for simulating Performance of sub systems of a combined cycle power plant 1 Introduction Overview Objective of the project Present scenario Proposed system Data gathering Methodology Literature survey Gas turbine Brayton cycle Development of gas turbine Applications of gas turbines Performance of gas turbine Atmospheric temperature Atmospheric humidity Fuel type Steam or water injection Air extraction HRSG Steam turbine Classification of steam turbine Condensing and non condensing steam turbines Single and double flow steam turbines Impulse and reaction steam turbines Condenser Direct contact condensers Surface condensers Cooling towers Wet cooling towers Dry cooling towers Plant overview Preface Heat recovery steam generator (HRSG) Fuel system Fuel storage Fuel centrifuge Fuel forwarding system...31 viii

9 4.4 Gas turbine fuel oil skid Fuel flow meter Fuel filter Fuel accumulator Gas Turbine Gas turbine building Accessory compartment Turbine compartment Exhaust compartment Load compartment Generator compartment Air intake system Gas turbine auxiliary systems Gas turbine accessory gearbox Fuel injection system Fuel pump High pressure fuel filter Lubrication oil system Atomizing air system Gas turbine cooling water system Heat recovery steam generator (HRSG) Low low pressure (LLP) circuit Low pressure (LP) circuit High pressure (HP) circuit Steam turbine Condensate system Cooling water system SWAS (Steam water analysis system) System modelling Combined cycle heat rate Gas turbine efficiency Gas turbine cooling water system Gas turbine fuel energy Gas turbine electrical power Gas turbine steam injection Gas turbine air intake and exhaust Gas turbine compartment convective losses Gas turbine ventilation system Turbine compartment ventilation system Exhaust compartment ventilation system Load compartment ventilation system Gas turbine efficiency Gas turbine compressor Heat recovery steam generator (HRSG)...61 ix

10 HRSG modules HP Economizer HP Economizer HP Economizer HP Super heater HP Super heater HP evaporator LP economizer LP evaporator LP supper heater LLP evaporator Steam Cycle Steam Turbine efficiency Steam Turbine isentropic efficiency Condenser efficiency Software architecture Data processing EES program structure Analysis and results Performances comparison 2007 vs Heat rate Gas turbine efficiency Compressor isentropic efficiency of gas turbine Turbine isentropic efficiency of gas turbine HRSG efficiency Steam cycle efficiency Heat transfer in the HP economizer 1 at full load Heat transfer in the HP super heater 1 at full load Heat transfer in the HP super heater 2 at full load Heat transfer in the HP evaporator Heat transfer in the LP economizer Heat transfer in the LP evaporator Heat transfer in the LP supper heater Heat transfer in the LLP evaporator Effects of NOx flow Heat input to GT from NOx steam Gas turbine power output Gas turbine efficiency Steam turbine power out Steam cycle efficiency Heat transfer in Condenser Plant overall heat rate Discussion and Conclusions Heat Rate Gas turbine HRSG Steam cycle Heat transfer in the condenser...91 x

11 14.6 Waste heat recovery potential Annexure Enthalpy of air and gasses from light fuel oil combustion Typical pretest stabilization period Heat rejection form gas turbine ventilation system Reference xi

12 1 Introduction 1.1 Overview Total grid installed capacity of Sri Lanka is around 2818MW [1], and thermal and hydro power plants are providing significant amount of electricity into the grid. Energy Installed capacity source Hydro Thermal Renewable Total Installed capacities of Sri Lanka [1] Electricity generation cost is directly associated with the generation efficiency, thus requiring the sustenance of power generation efficiency at satisfactory level with the provision of financial as well as environmental benefits. Poor generation efficiency results from ageing of equipment, equipment malfunction or design faults. Thus, it is necessary to monitor efficiency precisely and continuously because variations can be identified and possible corrective action can be taken to bring back the efficiency to original level. In a thermal power plant, a performance index called heat rate, which is the reciprocal of the efficiency, is used to measure the performance of the plant instead of thermal efficiency. Heat Rate = Heat energy consumed during fuel combustion (kj) / Electrical Energy produced (kwh) 1.2 Objective of the project Thesis research was done in one of the combined cycle power plant, which is 163MW in capacity, in Sri Lanka Present scenario Presently, heat rate is calculated on a monthly basis using monthly fuel consumption and electrical energy production. But, this figure does not reflect a correct picture on the thermal performance of the plant due to the following reasons. According to the demand the plant load is controlled and, generally, during day time the plant runs at full load whereas, at night, operates at part load or shut down. When the plant operates at part load, efficiency is less and higher heat rate exists. Thus, when heat rate is calculated on using monthly fuel consumption and monthly electrical energy production, only a generalized value can be obtained for the entire operated load spectrum. Hence, above heat rate varies with the monthly plant load factor. During plant start-ups significant amount of fuel consumption takes place affecting the heat rate in terms of the number of start-ups that take place during the month. Due to above reasons heat rate, which is calculated using present method, fluctuates from month to month. Thus, it is not possible to get a clear picture regarding the actual efficiency of the plant, hence not possible to compare the present performance of the plant with the past figures. One objective of this project is to develop a system which will calculate the heat rate excluding part load conditions and star-up. 12

13 The heat rate is giving only a macro picture about the thermal performance of the power plant. But, whole power plant consists of several sub systems such as Gas turbine, Lubrication oil system of gas turbine, Steam turbine, Feed water pumping system, HRSG, etc. Overall plant heat rate depends on efficiencies of these sub systems. Thus, these sub systems must be analysed in order to get a micro picture about the plant performance. In other words heat rate can only be used to check the existence of decreasing trend in thermal performance. But, to find the root cause for decreasing trend, it is necessary to have micro picture of each sub system. Currently there is no tool to obtain the efficiencies of sub systems of the power plant. Developing a tool which provides efficiencies of each sub system will enable to figure out which sub systems affects the plant efficiency prominently. Having identified the sub systems which contribute most to the plant efficiency, a detailed analysis can be performed to identify the measures that can be taken to increase the efficiency of the subsystem and, thereby to increase the overall plant efficiency Proposed system Main objective of the project is to develop a tool which cab be used to analyse the efficiencies of sub systems of the power plant. This tool is suppose to provide energy inputs to sub systems, energy outputs from sub systems, losses and efficiencies of sub systems. Further proposed tool should not give generalized figures for the entire operated load spectrum. Instead it should give specific figures for a given load. 1.3 Data gathering All vital plant operating parameters such as temperatures, pressures, and flow rates are measured using electronic measuring devices, and values are sent as electronic signals to control room of the plant. Also these parameters are automatically saved in an access database. Parameters, which are less important for the operation, such as lubrication oil pressure in individual bearing, jacking oil pressure in individual bearing, discharge and suction temperatures of gas turbine cooling water pump are not sent to the control room as electronic signals. Hence, these readings are taken from local gauges by plant operators every four hours, and they record these readings in logbooks. 1.4 Methodology As the first step entire power generation process was analysed macroscopically. Then, the plant was divided into the following main systems with the aim of analysis to be done on each subsystem for performance. Gas turbine HRSG Steam cycle During the next step above main systems were divided further into following sub systems, and individual performance of these sub systems were also analysed. Compressor Turbine High pressure circuit Low pressure circuit Low low pressure circuit Steam turbine Condenser 13

14 Thus, the system was modelled to record the indices related to macroscopic performances as well as microscopic performance of the power plant. The system modelling was carried out using EES (Engineering equation solver) software. This software was selected based on availability of the software, user friendliness, and the possibility to link the software with Microsoft excel. 14

15 2 Literature survey 2.1 Gas turbine Brayton cycle, which is the fundamental concept behind the gas turbine, was first proposed by George Brayton in The cycle consists of four reversible processes. [2] 2.2 Brayton cycle Figure 2.1: Elements of gas turbine 1-2 Isentropic compression (in a compressor) 2-3 Constant pressure heat addition 3-4 Isentropic expansion (in a turbine) 4-1 Constant pressure heat rejection Work input in the compressor Figure 2.2: P-v and T-s diagrams of Brayton cycle = W in W in = m& air (h 2 h 1 ) Work output from the turbine = W out 15

16 W out = m& air (h 3 h 4 ) Heat input = Q in Q in = m& air (h 3 h 2 ) = m& air cp (T 3 T 2 ) Heat rejection = Q out Q out = m& air (h 4 h 1 ) = m& air cp (T 4 T 1 ) Thermal efficiency of ideal Brayton cycle η Brayton Applying steady state steady flow equation to whole system Wout Win = Qin - Qout η Brayton Qout T 4 T = 1 = 1 1 Qin T 3 T 2 Process 1-2 and 3-4 isentropic W out W = in Qin K -1 K -1 T2 P2 K T3 P3 = K T1 P1 = T4 P4 K - Specific heat ratio P2 = P3 P1 = P4 η Brayton 1 = 1 K -1 P2 K P1 Thus according to the ideal Brayton cycle, efficiency of a gas turbine depends on pressure ratio and the specific heat ratio of air. When pressure ratio increases, cycle efficiency raises. But, it is necessary to maintain temperature at point 3 below a certain value which is the maximum possible temperature that the turbine blade can withstand. Net work output increases with increasing pressure ratio, and start decreasing after reaching a maximum value. Hence, it is necessary to compromise thermal efficiency and the net work output. Performance indicator called heat rate is used in thermal power industry instead of thermal efficiency. Heat Rate = Heat Energy consumed during fuel combustion Electrical Energy produced ( kwh) ( kj ) Thus heat rate has units, and when thermal efficiency increases with decreasing heat rate. 16

17 Figure 2.3: T-s diagrams of actual gas turbine cycle h 2s h η 1 iso, compressor = h 2a h 1 h 3 h η 4a iso, turbine = h 3 h 4s Process of an actual gas turbine deviates from ideal Brayton cycle due to following two reasons. Brayton cycle consists of four processes, but in reality constant pressure heat rejection does not takeplace inside the gas turbine. Instead, flue gas coming from the turbine is exhausted to the atmosphere, and fresh air is inducted into the compressor from the atmosphere. Second reason for the deviation is the presence of irreversibilities in an actual gas turbine. Actual work input of the compressor is higher than the ideal conditions, and the actual work output of the turbine is less than the ideal conditions. Deviation of the actual process from the ideal condition can be analysed using isentropic efficiencies of the turbine and the compressor. Apart from above mentioned irreversibilities, pressure drops during heat addition and heat rejection also take place. [3] 2.3 Development of gas turbine Gas turbine started to evolve rapidly since Efficiencies of compressor and turbine of early gas turbines were relatively lower; and the turbine inlet temperature could not be increased due to metallurgical limitations. Hence, efficiencies of early gas turbines were around 17 percent. Many research works were done to improve the efficiency of gas turbine, and it was possible to increase the turbine inlet temperature due to the invention of new cooling techniques and materials. Today turbine inlet temperature of a gas turbine is around 1425 o C, and in 1940s this value was around 540 o C. But, at higher combustion temperatures, NO x formation is significant, and hence steam injection was carried out into combustion chamber to reduce the flame temperature; and thus NO x formation reduces. Other than turbine inlet temperature, efficiencies of components of gas turbine were improved in order to increase the overall efficiency of gas turbine. Intercooling, regeneration, and reheating were also used to improve the overall efficiency of gas turbine [4]. 2.4 Applications of gas turbines Gas turbines are used in many industries such as power generation, aircraft, ships, helicopters, locomotives, and even in automobiles. Power to weight ratio of gas turbine is very high compared to reciprocation engines. At higher altitudes performance of gas turbine is much higher than a reciprocating engine, and hence gas turbines are very popular in aeronautical industry. Gas turbines are generally more expensive than reciprocation engines, and efficiency of a gas turbine varies drastically with varying shaft rpm. Hence gas turbines are not very attractive for automobile industry, and less popular. [5] 17

18 2.5 Performance of gas turbine Atmospheric temperature Power output of a gas turbine varies with the mass flow rate, thus performance of a gas turbine significantly depends on the density of atmospheric air which means ambient temperature plays a vital role regarding the performance of gas turbine. When ambient temperature increases density drops; and this cause power output and efficiency drop. Thus, it is not possible to compare heat rates, which were calculated in two different occasions, of a particular gas turbine, unless ambient temperatures are same for both occasions. But a concept called corrected heat rate is used in thermal power industry as solution for the above problem. Actual heat rate that is also called uncorrected heat rate is divided by a factor called ambient temperature correction factor to obtain the corrected heat rate. This correction is only for ambient temperature. Corrected heat Actual rate = (corrected for ambient temperature) Correction ambient heat rate factor for temperature This correction factor is function of ambient temperature, and it compensates the effects of ambient temperature from the heat rate. This correction factor depends on the type and the model of the gas turbine and manufacturer specify the correction factors. [6] Figure 2.4: Effect of ambient temperature Atmospheric humidity Similarly relative humidity also affects the performance of the gas turbine. Since, density of humid air is less than the density of dry air, power out drops with increasing relative humidity. Heat rate also increases when relative humidity rises. It is possible to compensate the effects of relative humidity on dividing the actual heat rate by the relative humidity correction factor. [6] Actual Corrected heat rate (corrected for relativ humidity) = Correction heat factor relative humidity rate for 18

19 Figure 2.5: Effect of relative humidity Fuel type Work output of a gas turbine depends on the specific heat capacity ( cp ) of the combusted gas. Q out = m& air (h 4 h 1 ) = m& air cp (T 4 T 1 ) This specific heat capacity of the combusted air depends on the fuel composition, and hence fuel type also one of the major factor that can vary the performance of the gas turbine. [7] Figure 2.6: Effect of relative humidity 19

20 2.5.4 Steam or water injection In modern gas turbines water or steam injection is carried out into the combustion chamber in order to reduce the NOx generation. Since, this is an additional mass flow; output of the gas turbine increases with the steam (or water) injection. When water is injected, some amount of heat is absorbed by the water, in order to raise the temperature to combustor condition. Thus, heat rate deteriorates when water is used for NOx reduction. When steam is used for NOx reduction, heat rate improves if the heat consumed to produce the steam is neglected. But, in most of the applications this heat can t be neglected, and hence heat rate get worse even when steam is used for NOx reduction. [8] Figure 2.7: Effect of steam injection into gas turbine Air extraction Compressor of the gas turbine consists of number of stages, and in each stage air pressure increases. Hence, pressure of the 1 st stage is at the lowest pressure and the pressure of the last stage is at highest. Hence, it is possible to extract air from various stages of compressor; and extracted air can be used for applications such as bearing sealing, turbine blade cooling, atomizing air for combustion etc. When air extraction increases, mass flow rate through the turbine decreases, and turbine out put power drop.[8] 2.6 HRSG Heat recovery steam generator (HRSG) can be considered as a heat exchanger which recovers heat from hot gas in order to produce steam. In combined cycle power plants, heat recovery steam generators are used to recover heat from the gas turbine exhaust, and to generate steam that rotates the steam turbine. Cogeneration plants also use heat recovery steam generators to manage energy in an efficient manner. HRSG consist of three main components called economizer, evaporator and supper heater. Water that is at sub cooled state enters the economizer, and the exit condition is very close to saturated liquid state. In the evaporator, saturated water becomes saturated steam that will be further heated inside the supper heater to produce super heated steam. Finally flue gas, from which, heat absorbed into the circulation fluid, goes to the atmosphere through the HRSG stack. HRSGs can be classified into two types. HSRGs, in which, flue gas flows vertically, are called vertical type HRSGs. Similarly in horizontal type, HRSG flue gas flow takes place horizontally. 20

21 Figure 2.8: Vertical HRSG The main advantage of the vertical HRSG is that it requires comparatively less land area. Some HRSGs are associated with supplementary firing unit which supplies more heat into the HRSG, and thereby produce more steam. Some power plants often use supplementary firing to cater peak demand. Some power plants have additional stacks to bypass the HRSG if required. In this type of a plant, a diverter valve is installed in the bypass stack entrance, to control the flue gas. This allows operating the gas turbine, with out HRSG in operation. [9] Most of the HRSGs have drums which are used for several functions. 1) Acts as a water storage 2) Helps to maintain the chemical balance of the working fluid 3) Separate water and steam 2.7 Steam turbine Steam turbine is a device that absorbs energy from steam and coverts into mechanical work. Initially reciprocation piston steam engines were used in the industry, and later rotary steam turbines became popular due to its advantages over reciprocating type. Efficiency and power to weight ratio of rotary steam turbine are higher compared to reciprocation piston steam engines. [10] Classification of steam turbine Steam turbine can be classified in many different manners Condensing and non condensing steam turbines In condensing steam turbines exhaust pressure is maintained below atmospheric pressure. This type of steam turbines is commonly used in power plants, and steam quality of the exhaust is generally around 90%. Exhaust pressure of non-condensing steam turbines is at higher pressure than the atmosphere, and this type of steam turbines is commonly used in refineries, district heating units, pulp and paper plants, and desalination facilities. In these steam turbines, exhaust is used in the process. 21

22 Single and double flow steam turbines Steam flow produces forces on both tangential and axial directions, and tangential force generates a torque which rotates the turbine. The axial thrust is imparted onto the structure that holds the entire steam turbine assembly by mean of a thrust bearing, and this axial thrust increases with increasing turbine capacity. In double flow steam turbines, steam injection taken place at the center of the shaft and leaves at both ends. Figure 2.9: Double flow steam turbine The blades of a one half of the turbine face the opposite direction of the blades of the other half. Hence, tangential forces of both haves generate a torques on the same direction while the thrust force of one half of the steam turbine compensates the thrust force of the other half of the steam turbine. Figure 2.10: Double flow steam turbine [10] Impulse and reaction steam turbines In impulse turbines, velocity of steam increases as it flows through the nozzle. Hence, pressure drop taken place across. Exit steam of the nozzle impinges upon the blade of the turbine and cause deflection. This creates a momentum reduction in the steam jet, and hence momentum of the blade rises. Across the blade, no pressure drop take-place, and only velocity change exists. 22

23 Figure 2.11: Impulse and reaction steam turbines [10] In reaction turbines, velocity of steam increases and pressure drops in the nozzle similar to impulse turbine. But as the steam flows across the blade both velocity and pressure drops in the axial direction. This creates a pressure difference across the blade in the tangential direction, and hence a torque generates around the shaft. [11] 2.8 Condenser The main objective of a condenser is to act as a heat exchanger on transferring heat from steam turbine exhaust to cooling water. Due to heat transferring process, steam turbine exhaust get condensed in the condenser. Condensers are mostly associated with a temporally condensate storage called hot well, into which, condensate flows from the condenser. [12] Condensers can be categorized into two main types as follows. a) Direct contact condensers b) Surface condensers Direct contact condensers In this type of condensers, cooling water and condensate directly mix inside the condenser, and hence there is only one out flow stream. This type of condensers can be categorized into three sub types as spray condenser, barometric condenser and jet condenser. In spray condensers, cooling water is sprayed into the steam. In the barometric condensers, cooling water falls down through baffles, and steam inlet is located underneath these baffles. Steam gets condensed, and the mixture flows out through a tail pipe that is located at the bottom of the condenser. Only difference between barometric condenser and jet condenser is that tail pipe is replaced with a diffuser in the jet condensers. This diffuser raise the inside pressure within short distance. 23

24 Figure 2.12: Spray condenser [13] Figure 2.13: Barometric condenser [14] Surface condensers In contrast to direct contact type, mixing is not taken place in surface condensers, and there are separate two inlets and separate two outlets. In power plants, these types of condensers are frequently encountered, and they are essentially shell and tube heat exchangers. Steam flow outside the tubes, and cooling water flows inside the tubes. Figure 2.14: Surface condenser [15] 24

25 2.9 Cooling towers Cooling towers are used to cool the water that comes from the condenser, and cooling towers can be categorized into two main types as wet type and dry type. [16] Wet cooling towers Figure 2.14: Wet cooling tower [17] In this type of cooling towers, hot water spreads over a net of slats or bars called packing, through which, water flows downwards. Atmospheric air enters into the cooling tower through louvers that are located at bottom of the cooling tower, and leaves from the top. Thus, air and water flows in counter directions, and this will allow mixing water and air thoroughly. During the mixing process, water evaporates, and due to absorption of latent heat water gets cooled. Other than due to evaporation, convection heat transfer takes place from surfaces of the water droplets to the air. Cold water collects in a basin called cooling tower basin at the bottom of the tower. It is possible to again categorized wet cooling tower into two sub types, forced draught (FD) and induced draught (ID). Forced draught type cooling towers are associated with fans that creates the air flow. Density of air that is in the cooling tower is less than the out side air, and this density difference creates a flow through the cooling tower, and in forced draught cooling towers this natural air flow is sufficient to provide the cooling effects. Water evaporation in wet cooling towers is significant, and hence considerable amount of make-up water must be supplied into the cooling tower basin in order to maintain the water level. Dry type cooling tower are used when it difficult to maintain such a large amount of make-up water. 25

26 2.9.2 Dry cooling towers Figure 2.15: Wet cooling tower [18] Dry cooling towers consist of finned tubes, in which, warm water flows, while air flow is taken place outside of the tubes. Thus, heat transfers from warm water to air, through the finned tube walls. In dry cooling towers air flow is maintained by means of fans. 26

27 3 Plant overview 3.1 Preface Figure 3.1: Plant overview The plant consists of one gas turbine, heat recovery steam generator (HRSG) and a stem turbine. A water treatment plant, which is installed in the same premises, is fulfilling the water requirement of the power plant, and effluent treatment is also carried out in this water treatment plant. Fuel oil purification is carried out using centrifuges, and these centrifuges are also located in the premises of the water treatment plant. HRSG consists of three fluid circuits which are used to heat the working fluid (water); and three different pressure levels are maintained in these circuits Apart from HRSG, there are number of other sub systems which are associated with the power generation process. 1) Condenser - Exhausted steam, which are coming from the steam turbine, condense inside the condenser. 2) Feed water pumping system Pumping system is used for water circulation 3) Cooling towers Cooling water, which is used in the condenser, cools in cooling towers 4) Cooling water pumping system High capacity cooling water pumps were installed in the power plant to maintained cooling water circulation through condenser and the cooling towers. 5) Lubrication oil system of gas turbine Turbine shaft is mounted on six journal bearings. Lubricating oil is pumped into these bearings and oil layer is maintained between shaft and the bearing metal surface. A cooling system is placed to extract heat from the lubrication oil. 27

28 6) Hydraulic oil system of gas turbine Hydraulic oil is in placed to drive some equipment of the gas turbine. 7) Fuel system Fuel system is installed to maintain fuel flow of the gas turbine. 8) Lubrication oil system of steam turbine 9) Hydraulic oil system of steam turbine Gas turbine capacity of the plant is 110MW; where as the capacity of the steam turbine is 57MW, and net power output at rated condition is 163MW. The flue gas coming from the gas turbine goes to the HRSG (Heat recovery steam generator) via a by pass damper which is a two way damper. Bypass damper can be used to divert the flue gas into bypass stack while blocking the flue gas path to HRSG. It can also be used to send the flue gas into HRSG having blocked the bypass stack. During the plant start up, heat exchanger modules of HRSG may not filled with the water, and it is not be possible to heat the HRSG until the heat exchanger modules get filled with water. Therefore during the plant start up, flue gas is diverted through the bypass stack, until HRSG heat exchanger modules are filled. Another purpose of the by pass damper is to stop the heat input to the HRSG immediately, in case of an emergency shutdown. 3.2 Heat recovery steam generator (HRSG) HRSG consists of three fluid circuits which are used to transfer working fluid (water); and three different pressure levels are maintained in these circuits. Circuit, which maintains the highest pressure, is called higher pressure (HP) circuit, and the pressure inside this circuit is around 100bar. Circuit, which maintain the lowest pressure, is called Low Low pressure (LLP) circuit and the pressure in side this circuit is around 1.8bar. The other circuit is called low pressure (LP) circuit, and it maintains intermediate pressure around 15.5bar. There are three drums that are associated with each circuit, and these drums are called HP drum, LP drum, and LLP drum. Steam turbine exhaust is cooled using condenser, and the bottom part of the condenser is called hot well, and it contains condensate. This condensate is pumped back to the LLP drum by means of condensate extraction pumps (CEP). Separate demineralised water supply line is connected to hot well for makeup water requirement. Water pumped from LLP drum to HP drum using high pressure boiler feed pumps (HPBFP), and similarly water is pumped from LLP drum to LP drum by low pressure boiler feed pumps (LPBFP). High pressure and low pressure steam, which are coming form HP and LP circuits, are injected into the gas turbine. Steam, which is at 24bar, is extracted from the steam turbine, and injected into gas turbine in order to limit the combustion temperature. This is because NO X generation is high at higher combustion temperature, and hence in normal operating conditions steam injection is carried out into gas turbine to achieve environmental standard. Similarly, steam, which is at 5bar, is extracted from the steam turbine, and injected into LLP drum for pre heating. Deration also taken place in the LLP drum due to steam injection. 28

29 4 Fuel system 4.1 Fuel storage Figure 4.1: Fuel storage system High speed diesel (HSD) is used as the fuel for power generation in the power plant, and this fuel is stored in storage facility which consists of two raw fuel tanks and two treated fuel tanks. Each tank is 6100m 3 in capacity. Fuel that is taken from the fuel supplier is pumped into raw fuel tank through a filter which used to carry out preliminary filtering process. But this filtering process is not sufficient for combustion in the gas turbine. Hence, fuel, that is stored in raw fuel tanks, is pumped via centrifuges, where further purification taken place, into treated fuel tanks. For every fuel delivery, fuel suppler provides a lab report regarding the fuel quality, and this report gives the higher heating value of fuel. Usually when level of a raw fuel tank, goes below 5-10m, new fuel delivery is started. Hence, fuel from the new delivery mixed with fuel from the previous delivery in the raw fuel tanks. Thus, mixture of fuel from different deliveries, are stored in the raw fuel tanks. Hence, higher heating value given in the lab report is different from the actual heating value of fuel that is used in the gas turbine. But for the analysis it is necessary to know the higher heating value of the fuel. Hence, it was started taking fuel samples from the fuel tanks, in order to measure the actual higher heating value of the fuel. These samples are sent to an industrial laboratory to measure the higher heating value. Presently fuel sample is taken after every fuel tank change-over as an operational practice. This new practice was initiated due to recommendations raised after the project. 29

30 4.2 Fuel centrifuge Figure 4.2: Fuel centrifuges Figure 4.3: Fuel centrifuge system Centrifuges are used to purify fuel by removing unwanted material in the fuel. During initial stage of the centrifuge process, de-mineralised water is mixed with liquid fuel. Finally, added water and impurities remove from fuel as sludge. After centrifuge process, purified fuel is stored in treated fuel oil tank. 30

31 Advantages of fuel centrifuge [19] Reduce high temperature corrosion Reduce ash fouling deposits Remove trace metal contaminants Water Separation Filtration of fuel to remove solid oxides, silicates and other harmful contaminants 4.3 Fuel forwarding system Figure 4.4: Fuel forwarding system Centrifuged fuel that is stored in treated fuel tanks is pumped to gas turbine fuel oil skid. Two fuel forwarding pumps, which are parallel, are used to pump fuel from treated fuel tanks to gas turbine fuel oil skid. When the power plant is running, only one fuel forwarding pump is in operation, and the other one acts as the standby pump. Return line is connected to pump discharge line, and fraction of fuel returns to treated fuel tank. A control valve, which is located in this return line, controls the fuel supply pressure. 31

32 Figure 4.5: Duplex fuel filter Fuel filter, which is 40micron in filtration rating, is used in the suction side of the forwarding pumps. This fuel filter is duplex type which means two filters are connected in parallel, and one filter is in operation and the other one acts as a standby filter. With time, particles accumulate inside the filter, and this cause higher differential pressure between the filter, and when this differential pressure rise beyond 0.5bars filter change over must be done. This filter change over can be done without interrupting the fuel flow. Similar filter, which is 20micron in filtration rating, is used in the discharge side of the forwarding pumps for further purification. 4.4 Gas turbine fuel oil skid Fuel forwarding system pumps fuel to gas turbine fuel oil skid, which consists of flow meter, fuel accumulator, and the duplex fuel filter Fuel flow meter Fuel flow meter is used to measure the fuel flow rate of the gas turbine. As far as energy analysis is concern, this meter is very important, since total energy input to the plant can be calculated using this flow meter Fuel filter After fuel flow meter, another duplex filter is located in the fuel line for the filtration, and it uses 5 micron paper elements for filtration Fuel accumulator In case of an emergency shutdown of the plant, a valve called fuel oil stop valve (FOSV) which is in the down stream of the gas turbine fuel oil skid, closed immediately in order to stop the fuel flow into gas turbine. When this happened, pressure inside the fuel line rises momentarily. This sudden pressure rise is dampened by the fuel accumulator which is connected to the outlet of the filter. 32

33 5 Gas Turbine 5.1 Gas turbine building All the equipments associated with the gas turbine are located in the gas turbine building, and equipment contained in several compartments called accessory compartment, turbine compartment, load compartment, and generator compartment Accessory compartment There are several auxiliary systems such as fuel injection system, lubrication system, atomizing air system, hydraulic system, turning gear system, starting system, and these systems are necessary for the operation of the gas turbine. Most of the electrical and mechanical equipments of above auxiliary systems are contained within the accessory compartment Turbine compartment Compressor, turbine, combustion wrapper and some auxiliary equipment are contained within the turbine compartment. Fuel is injected into compressor discharge through units called combustion cans, and there are fourteen number of combustion cans in the turbine compartment. Main fuel supply line splits into fourteen numbers of sub lines, and provide fuel for combustion cans. In order to combust, liquid fuel must be atomized, and this is done by injecting fuel and pressurized air through a nozzle at a high velocity. This pressurized air is called atomizing air, and a compressor, which is in the accessory compartment, provides atomizing air for combustion cans. There are 17 numbers of compression stages in the compressor, and each stage consists of stationary blade and a rotary blade. When air flows through these compression stages air get compressed, and compressor discharge pressure is around 10.5bar when plant is at full load. This compressed air then comes to a chamber called combustion wrapper. Figure 5.1: Stationary blades of the compressor during maintenance work 33

34 Figure 5.2: Rotary blades of the compressor during maintenance work This combustion wrapper is the source, which provides the secondary air to the combustion cans. Atomizing air, which is the primary air, and liquid fuel also enters into combustion cans, in which, combustion take place. Then heated air and combustion products are directed to the turbine. There are three number of turbine stages exists in the turbine; each stage consists of stationary blade and a rotary blade. This stationary blade, which is also known as turbine nozzle, changes the direction of air flow. Energy transfer takes places in the rotary blades which are also known as buckets. Third stage bucket Second stage bucket First stage nozzle Figure 5.3: Nozzles and buckets First stage nozzles and buckets are exposed to very high temperature, and hence cooling air flow is maintained through these nozzles and buckets. 34

35 Second stage nozzle Third stage nozzle First stage bucket Second stage bucket Third stage bucket Figure 5.4: Nozzles and buckets At high combustion temperatures NOx generation take place, and hence steam is injected onto the flame of the gas turbine to reduce NOx generation. This is called NOx steam, and a steam header that comes from steam cycle provides NOx to combustion cans of the gas turbine Exhaust compartment Gas turbine exhaust goes into the HRSG through a duct that is right-angles to the axis of the gas turbine. Thus, flow direction of flue gas must be changed by 90degree angle, and this is done by the exhaust plenum which is located inside duct close to the gas turbine outlet. The segment of the duct, which contains the exhaust plenum, is enclosed by another compartment called exhaust compartment Load compartment Gas turbine shaft and the turbine shaft are connected by a coupling called load coupling, and it is located in the load compartment Generator compartment The main item in the Generator compartment is the generator which is driven by the gas turbine. Since, this generator is driven by the gas turbine it is called gas turbine generator. 35

36 5.2 Air intake system Figure 5.5: Filter house Air, that is necessary for the gas turbine, is taken through a compartment, which is located outside the gas turbine building. This compartment is known as filter house, and there are two types of filter elements inside the filter house. Initially air comes through pre filter elements, and then goes through fine filter elements. Both pre and fine filters are 5microns in filtration rating. But some dust partial, which could not be blocked by the pre filter, are trapped in the fine filter. Differential pressures across both pre and fine filter elements are measured electronically, and recorded in the control room data base. When filter elements get choked, this differential pressure rises, and 60mmWC is the alarm limit for filter replacement. 5.3 Gas turbine auxiliary systems Gas turbine accessory gearbox Some equipment such as fuel pump, main oil pump, and atomizing air compressor are driven by the accessory gear box, and the input shaft of the gear box is coupled with the gas turbine shaft. Thus, accessory gear box takes input power from the gas turbine shaft Fuel injection system Fuel injection system is mainly consists of fuel stop valve, fuel pump, higher pressure fuel filter and the flow divider. All these equipment are in the accessory compartment. 36

37 Accessory Gearbox Fuel pump Figure 5.6: Gas turbine auxiliaries Fuel pump Fuel that comes from the gas turbine fuel oil skid goes through fuel oil stop valve, and pressure of the gas turbine fuel oil skid is around 4.2bar, and it is necessary to increase this pressure, since eventually fuel must be injected into combustion wrapper which is at around 15bar. Thus, outlet of the fuel stop valve is connected to a fuel pump, which is a gear pump, driven by the gas turbine accessory gear box. This fuel pump increases the fuel pressure. Fuel flow is controlled using a bypass valve that is connected across the inlet and the outlet lines of the fuel pump High pressure fuel filter Pressurized fuel, which is coming from the fuel pump, goes through a high pressure fuel filter. In this filter, contaminants such as wax and sand are retained. Filtration rating of the paper elements of this filter is 5 micron. This filter is not a duplex type (no standby filter); hence filter elements can not be replaced when the gas turbine is in operation Lubrication oil system Gas turbine and the generator are mounted on six bearings, and continuous lubrication oil supply is required when the plant is in operation. Not only for bearings but also for the accessory gear box, lubrication oil supply is an essentials requirement. Hence, lubrication oil system, which consists of tank, pumps, heat exchangers, filters, valves and various control devises, is located close to the gas turbine building. Apart from lubrication effect, it absorbed heat from above equipment Atomizing air system Discharge pressure of the atomizing air compressor is around 15bars, and the suction side is attached to the combustion wrapper. This is to reduce the differential pressure across the atomizing air compressor Gas turbine cooling water system It is required to absorbed heat from lubrication oil, atomizing air, turbine legs, and flame detectors in order maintain constant temperature. Thus, cooling water system is associated with gas turbine. Both cooling water and the lubrication oil flow through a shell & tube heat exchanger, in which, heat transfer taken place from lubrication oil to cooling water. Similarly, heat absorbs from atomizing air to cooling water using another heat exchanger. 37

38 Temperature of turbine support legs are maintained around 40 o c. This is to limit the expansion which could cause misalignment. Hence, turbine support legs are surrounded by jackets, through which, cooling water flow. Hence, heat transfers from turbine support legs to cooling water; and temperature of turbine support legs are maintained in an acceptable range. There are four flame detectors that are used to measure the intensity of the flames inside the gas turbine. These flame detectors are also cooled using cooling water. Fuel pump has inbuilt lubrication system which supplies lubrication oil to fuel pump. Finally the heated cooling water is cooled using an off-base cooling water module witch comprise of radiator (heat exchanger) bank, fans, pumps, surge tank, valves and instrumentations. Radiator modules are made of horizontal finned tubes that are fixed between two headers. Cooling air is blast across the radiator banks using six numbers of fans. It was necessary to calculate the heat rejection from the gas turbine cooling water for analysis. For this calculation, inlet/outlet cooling water temperatures of the radiator, and the flow rate of cooling water are required. There was no flow meter installed in the system to measure the cooling water flow rate. But, there were two local pressure gauges to measure the suction and, discharge pressures of the cooling water pump. When the plant is in operation reading are taken every 8 hours by plant operators and recorded in log sheets. Thus, cooling water flow rate was obtained form the pump characteristic curve using differential pressure across the pump. Similarly there was a local temperature gauge to measure the temperature of the outlet cooling water temperatures of the radiator, and readings have been recorded in log sheets. But the local temperature gauge, which was there to measure the inlet temperature, was not easily accessible. Hence, this temperature gauge has not been calibrated for longer perid. Thus, inlet water temperature measurement was carried out using a digital thermometer. 38

39 6 Heat recovery steam generator (HRSG) The exhaust flue gas temperature of the gas turbine is around 550 o c, and the flue gas flows through the a vertical HRSG. Energy that is available in the flue gas transfers into water, and production take place at the HRSG. As explain in the plant overview chapter the HRSG consist of three circuits (HP, LP and LLP circuit), through which, working fluid (water or steam) flows. 6.1 Low low pressure (LLP) circuit HRSG exhaust Flow station control Gas turbine exhaust Figure 6.2: LLP circuit Condensate pumped into the LLP circuit using condensate extraction pumps (CEP), and discharge line of these pumps terminates at a water storage tank called LLP drum. This drum acts as the storage, from which, water is taken for HP and LP circuits. LLP flow control station, that is located in between LLP drum and the CEP, is used to control the condensate flow. This control station consist of two control valves that are driven by pneumatic supply and two isolation valves that are driven by electrical motors. Steam lines that are coming from steam turbine, HP circuit, and LP circuit are connected to a common header; and this header terminates at LLP drum. This common header injects steam to the drum for preheating and dearation. When the plant is in normal operation, steam line, which is coming from steam turbine, fulfills the steam requirement for pre-heating and dearation. But, during the plant start-up, steams coming from the HP and LP circuits are used for pre-heating and dearation. This is because steam can t be extracted from the steam turbine during plant start-up. Blow-down line that connects the LLP drum and the IBD (intermittent blow down) tank is used to carryout drum blow-down in order to maintain chemical balance in the drum. 39

40 It is required to fill the LLP drum with fresh water before plant star-up if the shut down period is more than one week. Thus, separate filling line is also attached to LLP drum for initial filling. Two water pumps called LLP re-circulation pumps are associated with LLP drum in order to circulate water through the LLP evaporator which is a heat exchanger located in the HRSG. Hence, when drum water flow through the LLP evaporator, heat transfer take place from flue gas to water. 6.2 Low pressure (LP) circuit LP ejector Figure 6.4: LP circuit LLP drum provides the water to the LP circuit. Since, pressure of the LP circuit is higher than the LLP drum, feed water pumps are used to maintain water flow from LLP drum to LP circuit. One LP feed water pump is normally in operation and another one acts as the stand-by pump. Discharge line of the LP feed water pumps, which goes through the LP economizer module where heat absorption take place, connect with the flow control station. Outlet of the flow control station divides into two lines, and one goes to LP drum. The other line goes to the LP ejector. The LP ejector has two suction lines and one discharge line. As explain above one suction line connects with flow control station, and other suction line of the ejector connects with the LP drum. When, water that comes from flow control station goes through the ejector, water suck from the LP drum to ejector through the other line due to Venturi effect. Hence, inside the ejector, water that comes from both inlet ports mix and goes back to the LP drum through the LP evaporator module. Hence, water in the LP drum re-circulates through the LP evaporator where water evaporates due to heat absorption. During the plant start-up, flow through the ejector is not sufficient to produce the Venturi effect, and hence start-up pump is provided to maintain necessary flow rate during the start-up. During normal operation, bottom half of the drum filled with water, and top half contain saturated steam. A steam line, which is connected to the top of the LP drum, allows saturated steam to flow through the LP supper heater module where steam supper heating takes place. The steam that comes from LP supper heater goes to the steam turbine through two motor operated isolation valves which are parallel to each other. One valve is called main valve where as the other one is called integral by-pass valve. The main valve is 250mm in diameter and the integral by-pass valve is 25mm in di- 40

41 ameter. During the plant star-up down stream pressure of the above valves is around 1bar, and the up stream pressure is around 50bar. Due to this high differential pressure, large force acts on the stem of the main valve. But, the fore that acts on the integral by-pass valve is relatively small, since the diameter is only 25mm. Thus, during the plant start-up integral by-pass valve is opened, and water flows through the integral by-pass valve so that down stream pressure rises. When the down stream pressure increased, differential pressure drops; and the main steam valve is operated. Both supper-heated and saturated steam lines are provided vent valves that operate in case of a sudden pressure rise. Two blow-down lines are provided in order to maintain chemical balance of the drum, and water that comes from these blow-down lines goes to IBD (intermittent blow down) and CBD (continues blow down) tanks. Similar to LLP drum, separate filling line is provided for initial filling. 6.3 High pressure (HP) circuit Similar to LP circuit water flows to HP circuit from LLP drum through the HP feed water pumps. Figure 6.5: HP Boiler feed pump The discharge line of the HP feed water pumps goes to the HP economizer 1 module through the HP feed water main isolation valve which is associated with a integral by-pass valve. HP circuit consists of three economizers, one evaporator and two supper heaters. 41

42 Figure 6.6: HP circuit Feed water is injected to the outlet of the supper heater 1 module to control the temperature at the outlet of the supper heater 2 modules. Similar to LP circuit HP circuit also consists of feed water control station, ejector, startup pump, blow-down lines, and an initial filling line. 42

43 7 Steam turbine Figure 7.1: Steam turbine HP and LP superheated steam line, which come from HRSG, can carry condensate droplets especially during plant star-up. It is very important to maintain steam quality, before admit steam into steam turbine. Automated steam drip legs and manually operated steam drains are attached to the HP and LP steam headers. Condensate that collects in the drip legs of HP steam header, goes to a tank called HP flash tank. Similarly LP flash tank is incorporated in the system to collect condensate that comes from drip legs of LP steam header. Both flash tanks are connected with the condenser, and hence condensate that comes from drip legs will go back into the system through the condenser. To reduce NOx formation in the gas turbine, steam is injected into the combustion cans of the gas turbine, and the steam requirement is fulfilled by extracting steam from the turbine. But, when the plant is operated at low loads, pressure of the extraction steam is not sufficient. Hence, HP steam header provides steam for NOx reduction during part load operation. HP steam header divides into two branches, and there are two steam injection ports in the turbine. Each branch is associated with an isolation valve called HPSV (High Pressure Steam Valve) and another valve called HPCV (Higher Pressure Control Valve). In case of an emergency shutdown HPSV valve closed immediately to stop the steam supply the steam turbine. HPCV valve is used to control the steam flow rate, and hence it used for load controlling of the turbine. LP steam line does not have branches, and it is also associate with a LPSV (Low Pressure Steam Valve) and a (Low Pressure Control Valve) 43

44 Figure 7.2: HPCV All above mentioned HPSV, HPCV, LPSV, and LPCV valves are operated using hydraulic oil system. This hydraulic oil system consists of pumps, filtering system, and cooling system. Figure 7.3 Hydraulic oil system Similar to gas turbine, the steam turbine and the generator are mounted on six numbers of bearings, and hence lubrication oil system is provided for the bearings. This lubrication oil circulates through a cooler where heat absorption taken place. In case of a steam turbine shutdown, HPSV and LPSV valves closed immediately. This creates a sudden pressure rise in HP and LP lines. To overcome this problem, HP and LP by-pass valves are installed to bypass the steam turbine and to start dumping steam into condenser during a plant shutdown. At this condition both HP and LP superheated steam directly go into the condenser which is not designed for such situation. Hence, water is injected to the outlet of the bypass valves, in order to de-superheat HP and LP steam. This water line is taken from the CEP (condensate extraction pump) discharge. 44

45 Figure 7.4: Steam Turbine 45

46 8 Condensate system Figure 8.1 Condensate extraction pump Steam turbine is located on the condenser, in which, exhausted steam condensed due to heat transfer from steam to cooling water. Bottom part of the condenser is called hot-well, and the make up water requirement of the system is carried out by hot-well makeup pumps. This make up water is taken from the CST (Condensate storage tank). 46

47 Figure 8.2 Condensate system Major portion of the CEP discharge is taken for LLP drum, and small amount is used in vacuum pump, steam turbine gland steam system,, steam turbine exhaust hood spray, LP and HP by-pass valves, SWAS (Steam water analysis system) and in vacuum breaker valve. Steam turbine exhaust temperature is around 40 o c and this temperature should not increase to a very high level since, condenser is not designed for such situation. Hence, provision is provided to spray water to the steam turbine exhaust when its temperature increases beyond 92 o c. This is called steam turbine exhaust hood spray, and this spray water line is taken from CEP discharge. During the plant shut down condenser vacuum reduces from 0.074bar to atmospheric pressure. This is done by opening a valve called vacuum breaker valve which is installed in a line whose one end is connected to the condenser and the other end exposed to atmosphere. During normal operation one side of the vacuum breaker valve exposed to vacuum and the other end exposed to atmospheric pressure. Thus, water line that starts from CEP discharge provides water to the vacuum breaker valve in order to have a better sealing. One other important function that takes place in the condensate system is chemical dosing. This is to maintain the ph level of the water of the system. During normal operation condenser inside pressure is around 0.074bar, and before start-up this is at atmospheric pressure. Hence, during the start-up, air inside the condenser is sucked using vacuum pumps that are connected to condenser. 47

48 Figure 8.3 Vacuum pump Both these pumps are operated during the plant start-up to create the vacuum quickly, and once the inside pressure reach 0.074bar one pump is switch off while the other one remains in operation. It is not possible to have a perfect sealing everywhere in condensate system, and hence air can leaks into the system. Thus, one vacuum pump is in continues operation in order to remove air that leaks into the system. Vacuum pump is also requires seal water, and this requirement is fulfilled by CEP discharge water. 48

49 9 Cooling water system Figure 9.1 Cooling water system The main objective of the cooling water system is to provide cooling water for the condenser. This system consists of CW (cooling water) pumps, CT (cooling tower) fans, CWST (cooling water storage tank), and CT makeup pumps. To avoid algae generation in the cooling tower basin, calcium hypo-chloride is injected, and the concentration of this chemical in the cooling water is measured using a parameter called FRC (free residual chlorine). This FRC level is maintained around 0.5PPM in the CT basin. But when it is need to carry out CT blowdown, FRC must be reduced due to environmental concerns. Thus, sodium sulphate is injected into CT blow-down header to reduced FRC of the blow-down water. Figure 9.2 Cooling tower Cooling tower make-up pumps, which take water from the cooling water storage tank, pump water into CT basin to maintain the makeup water flow. Four numbers of cooling tower fan generates the necessary draft to cool the water, and when the plant is in full load operation all the fans must be used, and during part load operation one CT fan can be switch off. 49

50 Figure 9.3 Cooling water pump Three cooling water pumps are installed, and when plant is at full load, two numbers of pumps are sufficient to provide the necessary flow, and the other one is the standby pump. 50

51 10 SWAS (Steam water analysis system) Raw water contains number of minerals, and it is not possible to use this raw water for the steam cycle without treatments. This is because mineral could get deposited on the internal surfaces of the pipes that carry steam and water. Hence, raw water is treated in a water treatment plant to produced de-mineralized water. When concentration of mineral increase, electrical conductivity of water increases. Thus, mineral concentration is measured using a parameter called siemens per meter (S/m). 1 Siemen = Oms One other important parameter is the ph level of the fluid. Generally ph level of the fluid should be maintained around 9 in order to minimize the corrosion of the pipes. This is done by injecting various chemicals into the system, and the concentrations of these chemicals should also be measured frequently. Figure 10.1 SWAS panel Thus, water samples are taken from different location such as drums, supper heater outlets, CEP discharge, for testing purposes. Separate lines which are small in diameter are used to extract steam and water from above mentioned locations to collect samples. The fluid pressure is reduced using valves that are installed in these lines, and a cooler is used to condense steam at sample collection point to. This whole system is called steam water analysis system (SWAS), and offline and online measurements are carried out to monitor the quality of the steam and the water in the steam cycle. 51

52 11 System modelling 11.1 Combined cycle heat rate Both corrected and actual power outputs are available in the control room data base, and actual power output is used to calculate actual heat rate. This actual heat rate is divided by correction factors to obtain the corrected heat rate. Net actual power output of the plant Fuel temperature = TFUEL TFUELF = 1.8 TFUEL Fuel_densi ty API = [20] Fuel_densi ty Fuel reference temperature = TFUELREF Specific enthalpy of fuel at fuel reference temperature = H1 = PTNET (Acquired from control room data base) Specific enthalpy of fuel at combustion chamber inlet temperature = H2 [20] PATM1 = Actual combined cycle heat rate = HR_UNH m fuel (LHV + H2 H1) HR_UNH = & [20] PTNET 52

53 11.2 Gas turbine efficiency In order to analyze the gas turbine efficiency, a system boundary must be established. Since, there are several numbers of subsystems are involved, system boundary can be selected such away that, all the subsystems are inside the system boundary, so that it is not necessary to analyze each and every sub system in detail. Outer surface of the gas turbine compartments is selected as the boundary of the system. Fuel Steam Injection Heat removal through cooling water Convective and radiation losses 2 3 Electrical power Turbine 1 4 Air Heat removal from ventilation fan Flue gas & Super heated steam Figure 11.1 System boundary Fuel flow rate can be taken from the control room data base, and hence rate of heat input from the fuel, can be calculated using the calorific value Temperature and the pressure of steam that is injected into gas turbine can be taken from the control room data base, and thus heat input from the steam can be calculated. Heat removal through cooling water, Convective and radiation losses, and heat removal from ventilation fans can be calculated by local measurement Gas turbine cooling water system Cooling water pump discharge pressure = 4.11 bars Cooling water pump suction pressure = 0.3 bars Above values were taken from log sheets Differential pressure across the cooling water pump = 3.81 bars Differential pressure across the cooling water pump = m 53

54 Figure 11.2 Cooling water pump characteristic curve of Gas turbine Hence, flow rate through the gas turbine cooling water = 194 m 3 /hr Inlet cooling water temperatures of the radiator = 49.4 o c (measured) Outlet cooling water temperatures of the radiator = 42.2 o c (measured) Density of cooling water at above temperature = 998 kg/ m 3 Specific heat capacity of water = 4.18kJ/kg o c Heat rejected from gas turbine cooling water system= Heat_CW_GT Heat_CW_GT = Heat_CW_GT = 5.83 GJ/hr Heat_CW_GT = 1.62 MW ( ) Fuel flow rate Gas turbine fuel energy = m& [TPH] (Acquired from control room data base) fuel Higher heating value of fuel = HHV [MJ/kg] As explained in the fuel system chapter, samples are sent to an industrial laboratory to measure the calorific value of fuel. Since, fuel flow rate and heating value is known, rate of heat input to the gas turbine can be calculated. 54

55 Gas turbine electrical power Measured power output of gas turbine generator = PGTM [MW] (Acquired from control room data base) Frequency correction for gas turbine generator Voltage correction for gas turbine generator = PGTFL [MW] = PGTVL [MW] Power output after correcting for voltage and frequency PGTFL + PGTVL PGT = PGTM - [21] 1000 Electrical power output of gas turbine generator 3600 Power_gene rator_gt = PGT 100 = PGT [MW] = Power_generator_GT[GJ/h] Gas turbine steam injection Flow rate of steam that injected to gas turbine = WINJ1 [kg/s] (Acquired from control room data base) Flow rate of steam that injected to gas turbine WINJ = WINJ 3600 Pressure of steam that injected to gas turbine Temperature of steam that injected to gas turbine = WINJ [kg/h] = PINJ [bar] (Acquired from control room data base) = TINJ [ o c] (Acquired from control room data base) Since, temperature and the pressure of the steam that is injected into gas turbine is known, enthalpy can be calculated. Enthalpy of steam that injected to gas turbine = HINJ [kj/kg] Gas turbine air intake and exhaust Using ambient temperature TA, compressor inlet air enthalpy can be calculated. Ambient temperature = TA Compressor inlet air enthalpy = h 1_air Flue gas pressure at gas turbine outlet Flue gas temperature at gas turbine outlet Partial pressure of steam in flue gas = PFG [bar] (Acquired from control room data base) = TFG [bar] (Acquired from control room data base) = PSTMFG [bar] WINJ PSTMFG = (PATM + PFG) 1000 WINJ WGT_EXH Steam that was injected into the gas turbine become superheated and leaves with flue gas. The enthalpy of this superheated steam can be calculated since, pressure and temperature known. Enthalpy steam at gas turbine exhausts = HINJ_GT_EXT [kj/kg] 55

56 Air flow rate at compressor inlet = m& air [TPH] Flue gas flow rate at gas turbine outlet m & air + m& fuel = WGT_EXH Stoichiometric air to fuel ratio of fuel f = 14.6 [22] m& β = fuel [22] m& air β x = (1 + f ) [22] 1+ β Gas turbine exhausts temperature = WGT_EXH [TPH] = f = TTXM [ o c] (Acquired from control room data base) Since, ( TTXM ) is known, enthalpy can be obtained from table 14.1 Enthalpy of flue gas at turbine outlet = HFG Enthalpy of air at turbine outlet HFG = h 4_air + x DH4 [22] = h 4_air Gas turbine compartment convective losses Summation of length of the perimeters of gas turbine compartments = P_GT_compa rtment = 150 (Measured) Height of the GT compartment H_GT_compa rtment = 6.5(Measured) = H_GT_compartment Surface temperature of GT compartments = T_s T_s = 45 (Measured) P_GT_compartment Temperature of the boundary layer between atmosphere and the GT compartment surface = T_boundary T_s + TA T_boundary= 2 Volumetric expansion coefficient of boundary layer = VolExp Dynamic viscosity of air in the boundary layer Density of air in the boundary layer = ρ boundary Kinematic viscosity of air in the boundary layer = μ boundary = ν boundary 56

57 ν boundary μ = ρ boundary boundary Thermal conductivity of air in the boundary layer Prandtl number of air in the boundary layer = Pr = k boundary Since, temperature of the air in the boundary layer is known, volumetric expansion coefficient, dynamic viscosity, density, thermal conductivity, and Prandtl number can be obtained. g VolExp (T 3 S TA) H_GT_compartment Grashof number G rl = [23] 2 ν Rayleigh Number R = G Pr [24] Nusselt number ( Nu al rl L ) can be obtained from following table. Heat transfer coefficient for free convection h convection H_GT_compa rtment Nu L = [27] k boundary Convective heat loss from gas turbine compartments Figure 11.3 [27] = h convection = Heat_convection_GT[GJ/h] Heat_conve ction_gt = h convection P_GT_compa rtment H_GT_compa rtment (T S 3600 TA) Gas turbine ventilation system Ventilation system is incorporated to gas turbine in order to remove heat from all the compartments. Inside temperatures of the accessory and turbine compartments are very close to ambient condition; and hence, it is possible to assume that the heat dissipation from ventilation fans of above two compartments is negligible. But, inside temperatures of turbine, exhaust and load compartments are very high, and hence heat removal from ventilation fans of above three compartments must be estimated. 57

58 Turbine compartment ventilation system Volumetric air flow rate of turbine compartment ventilation fan = (measured) V& GT_turb_co mpartment = 21425[CFM] (measured) V & GT_turb_co mpartment Exhaust temperature of ventilation fan of turbine compartment = (measured) T air_gt_tur b_compartm ent Tair_GT_tur b_compartm ent = 154 [ o c] (measured) Density of exhaust air of ventilation fan of turbine compartment = ρ air_gt_tur b_compartm ent Since, ρ can be obtained T air_gt_tur b_compartm ent is known air_gt_tur b_compartm ent Specific heat capacity of exhaust air of ventilation fan of turbine compartment = cp air_gt_tur b_compartm ent Since, T air_gt_tur b_compartm ent is known cp air_gt_tur b_compartm ent can be obtained Heat rejection from turbine compartment ventilation system = Heat_Venti lation_gt_ turb_comp [GJ/h] Heat_Venti lation_gt_ turb_comp V& GT_turb_co mpartment = ρ (T air_gt_tur b_compartm ent air_gt_tur b_compartm ent 3600 TA) 6 10 cp air_gt_tur b_compartm ent Exhaust compartment ventilation system Volumetric air flow rate of exhaust compartment ventilation fan = (measured) V & [CFM] (measured) GT_exh_com partment = V & GT_exh_com partment Exhaust temperature of ventilation fan of exhaust compartment = (measured) T air_gt_exh _compartme nt Tair_GT_exh _compartme nt = 138 [ o c] (measured) Density of exhaust air of ventilation fan of exhaust compartment = ρ air_gt_exh _compartme nt Since, ρ can be obtained T air_gt_exh _compartme nt is known air_gt_exh _compartme nt Specific heat capacity of exhaust air of ventilation fan of exhaust compartment = cp air_gt_exh _compartme nt Since, T air_gt_exh _compartme nt is known cp air_gt_exh _compartme nt can be obtained Heat rejection from exhaust compartment ventilation system = Heat_Venti lation_gt_ exh_comp [GJ/h] Heat_Venti lation_gt_ exh_comp V& GT_exh_com partment = ρ (T air_gt_exh _compartme nt air_gt_exh _compartme nt 3600 TA) 6 10 cp air_gt_exh _compartme nt 58

59 Load compartment ventilation system Volumetric air flow rate of load compartment ventilation fan = (measured) V & [CFM] (measured) GT_Load_co mpartment = V & GT_Load_co mpartment Exhaust temperature of ventilation fan of load compartment = (measured) T air_gt_loa d_compartm ent Tair_GT_Loa d_compartm ent = 162 [ o c] (measured) Density of exhaust air of ventilation fan of load compartment = ρ air_gt_loa d_compartm ent Since, ρ can be obtained T air_gt_loa d_compartm ent is known air_gt_loa d_compartm ent Specific heat capacity of exhaust air of ventilation fan of load compartment = cp air_gt_loa d_compartm ent Since, T air_gt_loa d_compartm ent is known cp air_gt_loa d_compartm ent can be obtained Heat rejection from load compartment ventilation system = Heat_Venti lation_gt_ load_comp [GJ/h] Heat_Venti lation_gt_ load_comp V& GT_Load_co mpartment = ρ (T air_gt_loa d_compartm ent air_gt_loa d_compartm ent 3600 TA) 6 10 cp air_gt_loa d_compartm ent Heat_Venti lation_gt = Heat_Venti lation_gt_ turb_comp + Heat_Venti lation_gt_ load_comp + Heat_Venti lation_gt_ exh_comp Applying heat balance equation to system, following equation can be obtained, and m& air can be calculated. States of compressor inlet, compressor outlet, and gas turbine outlet can be obtained using the control room data base. Thus, gas turbine overall efficiency, compressor isentropic efficiency, turbine isentropic efficiency can be calculated. m& fuel HHV WINJ HINJ m& air h 1_air 1000 HFG HINJ_GT_EX T = WGT_EXH + WINJ + Power_gene rator_gt Heat_Convection_GT + Heat_Venti lation_gt + Heat_CW_GT Gas turbine efficiency Gas turbine heat rate = HRGTG [kj/kwh] WINJ HINJ m& fuel HHV + HRGTG = 1000 PGT Gas turbine efficiency = EFFGT 59

60 EFFGT = HRGTG Gas turbine compressor Gas turbine overall inlet filter differential pressure = DPGTFLT [mmwc] (Acquired from control room data base) Gas turbine overall inlet filter differential pressure DPGTFLT 9.81 DPGTFLT_bars = Compressor discharge pressure Compressor discharge temperature Isentropic exponent for compressor = DPGTFLT_bar [bar] = CPD [bar] (Acquired from control room data base) = CTD [ o c] (Acquired from control room data base) = K c Average of ambient temperature ( TA ) and the compressor discharge temperature, is considered as the temperature for compression in the compressor. Using this average temperature isentropic exponent for compressor ( K c ) can be obtained from following graph. Gas content for the compressor is zero. Figure 11.4 [22] Gas turbine exhausts temperature Turbine inlet temperature = TIT [ o c] = TTXM [ o c] (Acquired from control room data base) 60

61 CTD - TA K c 1 TA CPD + PATM ( ) = (( ) Kc 1) [22] η iso_compssor PATM DPGTFLT_bar ( ) 100 Isentropic exponent for turbine = K Average of turbine inlet temperature ( TIT ) and the turbine exhausts temperature ( TTXM), is considered as the temperature for expansion in the turbine. Using this average temperature isentropic exponent for turbine ( K) can be obtained from the above graph η iso_turbine 1 TIT - TTXM = (TIT + 273) (1 ) [22] 100 K 1 CPD + PATM ( ) ( ) K PFG + PATM Since, temperature of flue gas at turbine inlet is known enthalpy can be obtained from table Enthalpy of flue gas at turbine inlet = h 3 Enthalpy of air at turbine inlet h 3 = h 3_air + x DH3 Turbine power output = PT = h 3_air P T (m& air + m& fuel ) = (h 3 HFG) [MW] 3600 Compressor power output = PC Since, compressor discharge temperature CTD known, compressor outlet enthalpy ( h 2_air ) can be obtained from table P C m = & air (h 2_air h 1_air ) [MW] 3600 PGT = P T P C Heat recovery steam generator (HRSG) HP steam flow from HRSG to steam turbine = WHP [TPH] HP steam flow from HRSG to steam turbine WWHP = WHP 1000 HP superheated steam temperature = THP [oc] HP superheated steam pressure = WWHP [kg/h] = PHP [bar] - gauge pressure 61

62 HP superheated steam enthalpy HP drum pressure HP drum saturated water enthalpy = HHP [kj/kg] = PHPBD [bar] - gauge pressure = HHPSATWTR [kj/kg] HP drum blow down flow rate = WHPBD [TPH] HP drum blow down flow rate = WWHPBD [kg/h] WWHPBD = WHPBD 1000 LP steam flow from HRSG to steam turbine LP steam flow from HRSG to steam turbine WWLP = WLP 1000 LP superheated steam temperature = TLP [oc] LP superheated steam pressure LP superheated steam enthalpy LP drum pressure LP drum saturated water enthalpy = WLP [TPH] = WWLP [kg/h] = PLP [bar] - gauge pressure = HLP [kj/kg] = PLPSAT [bar] - gauge pressure = HLPSATWTR [kj/kg] LP drum blow down flow rate = WLPBD [TPH] LP drum blow down flow rate = WWLPBD [kg/h] WWLPBD = WLPBD 1000 Condensate pressure at LLP inlet = PFWLLP Condensate temperature at LLP inlet Condensate enthalpy at LLP inlet Condensate flow at LLP inlet Condensate flow at LLP inlet WWFWLLP = WFWLLP 1000 = TFWLLP [bar] - gauge pressure = HFWLLP[kJ/kg] = WFWLLP [TPH] = WWFWLLP[kg/h] Pegging steam flow to LLP drums from steam turbine Pegging steam flow to LLP drums from steam turbine WWSPEG = WSPEG 1000 Temperature of extracted pegging steam at LLP drums Pressure of extracted pegging steam at LLP drum Enthalpy of extracted pegging steam at LLP drums = WSPEG [TPH] = WWSPEG [kg/h] = TSPEG [oc] = PSPEG [bar] - gauge pressure = HSPEG [kj/kg] 62

63 Heat received to working fluid in the HRSG = Heat_Received_W_HRSG_Uncor [MJ/kg] WWHP HHP + WWHPBD HHPSATWTR + WWLP HLP + WWLPBD HLPSATWTR - HFWLLP WWFWLLP- HSPEG WWSPEG = 1000 HP feed water temperature = TFWHP [oc] HP feed water pressure = PFWHP [bar] Hp economizer 1 inlet feed water enthalpy = HFWHP [kj/kg] Pegging steam flow to LLP drum from LP drum Pegging steam flow to LLP drum from LP drum LP saturated steam enthalpy = HLPSATSTM WLPSAT = WLPSAT 1000 = WLPSAT [TPH] = WWLPSAT [kg/h] Heat addition to LLP drum from LP drum pegging steam = Heat_WF_LPSPEG [kj/h] Heat_WF_LP SPEG = WWLPSAT HLPSATSTM LP economizer inlet feed water temperature = TFWLP [oc] LP economizer inlet feed water pressure = PFWLP [bar] LP economizer inlet feed water enthalpy = HFWLP [kj/kg] HRSG HP de super heater spray flow HRSG HP de super heater spray flow LLP drum temperature = TLLPST [oc] LLP drum pressure = PLLPST [bar] LLP drum saturated water enthalpy = WHPDEH [TPH] = WWHPDEH [kg/h] = HLLPST [kj/kg] Heat input to HRSG from GT exhaust and from pegging steam = Heat_Input_HRSG [MJ/h] WGT_EXH HFG + WWSPEG (HSPEG HLPSATWTR) Heat_Input_HRSG = 1000 HRSG efficiency = HRSG_Effic iency Heat_Recei ved_w_hrsg _Uncor HRSG_Effic iency = 100 Heat_Input _HRSG 63

64 HRSG modules Figure 11.5 HRSG consists of 10 numbers of heat exchanger modules, and up stream and down stream flue gas temperatures of each module is measured. Two thermo couples (left and right) are located in between heat exchangers to measure these temperatures. Flue gas flow rate was calculated during the gas turbine analysis, and hence it is possible to calculate the heat dissipated from the flue gas at heat exchangers. Specific heat of flue gas is taken from following graph. Figure 11.6 [25] Temperature, pressure and flow rates of the fluid that flows inside the modules are also measured, and hence, it is possible to calculate the heat gained by the working fluid. 64

65 HP Economizer HP Economizer 3 outlet temperature = TFWHP_ECO3_OUT [oc] HP Economizer 3 outlet pressure = PFWHP_ECO3_OUT [bar] HP Economizer 3 outlet enthalpy = HFWHP_ECO3_OUT [kj/kg] Heat received to working fluid in HP Economizer = Heat_WF_HPECO [MW] WWFWHP HFWHP_ECO3 _OUT - HFWHP Heat_WF_HP ECO = HP Economizer 1 HP Economizer 1 outlet temperature = TFWHP_ECO1_OUT [oc] HP Economizer 1 outlet pressure = PFWHP_ECO1_OUT [bar] HP Economizer 1 outlet enthalpy = HFWHP_ECO1_OUT [kj/kg] Heat received to working fluid in HP Economizer 1 = Heat_WF_HPECO1[MW] WWFWHP HFWHP_ECO1 _OUT - HFWHP Heat_WF_HP ECO1 = Flue gas temperature after LP evaporator (right side probe) Flue gas temperature after HP economizer 1 (right side probe) Specific heat capacity of flue gas in the HP economizer 1 = TFG_AFT_LPEVP_2 = TFG_AFT_HPECO1_2 = Cp_HPECO1[MJ/kg] Heat transferred from flue gas at the HP economizer 1 = Heat_FG_HPECO1 WGT_EXH ( ) Heat_FG_HP ECO1 = Cp_HPECO1 (TFG_AFT_LPEVP_2 - TFG_AFT_HPECO1_2) 3600 MW] [ HP Economizer 3 Flue gas temperature after HP economizer 3 (left side probe) Flue gas temperature after HP economizer 3 (right side probe) Flue gas temperature after HP economizer 3 = TFG_AFT_HPECO3 TFG_AFT_HPECO3_1 + TFG_AFT_HPECO3_2 TFG_AFT_HPECO3 = 2 Specific heat capacity of flue gas in the HP economizer 3 = TFG_AFT_HPECO3_1 = TFG_AFT_HPECO3_2 = Cp_HPECO3[MJ/kg] WGT_EXH Heat_FG_HP ECO3 = 1000 Cp_HPECO3 (TFG_AFT_HPEVP - TFG_AFT_HPECO3)

66 HP Super heater 2 HP Super heater 2 inlet temperature = THP_SH2_IN[oc] HP Super heater 2 inlet enthalpy = HHP_SH2_IN [kj/kg] Heat received to working fluid in HP super heater 2 = Heat_WF_HP SH2 WWHP - WWHPDEH HHP - HHP_SH2_IN WWHPDEH HHP - HFWHP Heat_WF_HP SH2 = + [MW] Flue gas temperature after super heater 2 (left side probe) Flue gas temperature after super heater 2 (right side probe) Flue gas temperature after super heater 2 = TFG_AFT_HPSH2 TFG_AFT_HPSH2_1+ TFG_AFT_HPSH2_2 TFG_AFT_HPSH2 = 2 = TFG_AFT_HPSH2_1 = TFG_AFT_HPSH2_2 Specific heat capacity of flue gas in the HP super heater 2 = Cp_HPSH2 [MJ/kg] Heat transferred from flue gas at the HP super heater 2 = Heat_FG_HPSH2 WGT_EXH Heat_FG_HP SH2 = 1000 Cp_HPSH2 (TFG - TFG_AFT_HPSH2) HP Super heater 1 HP Super heater 1 outlet temperature = THP_SH1_OUT [oc] HP Super heater 1 outlet pressure = PHP_SH1_OUT [bar] PHP_SH1_OU T = PHP + PHPBD 2 HP Supper heater 1 output enthalpy = HHP_SH1_OUT [kj/kg] HP drum saturated steam enthalpy = HHPSATSTM [kj/kg] Heat received to working fluid in HP super heater 2 = Heat_WF_HPSH1[MW] WWHP - WWHPDEH Heat_WF_HP SH1 = 3600 Flue gas temperature after super heater 1 (left side probe) Flue gas temperature after super heater 1 (right side probe) HHP_SH1_OU T - HHPSATSTM 1000 Flue gas temperature after super heater 1 = TFG_AFT_HPSH1[oc] TFG_AFT_HPSH1_1+ TFG_AFT_HPSH1_2 TFG_AFT_HPSH1 = 2 = TFG_AFT_HPSH1_1[oc] = TFG_AFT_HPSH1_2[oc] Specific heat capacity of flue gas in the HP super heater 1 = Cp_HPSH1 [MJ/kg] 66

67 Heat transferred from flue gas at the HP super heater 1 = Heat_FG_HPSH1 [MW] WGT_EXH Heat_FG_HP SH1 = 1000 Cp_HPSH1 (TFG_AFT_HPSH2 - TFG_AFT_HPSH1) HP evaporator Heat received to working fluid in HP super heater 2 = Heat_WF_HPEVP[MW] Heat_WF_HP EVP = WWHP - WWHPDEH 3600 HHPSATSTM HFWHP_ECO3 _OUT 1000 Flue gas temperature after HP evaporator (left side probe) Flue gas temperature after HP evaporator (right side probe) Flue gas temperature after HP evaporator Specific heat capacity of flue gas in the HP evaporator Heat transferred from flue gas at the HP evaporator = TFG_AFT_HPEVP[oc] = TFG_AFT_HPEVP_1[oc] = TFG_AFT_HPEVP_2[oc] = Cp_HPEVP [MJ/kg] = Heat_FG_HPEVP [MW] WGT_EXH Heat_FG_HP EVP = 1000 Cp_HPEVP (TFG_AFT_HPSH1- TFG_AFT_HPEVP) LP economizer LP economizer outlet temperature = TFWLP_ECO_ OUT [oc] LP economizer outlet pressure LP economizer outlet enthalpy = PFWLP_ECO_ OUT[bar] = HFWLP_ECO_ OUT[KJ/kg] Heat received to working fluid in LP economizer = Heat_WF_LPECO [MW] WWFWLP HFWLP_ECO_ OUT HFWLP Heat_WF_LP ECO = Flue gas temperature after LP evaporator (left side probe) Flue gas temperature after LP economizer (left side probe) Specific heat capacity of flue gas in the LP economizer Heat transferred from flue gas at the LP economizer = TFG_AFT_LPEVP_1[oc] = TFG_AFT_LPECO_1[oc] = Cp_LPECO [MJ/kg] = Heat_FG_LPECO [MW] WGT_EXH Heat_FG_LP ECO = 1000 Cp_LPECO (TFG_AFT_LP EVP_1 - TFG_AFT_LP ECO_1) LP evaporator Heat received to working fluid in LP economizer = Heat_WF_LPEVP [MW] Heat_WF_LP EVP = WWLP 3600 HLPSATSTM HFWLP_ECO_ OUT 1000 Flue gas temperature after HP economizer 2 (left side probe) = TFG_AFT_HPECO2_1[oc] 67

68 Flue gas temperature after HP economizer 2 (right side probe) Flue gas temperature after HP economizer 2 = TFG_AFT_HPECO2[oc] = TFG_AFT_HPECO2_2[oc] TFG_AFT_HP ECO2 = TFG_AFT_HP ECO2_1 + TFG_AFT_HP ECO2_2 2 Flue gas temperature after HP economizer 2 = TFG_AFT_LPEVP[oc] TFG_AFT_LP EVP_1 + TFG_AFT_LP EVP_2 TFG_AFT_LP EVP = 2 Specific heat capacity of flue gas in the LP evaporator Heat transferred from flue gas at the LP evaporator = Cp_LPEVP [MJ/kg] = Heat_FG_LPEVP [MW] WGT_EXH Heat_FG_LP EVP = 1000 Cp_LPEVP (TFG_AFT_HPECO_2 - TFG_AFT_LP EVP) LP supper heater Heat received to working fluid in LP supper heater WWLP HLP HLPSATSTM Heat_WF_LP SH = LLP evaporator Flue gas temperature after LLP evaporator (probe 1) Flue gas temperature after LLP evaporator (probe 2) Flue gas temperature after LLP evaporator (probe 3) Flue gas temperature after LLP evaporator (probe 4) Flue gas temperature after LLP evaporator = Heat_WF_LPSH [MW] = TFG_AFT_LLPEVP_1[oc] = TFG_AFT_LLPEVP_2[oc] = TFG_AFT_LLPEVP_3[oc] = TFG_AFT_LLPEVP_4[oc] = TFG_AFT_LLPEVP TFG_AFT_LLPEVP_1+ TFG_AFT_LLPEVP_2+ TFG_AFT_LLPEVP_3+ TFG_AFT_LLPEVP_4 TFG_AFT_LLPEVP = 4 Specific heat capacity of flue gas in the LLP evaporator = Cp_LLPEVP[MJ/kg] Heat transferred from flue gas at the LLP evaporator = Heat_FG_LLPEVP [MW] TFG_AFT_LP ECO_1 TFG_AFT_HP ECO1_2 = WGT_EXH 1000 Cp_LLPEVP ( Flue gas temperature at HRSG inlet (probe 1) Flue gas temperature at HRSG inlet (probe 2) Flue gas temperature at HRSG inlet (probe 3) Flue gas temperature at HRSG inlet (probe 4) Flue gas temperature at HRSG inlet = TFG_INLET_ HRSG_1 = TFG_INLET_ HRSG_2 = TFG_INLET_ HRSG_3 = TFG_INLET_ HRSG_4 = TFG_INLET_ HRSG - TFG_AFT_LLPEVP) 68

69 TFG_INLET_HRSG_1+ TFG_INLET_HRSG_2 + TFG_INLET_HRSG_3 + TFG_INLET_HRSG_4 TFG_INLET_HRSG = 4 Specific heat capacity of flue gas in the HRSG = Cp_HRSG [MJ/kg] Heat transferred from flue gas at the entire HRSG = Heat_FG_HRSG [MW] WGT_EXH Heat_FG_HR SG = 1000 Cp_HRSG (TFG_INLET_ HRSG - TFG_AFT_LLPEVP) Steam Cycle HP steam temperature at the turbine inlet = THPST [oc] HP steam pressure at the turbine inlet = PHPST [bar] HP steam enthalpy at the turbine inlet = HHPST [kj/kg] LP steam temperature at the turbine inlet = TLPST [oc] LP steam pressure at the turbine inlet = PLPST [bar] LP steam enthalpy at the turbine inlet = HLPST [kj/kg] Flow rate of steam injected into gas turbine to reduced NOx generation = WEXT_NOX [kg/s] Flow rate of steam injected into gas turbine to reduced NOx generation = WWEXT_NOX [kg/h] WWEXT_NOX = WEXT_NOX 3600 Temperature of steam injected into gas turbine Pressure of steam injected into gas turbine Enthalpy of steam injected into gas turbine HP circuit feed water flow HP circuit feed water flow WWFWHP = WFWHP 1000 LP circuit feed water flow LP circuit feed water flow WWFWLP = WFWLP 1000 = WFWHP [TPH] = WWFWHP [kg/h] = WFWLP [TPH] = WWFWLP [kg/h] = TEXT_NOX [oc] = PEXT_NOX [oc] = HEXT_NOX [kj/kg] Heat input to steam cycle = Power_in_S T [MW] WWHP HHPST + WWLP HLPST + (WWHP + WWLP) HLLPST - (WWEXT_NOX HEXT_NOX + WWFWHP HFWHP + WWFWLP HFWLP + WWFWLLP HFWLLP) =

70 Steam turbine power output Steam cycle efficiency = PST [MW] = η_st_uncorr ected PST η _ST_Uncorr ected = 100 Power_in_S T Steam Turbine efficiency Steam turbine efficiency = EFFST [%] PST EFFST = WWHP HHPST + WWLP HLPST - (WINJ HINJ + WWSPEG HSPEG) Thermal power input to steam turbine from HP steam = Power_in_ST_HP [kj/h] Power_in_S T_HP = WWHP HHPST Thermal power input to steam turbine from LP steam = Power_in_ST_LP [kj/h] Power_in_S T_LP = WWLP HLPST Thermal power output from steam turbine due to steam extraction for gas turbine steam injection Power_out_ ST_INJ = WINJ HINJ [kj/h] Thermal power output from steam turbine due to steam extraction for LLP drum steam injection Power_out_ ST_SPEG = WWSPEG HSPEG [kj/h] Enthalpy of exhausted steam from steam turbine = HSTEXT [kj/kg] Power output from the steam turbine due to exhausted steam Power_out_ ST_EXT = (WWHP + WWLP - WINJ - WWSPEG) HSTEXT [kj/h] Total thermal power output from the steam turbine due to mass flow Perimeter of the steam turbine compartment P_ST_compa rtment = 75(measured) Height of the steam turbine compartment H_ST_compa rtment = 3.5 (measured) = P_ST_compartment[m] = H_ST_compartment [m] = Power_out_ ST [kj/h] Convective heat loss from steam turbine compartment = Power_out_ ST_convect ion 3600 Power_out_ ST_convect ion = h convection P_ST_compa rtment H_ST_compa rtment (TS TA) 3 10 Power_out_ ST = Power_out_ST_INJ + Power_out_ST_SPEG + Power_out_ST_EXT + Power_out_ST_convection Electrical power output from the steam turbine generator Power_out_ ST_Electrical = PST = Power_out_ST_Electrical [kj/h] Power_in_S T_HP + Power_in_ST_LP = Power_out_ST_Electrical + Power_out_ST 70

71 Dryness fraction of Steam turbine exhaust = x_stext Steam Turbine isentropic efficiency HP steam entropy at the steam turbine inlet = SHPST [kj/kg] LP steam entropy at the steam turbine inlet = SLPST [kj/kg] Exhaust steam entropy at the steam turbine outlet = SSTEXT [kj/kg] Exhaust steam enthalpy at the steam turbine outlet for isentropic expansion Isentropic efficiency of steam turbine = ST_Efficie ncy_iso HHPST - HSTEXT ST_Efficie ncy_iso = 100 HHPST - HSTEXT_ISO = HSTEXT_ISO [kj/kg] Condenser efficiency Cooling water temperature at condenser outlet = TCW_OUTLET_Condenser Cooling water temperature at condenser inlet = TCW_INLET_ Condenser Heat transfer into cooling water at the condenser Enthalpy of condensate at the condenser = HCondensate Heat_CW_Condenser = HSTEXT - HCondensate 1000 = Heat_CW_Condenser WWHP + WWLP - WINJ - WWSPEG 3600 Cooling water flow rate through the condenser = WCW WCW 4.2 Heat_CW_Co ndenser = 1000 (TCW_OUTLET_Condenser - TCW_INLET_Condenser)

72 12 Software architecture EES was selected for the development of the simulation model. Availability of the software, user friendliness, cost and the possibility to link the software with common packages such as Microsoft excel, were considered during the selection process Data processing At the end of the system modelling it was found that 89 plant operating parameters are required for the analysis, and all of these parameters can be obtained from a Microsoft Access database that is maintained in the plant control room. Plant parameters are measured using sensors, and the electrical signals are sent to control room, and the data base is automatically updated less than one second interval. This data base contains large numbers of plant parameters, and it is linked with the control desk which is used to control the power plant. Hence, this data base has very complex architecture. Therefore an excel file, into which, parameters are extracted from the data base, was prepared. This excel file is the interface between the Microsoft Access database that is maintained in the plant control room and EES model. This interface excel file is named as equipment performance data (EPD) file, and it contains several excel macro programs for data handling. First macro program extract data from the data base and calculate the average of data of every one minute interval. Data stored in the Microsoft access data base Macro program 1 Data (average every one minute interval) stored in the Microsoft excel As explain mentioned above, plant parameters are measured using sensors, and the electrical signals are sent to control room, and the data base is automatically updated less than one second interval. Some times during this communication process, some instantaneous data bytes losses due to communication failures. Loosing a single data byte will not make any significant effect for the analysis. But, in the data base #VALUE! text will be stored in the cell which corresponds to the particular instances. Theses texts are also extracted into the Microsoft excel EPD file. 72

73 These #VALUE! texts must be deleted before further data processing. It is not realistic to delete these texts manually as there are 1440 numbers of rows and 89 columns in a file which contain data of a day. Hence, another macro program was developed to locate all the #VALUE! texts in the excel EPD file and to remove them. Data with #VALUE! texts Macro program 2 Data without #VALUE! texts It is necessary to have further data processing, because data at this stage corresponds to a single day, and data updated during load changes are also present. Once a load change is done it will take an hour to stabilize the plant [25]. Hence, not only data updated during the load changes, but also data correspond to one hour just after load change must be removed form the file. Therefore another macro program was developed to remove the data correspond to unstable plant conditions. Data with unstable conditions Macro program 3 Data without unstable conditions 73

74 At this stage data can be fed into the EES model. There are 89 input and 52 outputs related to EES model. Thus, input feeding into the EES model and results extraction from the EES file are time consuming processes if there are done manually. Hence, another macro program was developed to feed data into the EES model. The execution of the EES model is also done by the same macro program, into which, results are extracted. Excel file Input data Macro program 3 Results EES model 12.2 EES program structure For the EES model following tables and windows were used. Lookup table 1 L Lookup up table 1 consists of 89 rows and refined data is loaded into this table by Macro program 3 of EPD excel file. Lookup table 2 h air Lookup table 3 Dh For the calculations it is necessary to have enthalpies of air at different temperatures, and these enthalpy values are stored in the Lookup table 2 & 3. Lookup table 4 k_gas Lookup table 5 cp Lookup table 6 Nusselt Similarly higher heating value of the fule is loaded into the Lookup up table 7 which consists of single row. Graph 11.4, Graph 11.3 and Graph 11.6 are tabulated in Lookup table 4, 5 & 6 respectively. Lookup table 7 HHV Lookup table 7 is used to store HHV value of fuel Parametric table 1 Table 1 Results are tabulated into parametric table 1 Equation window Equation window is used to write all the equations Diagram window Diagram window used visualize the resulats. 74

75 75

76 13 Analysis and results 13.1 Performances comparison 2007 vs EES system model was executed using present plant operating parameters, and the performances of each sub system were obtained. Similarly past data was used to obtain the past performances, and results were compared. Performance indicators were plotted against the ambient temperature due to the fact that ambient temperature is the most prominent uncontrollable operating parameter which affects performance. Performances of 10 th November 2011 and 20 th January 2007 were used for the comparison Heat rate Figure 13.1 It is possible to conclude that heat rate has improved from 2007 to Further, heat rate increases with raising ambient temperature. 76

77 Gas turbine efficiency Figure 13.2 It is possible to conclude that gas turbine efficiency has increased from 2007 to 2011, and gas turbine efficiency improves with decreasing ambient temperature Compressor isentropic efficiency of gas turbine Figure

78 When gas turbine is in operation, air flows into the compressor through the filter house. As explain in chapter 5.2 pre and fine filters are use to block dust particles that exist in air. But, some dust particles which are less than 5 microns in diameter penetrate filters. These due to the fact that both pre and fine filters are 5 microns in filtration rating. Dust particles, which could not be blocked at the filter house, enter into the compressor. Most of these particles escape the compressor, but few accumulate inside. This will increase the surface roughness of the compressor, and cause friction between air and internal surfaces. This affects the compressor isentropic efficiency. Thus periodic compressor cleaning methods are practiced to enhance the compressor performance. One method is called online water wash; and in this process, water is injected into the compressor inlet while the gas turbine is in operation. Injected water creates a mist that flows through the gas turbine with the air. This water vapour mist removes the dust particles in the compressor. Generally online water is carried out weekly, and water injection is continued for around 20 minutes. Some times, this method will not remove all the dust in the compressor. Thus, another method called offline water wash is done to remove the dust particles which could not be removed from online water wash. This method can only be done only when the gas turbine is not in operation. Thus, only few offline water washings are done annually. In this method water injection is done when the gas turbine rpm is around 600 and no mist formation take-place. Water droplets remove the dust. Initial water injection is done for 10 minutes followed by a detergent injection of 5 minutes. Subsequently several runs of water injections are done to remove the detergent. An offline water wash had been done on 20 th January 2007 just before test run, and hence higher isentropic efficiency was reported in But in 2011, the offline water wash, which was done prior to 10 th November, had been done on 23 rd June Hence, gas turbine isentropic efficiency on 10 th November 2011 is less than the value on 20 th January Turbine isentropic efficiency of gas turbine Figure

79 HRSG efficiency Figure 13.5 It is possible to conclude that HRSG efficiency has not change from 2007 to Steam cycle efficiency Figure

80 Significant drop in the steam cycle efficiency can be observed and the reason is explained below. Steam Cycle efficiency = Electrical Energy output from Net thermal energy input to ST generator steam cycle Net thermal energy input to steam cycle = Heat input to steam cycle from HRSG Heat extraction from ST due to NO x Heat extraction from ST due to NOx flow in 2007 is higher compared to Hence, reduction in NOx flow has increased the net thermal energy input to steam cycle. But electrical energy output from ST has not changed significantly, and thus steam cycle efficiency has dropped Heat transfer in the HP economizer 1 at full load Figure

81 Heat transfer in the HP super heater 1 at full load Figure Heat transfer in the HP super heater 2 at full load Figure

82 Heat transfer in the HP evaporator Figure Heat transfer in the LP economizer Figure

83 Heat transfer in the LP evaporator Figure Heat transfer in the LP supper heater Figure

84 Heat transfer in the LLP evaporator Figure In 2008 a modification was done on the HRSG, and the HP supper heater1 module was enclosed by steel bellows to reduce the heat transfer at this module. The objective of the modification is to reduce the desuper heater water spray that is carried out in between HP super heater 1 and 2. This is because rapid desuperheating creates thermal stresses in the system. Due to this modification, heat transfer at the HP supper heater 1 has decreased. This has increased HP supper heater 1 down stream flue gas temperature. Thus more heat transfer has taken place at the other modules. 84

85 13.2 Effects of NOx flow Steam that is injected into the gas turbine, in order to control NOx, affects plant performance. Thus, a trial was performed to analyze the affects of this steam flow. Currently steam flow of 4.8kg/s is maintained, and this flow rate was increased to 6.2kg/s, and the test run was started. The test run was done for 24 hours and then the system was brought back to original conditions. Plant operating parameters corresponds to test run were fed into the EES model and results were obtained Heat input to GT from NOx steam When steam is injected into the gas turbine, extra heat addition takes place other than the heat given from fuel. Hence, this extra heat increases with increasing NOx flow. Figure Heat addition to gas turbine, has increased significantly with increasing NOx flow. 85

86 Gas turbine power output Gas turbine power output increases when NOx steam flow increases. This is due to the extra heat addition and the increased mass flow. Figure Gas turbine power output has not increased proportionally with the increase in the heat addition. Hence, gas turbine efficiency has dropped slightly with increase in NOx flow Gas turbine efficiency 86

87 Figure Steam turbine power out Steam that is injected into gas turbine is extracted from the steam turbine. Hence, increase in NOx flow reduces the net heat input into the steam turbine. This causes drop in the steam turbine power output. Figure Net thermal energy input to steam cycle = Heat input to steam cycle from HRSG Heat extraction from ST due to NO x Steam Cycle efficiency = Electrical Energy output from Net thermal energy input to ST generator steam cycle Both numerator and the denominator of above equation have decreased with increasing NOx flow. But, steam turbine power output has not dropped proportionally. Thus, steam cycle efficiency has increased. 87

88 Steam cycle efficiency Figure Heat transfer in Condenser As explain above, steam cycle becomes less efficient when NOx flow reduces. This means, more heat transfer take place at the condenser. Figure