PDHengineer.com. Course M Industrial Gas Turbine Performance Engineering

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1 PDHengineer.com Course M Industrial Gas Turbine Performance Engineering This document is the course text. You may review this material at your leisure before or after you purchase the course. If you have not already purchased the course, you may do so now by returning to the course overview page located at: htm (Please be sure to capitalize and use dash as shown above.) Once the course has been purchased, you can easily return to the course overview, course document and quiz from PDHengineer s My Account menu. If you have any questions or concerns, remember you can contact us by using the Live Support Chat link located on any of our web pages, by at administrator@pdhengineer.com or by telephone tollfree at Thank you for choosing PDHengineer.com. PDHengineer.com, a service mark of Decatur Professional Development, LLC. M K9

2 Industrial Gas Turbine Performance Engineering By A.M.Y (Zak) Razak, Msc, PhD, Mem ASME Prepared for PDHEngineer.com (June 2007)

3 Table of Contents Part Introduction Thermodynamics of gas turbine cycles The first law of thermodynamics The second law of thermodynamics Entropy Steady flow energy equation Ideal simple cycle gas turbine Irreversibility Efficiency Isentropic efficiency Polytropic efficiency Temperature entropy diagram of a practical gas turbine cycle Thermal efficiency of a practical gas turbine cycle Component performance Axial compressor performance Compressor blade profile Stage characteristic Overall compressor characteristic Compressor surge Combustors Gas turbine combustors DLE/DLN Combustors Axial turbines Turbine blade profile Overall turbine characteristics Turbine creep life and cooling...29 Convective cooling...30 Impingement cooling...30 Transpiration cooling...30 Steam and mist cooling...30 Impact of turbine cooling technology on gas turbine performance Off-design behaviour of gas turbines Off design behaviour of the single shaft gas turbine Effects of ambient condition on engine performance Ambient temperature Ambient pressure Humidity effects Off-design behaviour of a two-shaft gas turbine operating with a free power turbine Effects of ambient condition on engine performance Ambient temperature Ambient pressure Humidity Performance deterioration Engine control systems Proportional, Integral and Derivative control

4 5.2 Low signal selection...49 Part Single shaft gas turbine performance and operation Effects of ambient temperature on gas turbine performance Effects of ambient pressure on gas turbine performance Simulating the effects of component deterioration on engine performance Compressor fouling Turbine damage Power augmentation Peak rating Power augmentation by water injection Turbine inlet cooling Simulation of engine control system performance PID Loop Proportional action Proportional and integral action Proportional, integral and derivative action Signal selection Optimising EGT Trips VIGV Control...84 Part Two-shaft gas turbine operating with a free power turbine performance and operation Effects of ambient temperature on gas turbine performance Effects of ambient pressure on gas turbine performance Simulating the effects of component deterioration on engine performance Compressor fouling Turbine damage Turbine hot end damage and blade rubs (power turbine) Power augmentation and power turbine performance Peak rating Maximum continuous rating Power augmentation at low ambient temperature Power augmentation by water injection Turbine inlet cooling Power turbine optimisation Power turbine performance Simulation of engine control system performance References

5 Part 1 4

6 1.0 Introduction Gas turbine performance is of paramount importance to both users and manufacturers, particularly in a deregulated market. It is effectively the end product that the users/operators of gas turbines purchase and the profitability of industries that employ them are becoming increasingly dependent on good engine performance. The performance of the gas turbine has increased considerably since its advent, which is credited to the Norwegian, Aegidius Elling, who in 1903 built the first successful gas turbine. Although the gas turbine was first developed for industrial use, it was not until its successful application in aircraft propulsion by Frank Whittle in Britain and Hans von Ohain and Max Hahn in Germany in the 1930s, that the importance of gas turbines was really appreciated. Today, the gas turbine is a formidable competitor as a prime mover achieving thermal efficiencies of over 40% in simple (open) cycle configuration and over 60% in combined cycle mode. It is a very popular choice for base load power generation and also widely used in mechanical drive applications, where the driven loads are process compressors and pump. Another application where the use of gas turbines is widespread is in naval propulsion. They were also actively considered in the 1960s for automotive applications. Advances made in gas turbine technology since then may see them being reconsidered for application in automotives. They are also considered for small scale and distributed power generation, particularly in conjunction with fuel cells, where overall thermal efficiencies of the order of 70% or higher are possible. Another industry that would consider them because of their superior performance is the nuclear power generation sector. This course is designed to give the reader a good insight into issues of gas turbine performance, both at design and at off-design conditions, where the engine operates for most of its life. We first start by describing the thermodynamics of the gas turbine cycle, know as the Joules cycle or the Brayton cycle. A brief description of gas turbine components such as axial compressors and turbines, and combustors are also included. The emphasis here is on component characteristics rather than the design of these components, as these characteristics plays a primary role in the performance behaviour of the gas turbine, particularly at off-design conditions. The off-design performance behaviour of gas turbines is described and the effects of ambient conditions, such as ambient pressure, temperature and, to a lesser degree humidity, are also discussed. For completeness, the impact of gas turbine performance deterioration and control are also included. The latter part of the course employs gas turbine simulators and revisits the material covered to reinforce understanding of gas turbine performance. These simulators are effectively virtual engines with respect to their performance, emissions, life usage and lifecycle cost. Therefore, we have also cover these aspects of gas turbine operation in the simulation section, including means of power augmentation and optimisation of gas turbine performance for a given site condition, stating their advantages and disadvantages. The reader is encouraged to repeat these simulations as this will help them in the quiz at the end of this course. Exercises using simulators differ from the more traditional numerical exercises, where simulator-based exercises give a holistic understanding, whereas numerical exercises give a somewhat narrower picture. 5

7 Although we have restricted our discussion to the single and two-shaft gas turbines operating with a free power turbine, these two configurations cover the vast majority of gas turbines operating in the field today. All figures included in this course have been taken from: 'Industrial Gas Turbines - Performance and Operability by A M Y Razak, 2007 and produced with the permission of: Woodhead Publishing Limited, Cambridge, UK, 6

8 2.0 Thermodynamics of gas turbine cycles Gas turbines are heat engines converting heat into useful work driving various loads such as electrical generators, pumps and compressors. The heat input is usually achieved by burning fuel in a combustion chamber also referred to as a combustor. Thus the performance analysis of a gas turbine is best achieved by applying the principles of thermodynamics. There are two laws of thermodynamics that concern us regarding gas turbine cycles. They are the first and the second laws of thermodynamics. There are several definitions of these laws, particularly the second law of thermodynamics and we shall use the following definitions. 2.1 The first law of thermodynamics The first law of thermodynamics states simply that energy cannot be created or destroyed but can only be converted from one type or form to another (e.g. conversion of heat energy into work). It is in effect a statement of the conversation of energy, where the work done by a thermodynamic system cannot exceed the heat input. For example, if we supply 10 MJ of heat into a thermodynamic system to produce work, then we can only produce up to 10MJ of work. 2.2 The second law of thermodynamics The second law imposes a further restriction on heat engines and states that not all the heat supplied to a heat engine can be converted into useful work. Some heat must be rejected. Therefore heat engines cannot have a thermal efficiency, which is the ratio of useful work to the net heat input, of 100% as implied by the first law of thermodynamics. The maximum thermal efficiency (ηthmax) of a heat engine is primarily determined by the minimum and maximum temperatures present in the cycle and given by the equation: T 2 η th max = T1 where T2 and T1 is the minimum and maximum cycle temperatures respectively. The maximum thermal efficiency is often referred to as the Carnot efficiency and no heat engine can exceed this efficiency. In fact, this is another statement of the second law of thermodynamics. For a heat engine to achieve the Carnot efficiency all the heat must be supplied at the maximum cycle temperature (T1) and all the heat rejected must occur at the minimum cycle temperature (T2). 2.3 Entropy The availability and accessibility of energy is important in producing work from a heat engine. The more accessible the energy is, the lower is its entropy. Conversely, the less available the energy is, the higher is its entropy. Entropy is a thermodynamic 7

9 property and is given the symbol S. The change in entropy during a thermodynamic process is defined as: S = dq T Steady flow energy equation The gas turbine cycle is a continuous flow process. Therefore the governing equation that satisfies the first law of thermodynamics is the steady flow energy equation. The steady flow energy equation may be described as: Q W = H Q represents the heat input into a steady flow thermodynamic system W represents the work done by the thermodynamic system H represents the change in the energy of the gas/fluid in the system as it has capacity to hold heat (specific heat). It is called the change in the stagnation or total enthalpy in the thermodynamic system. For an ideal gas the change in enthalpy can be represented by: H = m cp T where m is the mass flow rate of the fluid - usually air or products of combustion, cp is the specific heat of the gas at constant pressure and T is the total or stagnation temperature change in the thermodynamic system. We can therefore rewrite the steady flow energy equation 2.3 as: 2.5 Ideal simple cycle gas turbine Q W = m cp T The ideal gas turbine can be considered as a heat engine because it works in a cycle exchanging heat from a heat source and exhausting heat to a heat sink and producing work. The processes involved in the ideal gas turbine cycle, as shown in Figure 2.1, are: 1. Compression (Isentropic) 2. Heat addition (Constant pressure) 3. Expansion (Isentropic) 4. Heat rejection (Constant pressure) A thermodynamic process where the heat transfer is zero (i.e. adiabatic process) and reversible (i.e. entropy change is zero) is referred to as an isentropic process. From equation 2.5, the compressor and turbine work done is given by Wc = m cp ( T 2 T1) and Wt = m cp ( T3 T 4) respectively. 8

10 The heat transfers occur at constant pressure. Heat addition into as gas turbine combustor is a steady flow processes and the work transfer is zero. Therefore, from equation 2.5 the heat supplied Q = m cp ( T3 T 2) Figure 2.1 shows a schematic representation of a simple cycle gas turbine The gas turbine cycle is best illustrated on a temperature entropy diagram as shown in Figure 2.2, which shows the various thermodynamic processes. Figure 2.2 shows the representation of an ideal gas turbine cycle on a temperature entropy diagram Using the equations describing the compressor work, turbine work and heat input the thermal efficiency of an ideal simple cycle gas turbine is given by: 9

11 where γ 1 2 γ P c = and P 1 1 η th = c P 2 = Compressor pressure ratio. P1 The compressor temperature ratio for an isentropic process is given by: γ 1 γ T 2 P2 = where T2 and T1 is the compressor inlet and exit temperatures T1 P 1 respectively, and γ is the isentropic index (cp/cv), cv is the specific heat at constant volume. Equation 2.6 can be given by: T1 η th = T 2 Since T2 is less than T3, which is the maximum cycle temperature (also known as the turbine entry temperature), the ideal thermal efficiency of a gas turbine is less than the Carnot efficiency and therefore compliant with the second law of thermodynamics. The specific work Wnet, which is net work done by the gas turbine cycle per unit mass of working fluid, is given by: c 1 T3 Wnet = cpt1 c c T1 Figure 2.3 shows the effect of pressure ratio on the temperature-entropy diagram for an ideal gas turbine cycle when T3 is constant 10

12 From equation 2.6, we note that the ideal simple cycle gas turbine thermal efficiency increases with compressor pressure ratio. However, the thermal efficiency cannot increase indefinitely and the maximum thermal efficiency occurs when the compressor discharge temperature T2 tends to the turbine entry temperature T3. This condition would then correspond to the Carnot efficiency. From equation 2.8 the specific work will be zero. The other limiting case is when the compressor pressure ratio (P2/P1) tends to unity so the specific work and the thermal efficiency both becomes zero. Figure 2.3 illustrates these limiting cases on the temperature entropy diagram. The maximum compressor pressure ratio Rpmx occurs when T2 tends to T3 and is given by: γ T3 γ 1 Rpmx = T1 From the above, the maximum specific work occurs at some compressor pressure ratio between 1 and Rpmx and it can be shown that this condition occurs when the turbine exit temperature T4 equals the compressor discharge temperature T2. The compressor pressure at this maximum specific work condition is given by: γ 1 where C ( pr ) γ opt = opt T3 C opt = T1 The temperature entropy diagram for this maximum specific work case is also shown in Figure Irreversibility The above considered the compression and expansion processes as reversible and adiabatic (isentropic). These ideal processes never occur in practice. Although making such assumptions are useful in analysing ideal thermodynamic cycles, such as gas turbines, in practice we need to allow for irreversibilities. This give rise to the concept of efficiencies in the compression and expansion processes. We have stated in Section 2.3 that entropy is a measure of accessibility of energy. When compression and expansion processes are irreversible, entropy increases, therefore making energy unavailable. We illustrate this on the temperature entropy diagram (Figure 2.4), where we have shown a compression and expansion process when irreversibility s exists. (1 to 2 is an ideal compression process and 1 to 2 is the actual compression process. From 2 to 3 is the ideal expansion process and from 2 to 3 is the actual expansion process). During compression, the degradation of energy results in increased energy demand to maintain the required compressor pressure ratio, resulting from the increase in entropy during the compression process. Similarly, during expansion the degradation of energy results in a loss in work output from the expander, usually a turbine. Again, there is an increase in entropy to indicate the unavailability of this degraded energy in the expansion process. 11

13 Figure 2.4 showing the ideal and actual compression and expansion processes on the temperature-entropy diagram 2.7 Efficiency The degradation of energy discussed above results from the inefficiencies in the compression and expansion processes. For a compression process the efficiency is defined as the ratio of the ideal work to the actual work required to achieve a given pressure ratio. For an expansion process the efficiency is defined as the ratio of the actual work done to the ideal work done Isentropic efficiency Referring to Figure 2.4, the actual work absorbed (wca)during the compression process is: wca = cp ( T 2 T1) where cp is the mean specific heat at constant pressure, T1 and T2 is the compressor inlet and exit temperatures respectively. The ideal work required for compression (wci) is: wci = cp( T 2' T1) where T2 is the ideal compressor exit temperature. 12

14 For an isentropic process, γ 1 P2 γ Where c = P 1 T 2' = c T1 The isentropic efficiency is now given by: T1( c 1) η c =.2.11 T 2 T1 Similarly it can be that the isentropic efficiency of the turbine is given by: For an isentropic process T3'/T2= 1/c, then Polytropic efficiency T 2 T3 η t = T c The isentropic efficiency only considers the start and end states of the compression and expansion processes and pays no attention to the actual paths the compression and expansion processes take. Since the work is not a thermodynamic property and depends on the actual path, the polytropic analysis endeavours to account for the path taken during the compression and expansion processes in determining the actual work. In a polytropic process the compression or expansion process takes place in small steps (infinitesimally small steps). Calculating the work for the polytropic process involves the summation of the work for each step. To calculate the work for each infinitesimal step we use the isentropic analysis discussed in above. It can be shown [Saravanamutoo, Rogers and Cohen 2001]: For a compression process, the polytropic efficiency is given by: γ 1 P2 γ ln P1 η p = T 2 ln T1 and for an expansion process the polytropic efficiency is given by: 13

15 T3 ln 2 η T p = γ 1 P3 γ ln P2 γ = cp/cv, ratio of specific heats, also known as the isentropic index. 2.8 Temperature entropy diagram of a practical gas turbine cycle When irreversibilities are present in the compression and expansion processes, the increase in entropy results in an increased discharge temperature as discussed above. Furthermore, the heat addition in a practical gas turbine does not occur at constant pressure. There are two main sources of combustion pressure loss arising from the frictional loss and that due to the addition of heat resulting in the increase in gas temperature. The latter is often referred to as the fundamental loss and is associated with the decrease in gas density due to the increase in gas temperature. The inlet and exhaust systems also incur pressure losses due to frictional effects present in these systems. The temperature entropy diagram describing the above effects is shown in Figure 2.5. Figure 2.5 shows the temperature - entropy diagram for a practical gas turbine cycle 2.9 Thermal efficiency of a practical gas turbine cycle For a given working fluid, we have stated that the ideal simple cycle gas turbine thermal efficiency is only dependent on the compressor pressure ratio. It should be 14

16 evident from the above that when irreversibility and pressure losses are considered, the thermal efficiency of a practical gas turbine is also dependent on the maximum to minimum cycle temperature ratio. The thermal efficiency of a practical gas turbine can be given by: t 1 η th = 1 ct 2.15 t cc where t = T3/T1 (maximum to minimum cycle temperature ratio) 1 ct =, where rpt is the turbine expansion ratio, γ = cp/cp, and ηpt is the ( γ 1) ηpt ( rpt) γ turbine polytropic efficiency given by equation γ 1 cc = ( rpc) γηpc, where rpc is compressor pressure ratio, γ = cp/cv, ηpc is the compressor polytropic efficiency given by equation For a given compressor and turbine pressure ratio, and γ, the thermal efficiency described by equation 2.15 increase as t increases. Thus, for a given compressor inlet temperature (T1), the thermal efficiency increases with T3, which is the turbine entry temperature. This effect is illustrated in Figure 2.6. It should be noted that the thermal efficiency of a practical gas turbine does not continuously increase with compressor pressure. For a given turbine entry temperature, a continuous increase in compressor pressure ratio will result in a decrease in thermal efficiency at high compressor pressure ratios. This differs from the ideal simple cycle gas turbine thermal efficiency, which continuously increases with compressor pressure also shown in Figure 2.19 for comparison. Figure 2.19 shows the simple cycle gas turbine thermal efficiency varying with pressure ratio 15

17 Current simple cycle gas turbine designs operate at turbine entry temperatures of the order of 1850K and compressor pressure ratios of the order of 40. The thermal efficiency of such advanced simple cycle gas turbines exceeds 42%. 3.0 Component performance The various thermodynamic processes present in a gas turbine cycle are achieved using devices such as compressor, combustors and expanders. In gas turbines, the compression and expansion process is usually achieved using axial compressor and axial turbines respectively. The combustion process is achieved by burning fuel in a combustor consisting of various stages. In this section we shall briefly discuss the performance of such devices. 3.1 Axial compressor performance An axial compressor consists of a series of stages where each stage is made up of a rotor and a stator as illustrated in Figure 3.1. Work is entered via the rotor to increase the kinetic energy of the working fluid, which is usually air. The fluid is then diffused in the stage to increase its static pressure. The diffusion can occur in both the rotor and stator. When all the diffusion occurs in the rotor, the stage is said to be of 100% reaction design. When all the diffusion occurs in the stator, the stage is considered to be of impulse design (also known as a 0% reaction). Usually the compressor stages are designed for 50% reaction resulting in the diffusion being divided equally between the rotor and stator as this condition results in the best efficiency of the compressor stage. Figure 3.1 an axial compressor showing the first and last stages Compressor blade profile A typical compressor blade profile is shown in Figure 3.2. The air enters the rotor and is deflected through a suitable angle to prevent the air from stalling as it passes over the rotor and stator. The figure also shows the velocity triangles at rotor and stator inlets. The vectors V1 and C1 are the relative and actual velocity into the rotor at incident angles β1 and α1 respectively. Similarly, vectors V2 and C2 at exit angles β2 and α2 are the relative and actual velocities at exit from the rotor. The vector U is the blade velocity and Ca is the axial air velocity and is proportional to the compressor 16

18 inlet air flow rate. The profile of the rotor and stator blades influences these velocity triangles, and in turn thus affects the performance of the compressor. The annulus of the compressor shown in Figure 3.1 is usually convergent so that the axial velocity remains constant as the pressure increases along the annulus. In this case the velocity triangles can be superimposed as shown in Figure 3.3 Figure 3.2 shows the blade profile of a compressor stage Figure 3.3 shows the combined velocity triangle for the rotor and stator The symmetry of the superimposed velocity triangles implies that the stage reaction is 50%. Since there is no change in the axial velocity, the work input occurs due to the change in tangential velocity, Vw, which is the rate of change of momentum per unit mass flow rate. The work done per unit mass flow rate is now U Vw and will also 17

19 correspond to the specific enthalpy change. This can be related to the stage pressure ratio as follows: The temperature rise in the stage is determined by the stage stagnation enthalpy rise H. Assuming air as a perfect gas the stage temperature rise T = H / cp, where cp is the specific heat of air at constant pressure. The stage pressure ratio Rs can be calculated from the expression: Rs = 1+ η γ 1 T γ s T Where ηs is the stage isentropic efficiency, which closely approximates the compressor polytropic efficiency γ = cp/cv (ratio of specific heats) and T is the stage stagnation inlet temperature. Repeating the above calculation for all the compressor stages using the exit conditions (pressures and temperatures) of the previous stage as inlet to the current stage the overall pressure ratio of the compressor can be determined Stage characteristic The parameters H, U and Va can be grouped into non-dimensional parameters known as the stage loading and flow coefficients. These are defined as: H Va Ψ = Stage loading coefficient, Φ = Flow coefficient 2 U U Figure 3.4 shows an actual stage characteristic 18

20 In terms of air angles (β1 and β2): H U 2 = Ca (tan( β1) tan( β 2)) U Stage characteristics are normally represented by plotting the stage loading coefficient against the flow coefficient as shown in Figure 3.4. At low flows the high angle of incidence results in stalling whereas at high flows the inlet Mach number approaches unity and the inlet flow chokes causing the characteristic to become vertical. It should be noted that the velocity triangles change as we approach stall and choke conditions. The rotor incident increases as we approach stall and decreases as we approach the choke condition Overall compressor characteristic Figure 3.5 showing a typical compressor characteristic Using the stage characteristics discussed above, the overall compressor characteristic representing the effects of all the stages can be generated. This is particularly useful when predicting the performance of gas turbines at different operating conditions. A method known as stage stacking is used to generate the overall characteristic and is similar to that discussed when determining the overall pressure ratio of the compressor. However, we need to allow for the change in flow coefficient as well as stage loading coefficient for each compressor stage. Overall compressor characteristics are also produced on a non-dimensional basis but instead of using stage loading and flow coefficients we use non-dimensional flow, speed and pressure ratio. In fact, the non-dimensional flow and speed are Mach 19

21 numbers based on axial flow and rotational speed of the compressor respectively. The non-dimensional flow and speed are defined as: W RT / γ ND and 2 D P γrt respectively. Where W is the mass flow rate, R is the gas constant at compressor inlet, T is the compressor inlet stagnation temperature, P is the stagnation pressure at compressor inlet, γ is the ratio of specific heats (cp/cv) at compressor inlet, D is the reference diameter. For a give compressor, the reference diameter D is constant and often omitted from these non-dimensional groups. The efficiency of the compressor at different operating conditions is also drawn on the compressor characteristic and is shown as contours of constant isentropic efficiencies. Figure 3.5 shows an example of a typical compressor characteristic. The compressor speed lines become vertical at the high speed end of the characteristic due to the front or LP stages of the compressor choking as the axial velocity increases as the airflow rate increases. At low speed operation, the HP stages of the compressor chokes, due to lower pressure in these stages resulting in high axial velocities Compressor surge There is a region, known as the surge region, on the compressor characteristic where operation is not possible and this region is located to the left of the surge line, as shown in Figure 3.5. Surge is a violent aerodynamic instability which can easily destroy the compressor. It is associated with stalling, however, if the compressor efficiency falls rapidly then the discharge pressure cannot be maintained and the flow reverses. As the discharge plenum downstream of the compressor empties, due to flow reversal, normal compression will resume. However, as long as the condition that gave rise to the stall and surge prevails the flow will break down again and reverse. This flow reversal can generate pressure waves at frequencies close to the resonance frequency of the compressor and may destroy the compressor. Another feature of compressor surge is high compressor discharge temperatures. This is due to the recompression of what is effectively compressor discharge air, already at a high temperature, resulting in very high discharge temperatures during surge. We have stated that compressor surge is associated with compressor stall. We have also stated that at high speeds the choking of the compressor s LP stages occurs. This forces the HP stages to stall and surge at high speed operation is thought to be associated with stalled HP stages. This imposes an upper limit on the compressor pressure ratio, which is currently at about 20:1. Similarly, at low speeds, the operation of the compressor results in the stalling of the LP stages of the compressor and this is due to the choking of the HP stages. Thus, compressor surge at low speeds is associated with stalling of the LP stages of the compressor. 3.2 Combustors 20

22 The next major engine component is the combustor and it is responsible for achieving the required heat input. This is produced by burning hydro-carbon fuels such as diesel and kerosene, which are liquid fuels, or gaseous fuels such as natural gas. The combustion process requires the fuel to be heated to sufficiently high temperatures where species or radicals, such as CH 3, OH, etc are produced. For example, a methane molecule CH 4 would produce a methyl radical (CH 3 ) and a hydrogen ion (H+ ). In the absence of oxygen these radicals will revert to their original state at low temperatures. However, in the presence of oxygen, they will readily oxides producing heat and products of combustion. The products of combustion are mainly carbon dioxide and water vapour (steam). There are also small amounts of pollutants such as un-burnt hydrocarbons (UHC) and carbon monoxide (CO), also formed during combustion. The presence of these pollutants is due to combustion inefficiencies. However, gas turbine combustion efficiencies are very high and are of the order of 99.5%. Thus emissions of UHC and CO are small. Another pollutant produced during combustion is NOx (Oxides of nitrogen). NOx is formed primarily due to the oxidation of nitrogen in the combustion air. Combustion parameters such as pressures, temperatures, air-fuel ratio and fuel type (e.g. liquid or gaseous) have a strong influence in the formation of these pollutants. For example, high combustion pressure and temperature reduce the formation of UHC and CO emissions but increase the formation of NOx emissions. Liquid fuels normally produce higher levels of emissions compared with gaseous fuels. Another product of combustion is carbon dioxide (CO 2 ). Unlike NOx, CO and UHC, which are poisonous, CO 2 emissions are thought to be mainly responsible for global warming. The amount of CO 2 emissions are determined primarily by two factors. These correspond to the thermal efficiency of the gas turbine and the amount of carbon present in the fuel, namely the carbon to hydrogen ratio. The higher the thermal efficiency is, the lower is the fuel flow required to generate a given level of power output, thus high thermal efficiency reduces CO 2 emissions. Current design of gas turbines can achieve thermal efficiencies above 40%. When used as part of a combined cycle plant, thermal efficiencies of the order of 60% are possible. The higher the carbon-hydrogen ratio, greater is the content of carbon in the fuel and this will therefore increase the emissions of CO 2. We illustrate this with a simple example. The chemical reaction of methane with oxygen is: CH 4 + 2O > CO 2 + 2H 2 O Therefore one molecule (mole) of CH 4 (whose carbon-hydrogen ratio is about 0.25) will produce one mole of CO 2. Since the molecular (mole) weight of CH 4 and CO 2 are approximately 16 and 44, therefore 1 kg of CH 4 will produce 2.75 kg of CO 2. Similarly, kerosene can be represented by the hydrocarbon C 12 H 24 : C 12 H O > 12CO H 2 O The mole weight of kerosene is about 168. Therefore 1 kg of kerosene (whose carbon - hydrogen ratio is 0.5) will produce about kg of CO 2. But the lower heating value (LHV) of CH 4 is about 50MJ/kg and the corresponding value for kerosene is 21

23 about 43MJ/kg. Therefore, for a given heat input the fuel flow of kerosene is approximately 50/43 = times greater than when using CH 4. Thus the CO 2 emissions arising from kerosene is about (3.143/2.75 x 1.163) 1.33 times higher than when using methane Gas turbine combustors The above section has briefly described the combustion process in general. We now discuss how these combustion processes are implemented in a gas turbine combustor. There are two types of gas turbine combustors and they are referred to as diffusion and dry low emissions (DLE) or dry low NOx (DLN) combustors. The diffusion flame type is the original type of combustion system and many are still in use today. Diffusion flame type combustors consist of four distinct regions and correspond to the diffuser, primary zone, intermediate zone and the dilution zone as shown in Figure 3.6. Figure 3.6 shows a typical diffusion type gas turbine combustor The diffuser reduces the combustion air velocity sufficiently so that combustion can be sustained. The primary zone is where the fuel is burnt and it is the major heat release zone in the combustor. The primary and secondary air holes provide the combustion air for the primary zone. The air pattern through these holes form strong vortices in the primary zone and the fuel is injected through the fuel nozzle into the eye of these vortices. Igniters are used to initiate the combustion process. The products of combustion present in the primary zone may contain high levels of UHC and CO. The intermediate zone adds control amounts of air into the products of combustion leaving the primary so that UHC and CO are oxidised to CO 2 and H 2 O. The temperature of the gases leaving the intermediate zone may be too high for the turbine components (blades), which are located downstream of the combustor. The function of the dilution zone is to add more air into the gases leaving the dilution zone so reducing the gas temperature to that suitable for the turbine. Current practice allows turbine entry temperatures (TET) of the order of 1800K. 22

24 The distribution of the air and fuel in the primary zone (diffusion type combustor) is largely stratified and therefore there are regions where the air-fuel ratio is approximately stoichiometric (i.e. theoretically correct for complete combustion). The gas temperature in this region can be as high as 2500K, resulting in high NOx emissions. However, regions where the air fuel ratio is high (i.e. lean) the NOx will reduce. Since NOx increases exponentially with combustion temperature, high levels of NOx emissions result. Current legislation requires NOx levels to be reduced drastically (less than 25 ppmv). One method used widely in conventional (diffusion) combustion systems is water injection. Water has a high specific heat and therefore acts as a good heat sink. Thus injecting controlled amounts of water into the primary zone will result in reducing the primary zone flame temperature thereby reducing NOx emissions significantly. However, the decrease in primary zone flame temperature will increase CO emissions and the amount of water injection for NOx suppression would be limited by CO emissions. The additional heat required to heat the water to steam would result in an increase in fuel flow and this will reduce the thermal efficiency. In spite of these drawbacks water injection is an effective mean of NOx control. Another method of NOx suppression is to use steam injection. It is less effective than water injection in suppressing NOx but improves the thermal efficiency of the gas turbine. Both methods are currently in use DLE/DLN Combustors An alternative method of reducing NOx is to ensure that the primary zone air fuel ratio is lean. This results in reduced combustion temperature and thus NOx emissions. To ensure low NOx emissions, the primary zone temperature should not be allowed to exceed 1800K. Also, the air and fuel should be thoroughly mixed to ensure that the primary zone is uniformly lean (i.e. the equivalence ratio, which is the fuel air ratio to the stoichiometric fuel air ratio, is never greater than 0.6 in any part of the primary zone). The problem with DLE combustors occurs when we reduce the load of the engine. This results in the fuel-air ratio decreasing (or the air-fuel ratio increasing) and thus the equivalence ratio approaches the weak extinction limit. If the weak extinction limit is exceeded, the combustor will result in flameout and the engine will shutdown, which should be avoided. Ideally, we would like to switch more of the primary zone air into the dilution zone as the load is reduced, such that the equivalence ratio in the primary zone remains constant at low loads. Such variable geometry combustion systems have proved to be somewhat unreliable and this led to the development of stage combustion. However, some manufacturers have developed DLE combustion systems using effectively a single stage. These types of combustion systems require an overboard bleed to maintain the equivalence ratio at low loads but incur a thermal efficiency penalty due the overboard bleed. In staged combustion, we switch the fuel flow from one stage to another with engine load such that the equivalence ratio is in the region where NOx emissions are low. Two types of stage combustion systems have been developed and they correspond to the series and parallel types, as shown schematically in Figure 3.6a. With parallel staged combustion two chambers are positioned side by side and the fuel flow is reduced in one and then the other as the load is reduced. Such systems may also require an overboard bleed in order to prevent the weak extinction limit from being exceeded at low engine loads. However, the bleed would be smaller. In the series staged system, one chamber is positioned in front of the other. The advantage of series 23

25 staging is that the second stage has a very low weak extinction limit and would not require overboard bleeds. This is due to the high primary or first stage exit gas temperature. However, as the second stage combustion temperature decreases with engine load, the CO emissions may rise. Therefore, bleeds may be necessary to maintain the combustion temperature, thus controlling the CO emissions. But the bleeds needed would be smaller than for parallel staged combustion systems. Figure 3.6a shows a schematic representation of staged combustion Another problem with DLE combustion systems is that the heat release occurs at narrow acoustic frequencies. These frequencies may approach the resonance frequency of the combustion system resulting in serious damage to the combustion system and engine. This is largely overcome by changing the mass of the combustion system such that the natural frequency of the combustor is shifted away sufficiently to prevent resonance. 3.3 Axial turbines Figure 3.7 shows a typical axial turbine consisting of two stages As we found with axial compressors, which absorb power to increase the pressure and temperature ratio, turbines achieve the opposite. They produce power while reducing 24

26 the pressure and temperature ratio. A turbine stage consists of a stator (also knows as a nozzle guide vane or nozzle) and a rotor (sometimes referred to as a bucket), as shown in Figure 3.7. The expansion of the gas (usually products of combustion) can occur in the nozzle or rotor. When all the static pressure drop occurs in the nozzle the reaction of the stage is said to be zero (impulse turbine). When the entire static pressure drop occurs in the rotor the reaction of the stage is said to be 100%. Usually the stage reaction is 50% indicating that the expansion of the gases in the turbine stage is equally divided between the nozzle and the rotor Turbine blade profile A typical turbine blade profile is shown in Figure 3.8. The gas enters the nozzle and is deflected through a suitable angle onto the rotor to minimises losses and hence increase the efficiency of the turbine. The vectors C1 and V1 are the actual and relative velocities at entry to the rotor respectively. The actual and relative angles of incidence into the rotor are α1 and β1 respectively. Similarly, the vectors C2 and V2 are the actual and relative velocities at exit from the turbine rotor with exit angles corresponding to angle α2 and β2. The vector U corresponds to the blade velocity at some point on the rotor blade and this is usually at blade mid height. The velocity vector Va is the axial velocity along the turbine and is proportional to the gas flow rate. The profile of the turbine blades influence the velocity triangles and hence the performance of the turbine stage. Figure 3.8 shows typical turbine rotor and stator blade profiles The annulus of the turbine, as shown in Figure 3.7 is divergent in such a manner that the axial velocity remains constant. The divergence of the annulus is necessary to accommodate the decrease in pressure/density due to the expansion nature of turbines. 25

27 In this case we can superimpose the two velocity triangles as shown in Figure 3.9. The symmetry of the superimposed velocity triangles implies the reaction is 50% and usually results in the maximum efficiency of the turbine. As there is no change in the axial velocity (Va) at the inlet and exit of the turbine stage, for a given blade velocity (U), all the work done by the stage is achieved by the change tangential velocity ( Vw). The work done per unit mass flow rate of gas through the turbine stage is U Vw and will correspond to the stage specific enthalpy change H. Figure 3.9 shows superimposed velocity triangles for a turbine stage The stage specific enthalpy change can be used to determine the stage pressure ratio as follow: The temperature change in the stage is determined by the stage specific enthalpy drop H. Assuming products of combustion acts as a perfect gas, the stage temperature drop T = H / cp, where cp is the specific heat of air at constant pressure. The stage pressure ratio Rs can be calculated from the expression: Rs = 1 T 1 Tη s γ γ Where η s is the stage isentropic efficiency, which is approximately equal to the polytropic efficiency. γ = cp/cv (ratio of specific heats) and T is the stage inlet temperature (total or stagnation). As discussed with compressors, the work done and the resultant power produced by the turbine occurs in the rotor rather than the stator. The latter is used to deflect the gases onto the rotor to obtain the necessary reaction and efficiency, and this is dependent on the turbine blade profile, as shown in Figure 3.8. The power produced by the turbine (rotors) is used to drive the compressor and the load. Unlike compressors, where the pressure gradient is negative (i.e. the pressure increases along the compressor, which can result in stall and surge), in turbines the pressure decreases as we progress along the turbine. Therefore, surge does not occur in turbines and the stage loading can be significantly greater than that achieved by a compressor stage. This results in a few stages of turbine driving many stages of 26

28 compressor, as observed in practice. The higher stage loading used in turbines also results in larger rotor and stator deflection, as show in Figure 3.9. The stage enthalpy change ( H) and axial velocity (Ca) can be made dimensionless by dividing these terms by U 2 and U to produce the stage loading and flow coefficient respectively, where U is the blade velocity. Thus: H Ψ = Stage loading coefficient 2 U Ca Φ = Flow coefficient U In terms of gas angles the stage loading coefficient is given by: H Ca = (tan( β 1) + tan( β 2)) U U [Smith ] generated a correlation relating turbine efficiency with stage loading and flow coefficient. This is often referred to as the Smith's plot and is shown in Figure It is a useful source of data when designing a turbine. The impact of increasing the rotor deflection, axial velocity and blade speed are also shown in Figure Figure 3.10 shows a typical Smiths Plot From equation 3.4, increasing the rotor deflection (β1+β2) increases the rotor inlet velocities V1 and C1 respectively. This increases the losses in the stage and thus reduces the stage isentropic efficiency. It should be noted that increasing the rotor deflection increases the stage loading coefficient while the flow coefficient remains constant. Increasing the axial velocity increases the flow coefficient (Ca/U) and will also increase the stage loading coefficient. This results in the increase in all the 27

29 velocity vectors thus reduces the stage isentropic efficiency. Increasing the blade speed will increase the stage work done (i.e. increases H) but achieves this by reducing the stage loading coefficient ( H/U 2 ). The reduced stage loading coefficient improves the stage efficiency as illustrated in Figure 3.10, which shows the effect of increasing rotor deflection, axial velocity and blade speed on stage isentropic efficiency Overall turbine characteristics Figure 3.11 shows a typical turbine flow characteristic Figure 3.12 shows a typical turbine efficiency characteristic As discussed in Section 3.1.3, the use of overall characteristics is more useful when predicting the performance of gas turbines. Again we plot turbine characteristics on a 28

30 non-dimensional basis and these non-dimensional groups for flow and speed are Mach numbers based on inlet flow and rotational speed of the turbine. Thus the nondimensional flow and speed for turbines are: W RT / γ ND and 2 D P γrt with compressors. respectively. These relationships are the same as we found Typical turbine characteristics representing the flow and (overall) isentropic efficiency are shown in Figures 3.11 and 3.12 respectively, for a series of turbine nondimensional speeds. The non-dimensional mass flow increases with pressure ratio and beyond a certain pressure ratio the non-dimensional flow is constant. This is the case when the nozzle is choked, which is usually the case. When the rotor chokes there is a small variation of non-dimensional flow with speed Turbine creep life and cooling Figure 3.13 shows a typical Larson-Miller curve We have discussed in Section 2 the importance of achieving high turbine entry temperatures as this increases the thermal efficiency and specific work. Increased thermal efficiency reduces fuel consumption and thus fuel costs, whereas increased specific work reduces engine size. However, materials under stress and above a certain temperature will elongate plastically although the stress in the material is below the yield point. This phenomenon is know as creep and is measured as the rate of strain per hour for a given stress and temperature. Gas turbine blade materials operating at high temperatures suffer from creep and therefore materials employed in the manufacture of turbines are special creep resistance nickel-based alloys. One parameter that is often used in predicting creep is the Larson-Miller parameters. It is defined as: 29

31 LM = 1.8T (20 + ln( t)) Where T is the metal temperature in K and t is the creep life in hours. The Larson-Miller parameter is related to the turbine blade material stress as shown in Figure The demand for increased performance of gas turbines has resulted in turbine entry temperatures exceeding even the turbine material melting point. Thus a substantial amount of turbine cooling is required to keep the turbine material below its melting point and achieve acceptable turbine creep life. There are many turbine cooling technologies available and we shall very briefly describe them here. Convective cooling One of the earliest forms of air cooled turbines was convective cooling. Here cooling air is passed through the turbine blade internals to remove heat primarily by convection. The cooling air is normally ejected into the gas stream at the trailing edge of the turbine blade. Impingement cooling Convective cooling requires substantial amounts of cooling air. To reduce cooling the air requirement, impingement cooling was developed. Cooling air is impinged onto the internal surfaces of the turbine blade material to increase the turbulence and hence, the heat transfer coefficient. The higher heat transfer coefficient results in less cooling air being required to carry out the required cooling duty. As with convective cooling, the cooling air is added to the gas stream, usually at the tip and trailing edges of the turbine blade. Transpiration cooling Transpiration cooling is a later cooling technology where the cooling air is allowed to diffuse through pores or fine passages in the turbine blade material such that the blade is covered with a layer of cooling air. The insulating property of air is used to control the blade temperature. Transpiration cooling requires the least amount of cooling air. Steam and mist cooling Steam and mist cooling technology is a very recent development where steam is used to cool the turbine blade material. It is mostly applicable in combined cycle power plants because of the ready availability of steam for turbine cooling. Unlike air cooling systems discussed above, the steam and mist cooling systems are purely an internal cooling system, the steam being returned to the steam plant after the cooling task has been carried out. The increased energy in the coolant returning to the steam plant contributes to a worthwhile improvement in the steam plant performance. Impact of turbine cooling technology on gas turbine performance Applying turbine cooling technology improves the gas turbine performance significantly, due to the resultant increase in turbine entry temperature, but it does impose some losses in the turbine. The use of cooling air directly reduces the net flow through the turbine, thus reducing the turbine power output. The mixing of the 30

32 cooling air with the gas stream introduces turbulence into the gas stream leading to further losses. Also the mixing of the cooling air with the gas stream reduces the rotor inlet gas temperature resulting in reduced work done by the turbine. The gas temperature at the rotor inlet is important and manufacturers often refer to this temperature as the stator outlet temperature or first rotor temperature. Although transpiration cooling requires less cooling air, the turbulence generated by such cooling techniques result in the greatest loss in turbine isentropic efficiency. Also the fine cooling passages can get blocked by dirt so that the blades are vulnerable to failure due to a lack of turbine cooling. Convective cooling on the other hand is less efficient but also less likely to be affected by blockages. In practice all the above mentioned air cooling technologies are employed to give the best cooling with reliability. Steam and mist cooling do not generally have the drawbacks of air cooling systems. Since such cooling systems are internal they do not disrupt the gas flow patterns around the turbine blade. Also, the absence of mixing of the coolant and the gas stream minimises the temperature drop at the rotor. However, a steam source is required and this would add to cost and therefore it is considered primarily in combined cycle plants where a ready source of steam is available. 31

33 4.0 Off-design behaviour of gas turbines The discussion in Chapter 2 relates to the design point performance of simple or open cycle gas turbines. As part of the design process we choose a compressor pressure ratio, maximum cycle temperature or turbine entry temperature and compressor/turbine isentropic efficiencies. The use of these data in the design point calculation yields the cycle thermal efficiency and the specific work. The required power demand is used in conjunction with the specific work to determine the airflow rate through the engine and thus the engine size. This information is employed in the design of the compressors, combustors and turbines. The design process of these components generates the component characteristics as discussed in Sections 3.1 and 3.3 above. Figure 4.1 shows a schematic representation of a singe shaft gas turbine A Gas turbine designed as descried above would normally achieve its design point performance. However, gas turbines have to operate for prolonged periods away from their design point conditions. Means to predict their performance at off-design conditions are somewhat complicated and discussed by [Walsh and Fletcher 2000, Saravanamutoo, Rogers and Cohen 2001, Razak 2007]. Methods to determine or predict the off-design performance of gas turbines are implicit and the use of engine component characteristics is mandatory. They involve generating estimates or guesses for certain variables and checks are then made to see if these estimates are correct. For example, for a single shaft gas turbine as shown in Figure 4.1, we would specify: 1) The compressor inlet pressure 2) The compressor temperature 3) The humidity at compressor inlet 4) The turbine exhaust pressure 5) Gas turbine speed as determined by the driven load. For power generation this speed will be constant and would correspond to the synchronous speed of the generator. 32

34 6) Either the power output or fuel flow. Other parameters such as compressor pressure can also be specified but this is unusual. The estimates for a single shaft gas turbine are: 1) Compressor inlet airflow 2) Compressor pressure ratio 3) Turbine entry temperature The corresponding checks would be: 1) Turbine flow compatibility determined from the turbine characteristics 2) The required gas turbine power output or fuel flow as specified in item 6 above 3) Gas turbine speed compatibility with that specified in item 5 above When these estimates match the checks the off-design performance of the gas turbine is determined. The cycle calculations (also used in the design point calculation) is included in the prediction of the off design performance of gas turbines. As the number turbine shafts increases, so does the number of estimates and checks. It should be noted that the number of estimates should match the number of check when performing the off-design calculation of gas turbines. If a single shaft gas turbine incorporates a variable inlet guide vane (VIGV), then a further estimate of the VIGV position is necessary. The corresponding check will usually be a specified exhaust gas temperature (EGT). Note that different VIGV positions will generate different compressor characteristics and the appropriate compressor characteristic should be used in the prediction of the off-design performance of the gas turbine. Figure 4.2 shows a schematic representation of two-shaft gas turbine operating with a free power turbine. The gas generator produces high temperature and pressure gas which the power turbine converts to useful power to drive the load. Note the power turbine is mechanically independent of the gas generator turbine. For a two shaft gas turbine operating with a free power turbine (Figure 4.2), the estimates are: 33

35 1) Compressor inlet airflow 2) Compressor pressure ratio 3) Turbine entry temperature 4) Gas generator turbine pressure ratio The corresponding checks are: 1) Gas generator turbine flow compatibility determined from its turbine characteristics 2) Gas generator power balance the power balance between the compressor and gas generator turbine 3) Power turbine flow compatibility determined from the power turbine characteristic 4) The required gas turbine power output or fuel flow as specified in item 6 above Thus we have four estimates and checks for a two-shaft gas turbine operating with a free power turbine compared with three estimates and checks for a single shaft gas turbine. It should be noted that the power turbine speed must be specified and would normally be determined by the driven load. The prediction of the off-design behaviour of gas turbines is quite complex and computer programs are often used to predict the off design performance accurately. However, making some assumption will significantly simplify the understanding of the off-design performance behaviour of gas turbines. To a very good approximation, gas turbines behave in a non-dimensional manner. This implies that the nondimensional flows and speeds, and pressure and temperature ratios, as discussed in Chapter 3 are important, and this will be evident in the following discussion. 4.1 Off design behaviour of the single shaft gas turbine The single shaft gas turbine configuration (Figure 4.1) is widely used in power generation. As the compressor, turbine and the load (e.g. electrical generator) are mechanically connected the compressor acts as brake in the event of the electrical load being shed. In this event, the compressor helps prevent the gas turbine from overspeeding. The disadvantage of single shaft gas turbines is the requirement of high starting powers, particularly with large gas turbines. Variable inlet guide vanes (VIGVs) are normally provided and closed during starting to reduce the starting power requirements. This is due to the reduced compressor airflow rate, thus reduced power demand from the compressor during starting. VIGVs are also used to maintain high exhaust gas temperatures during part-load operation, particularly in combined cycle plants. Significant improvement in thermal efficiency/heat rate can thus be achieved by operating at high turbine entry temperatures at part-loads. These issues will be discussed later in the course where we use the single shaft gas turbine simulator to help augment the understanding of gas turbine performance and operation. Referring to Figure 4.1, we can write the (non-dimensional) flow compatibility equation between the compressor and turbine as: 34

36 W 3 T3 W1 T1 P1 P2 T3 W 3 =..4.1 P3 P1 P2 P3 T1 W1 W 3 T3 W1 T1 Where is the turbine inlet non-dimensional flow, is the compressor P3 P1 P2 P3 inlet non-dimensional flow, is the compressor pressure ratio, is the combustor P1 P2 pressure ratio and usually is less than unity due to the frictional losses known as the cold loss and the fundamental loss due to the addition of heat resulting from the T3 burning of fuel in the combustor. is the ratio of the maximum to minimum cycle T1 W 3 temperature ratio. T3 is also referred to as the turbine entry temperature. is the W1 ratio of the gas flow rates (W3) at inlet to the turbine to the compressor inlet airflow rate (W1). The flow rate (W3) at entry to the turbine will usually be different from that at entry to the compressor (W1). This is due to the addition of fuel in the combustion, which will pass through the turbine, while part of the cooling air bleeds, usually taken at the compressor exit, for turbine cooling as discussed in Section above, will bypass the turbine. We have ignored the gas property terms (γ ratio of specific heats and gas constant R) to simplify our discussions. These parameters usually have a small effect of engine performance and we shall discuss their effects where relevant. Also, we have omitted the dimensional term (D) as for a given compressor and turbine this parameter is constant. The flow compatibility described by equation 4.1 is particularly useful as it includes compressor ratio and cycle temperature ratio (T3/T1). These parameters strongly influence engine performance. Figure 4.3 shows operating point on the compressor characteristic due to the increase in T3/T1 35

37 If we restrict our discussion to the high operating power range of the gas turbine and the turbine is choked at these power ranges. Thus the turbine inlet non-dimensional flow (W3 T3/P3) will be constant, as shown in Figure 3.11 (this is assuming a single line turbine characteristic and would be the case if the nozzle guide vane is choked. There would be some variation of the turbine non-dimensional flow with nondimensional speed if the rotor choked. But this variation is quite small). Also, at normal compressor operating speeds the compressor speed line describing the change in compressor inlet non-dimensional flow with compressor pressure ratio is steep, as shown in Figure 4.3. Thus for a given non-dimensional speed the compressor inlet non-dimensional flow (W1 T1/P1) will be approximately constant. The air/gas flow ratio (W3/W1) does not vary a great deal in the normal operating range and can be assumed to be approximately constant. Therefore any increase in temperature ratio T3/T1 will result in a decrease in the ratio P1/P2 to satisfy the flow compatibility equation 4.1. A decrease in the ratio P1/P2 implies an increase in compressor ratio (P2/P1). Thus an increase in T3/T1 also increases the compressor pressure ratio P2/P1. This is illustrated in Figure 4.3. Figure 4.4 shows the variation of compressor non-dimensional compressor temperature rise with pressure ratio for a series of compressor non-dimensional speeds The increase in compressor pressure ratio will also increase the turbine pressure ratio. The increase in turbine entry temperature (T3) and turbine pressure ratio will increase the power output from the turbine section. The compressor temperature rise ( T21 = T2-T1) will also increase due to the increase in compressor pressure ratio. However, this is quite small as illustrated in Figure 4.4, which shows the compressor nondimensional temperature rise ( T21/T1) varying with compressor pressure ratio (P2/P1) for a series of compressor non-dimensional speeds (N1/ T1). For a given 36

38 compressor non-dimensional speed, the relatively flat nature of the non-dimensional compressor temperature rise is due to a significant change in compressor efficiency with pressure ratio. Thus at low compressor pressure ratios, the loss in compressor efficiency results in a small decrease in compressor temperature rise. At higher pressure ratios the improved compressor efficiency partly compensates for the higher pressure ratio, thus resulting in a small increase in compressor temperature rise with pressure ratio. These issues are discussed by [Saravanamutoo, Rogers and Cohen 2001]. Therefore, when operating along a constant compressor non-dimensional speed line the compressor specific work absorbed is approximately constant. Thus, the specific work output from the gas turbine increases significantly with increase in T3/T1. Due to the steep nature of the speed line on the compressor characteristic, the variation of the compressor non-dimensional flow (W1 T1/P1) with pressure ratio is small. Therefore, for a given ambient condition (P1 and T1), the change in airflow rate (W1) through the compressor with increasing pressure ratio is small. Thus the airflow rate (W1) through the compressor is approximately constant with increase in T3/T1 and compressor ratio (P2/P1). Therefore almost all of the power increase, due to the increase in the turbine entry temperature (T3), is due to the increase in specific work, which is achieved by increasing the fuel flow to the gas turbine. From our discussions in Chapter 1, the increase in T3/T1 and compressor pressure ratio will increase the thermal efficiency of the gas turbine (i.e. reduces the heat rate) Effects of ambient condition on engine performance Changes in ambient conditions have a profound effect on gas turbine performance. Ambient conditions can change due to changes in ambient temperature, pressure and humidity. We shall now consider the effects of changes in these parameters on engine performance Ambient temperature The speed (N1) of a single shaft gas turbine is normally constant (assuming power generation, which is usually the case). Therefore, a high ambient temperature results in a decrease in compressor non-dimensional speed (N1/ T1). The decrease in compressor speed (N1/ T1) will result in a decrease in compressor non-dimensional flow (W1 T1/P1). If we continue to operate the gas turbine at a constant turbine entry temperature (T3) by adjusting the fuel flow, the increase in ambient temperature will result in an increase in compressor inlet temperature (T1), thus T3/T1 will decrease. Making the assumptions regarding compressor non-dimensional temperature rise, the steepness of the compressor speed lines and gas property changes, as discussed in Section 4.1, the decrease in compressor non-dimensional flow and temperature ratio (T3/T1) will require an increase in P1/P2 to satisfy the flow compatibility equation 4.1. Thus an increase in ambient temperature, while operating at a constant turbine entry temperature, will result in a decrease in compressor pressure ratio (P2/P1). The reduction in compressor pressure ratio and T3/T1 will therefore reduce the specific work and thermal efficiency. (It should be pointed out that the increase in ambient temperature and thus compressor inlet temperature T1 will increase the specific work as implied by equation 2.8 in Chapter 2. However, the decrease in compressor ratio 37

39 and temperature ratio is sufficient to decrease the gas turbine specific work, which then reduces the power out of the gas turbine.) Figure 4.5 shows lines of constant gas turbine power output with ambient temperature on the compressor characteristic This picture is however not quite complete. The decrease in compressor nondimensional flow (W1 T1/P1), due to the decrease in compressor non-dimensional speed (N1/ T1), will decrease the airflow rate through the engine, thus reducing the power output of the gas turbine even more (i.e. W1 T1/P1 decreases due to the increase in T1, which will cause W1 to decrease). It should also be noted that the increase in T1 will also further reduce W1. The impact of ambient temperature on engine performance is shown in Figure 4.5, where lines of constant gas turbine power output with ambient temperature are shown on the compressor characteristic Ambient pressure If we assume that the ambient temperature (T1) and gas turbine speed (N1) is constant, a decrease in ambient pressure does not affect the compressor nondimensional speed (N1/T1). Therefore, the compressor non-dimensional flow (W1 T1/P1) is largely constant, due to the steep speed line on the compressor characteristic, as discussed in Section 4.3. Any decrease in ambient pressure and thus P1 will result in a corresponding reduction in compressor inlet airflow rate (W1). This will reduce the power output of the gas turbine. If we assume we continue to operate at a constant turbine entry temperature (T3), from the flow compatibility equation 4.1, the compressor pressure ratio is constant. Therefore there is no change in the thermal efficiency or heat rate. It should be pointed out that if external or auxiliary loads exist, such as power demand from inlet chillers (vapour compression refrigeration), they could demand approximately constant load as the ambient pressure decreases. Thus in the presence of such auxiliary loads a decrease in ambient pressure will decrease the thermal efficiency. (i.e. heat rate will have to increase). 38

40 Humidity effects Humidity reflects the amount of water vapour trapped in air. The presence of water vapour in air affects the gas properties of air and this is what affects the gas turbine performance. There are two types of humidity and are referred to as the relative and absolute or specific humidity. The first is a measure of the amount of water vapour that should be evaporated to saturate the air (i.e. make the humidity 100%) and it is the valve usually quoted in metrological forecasts. The second, specific humidity reflects the mass of water vapour trapped in a unit mass of dry air. Although there is a relationship between relative and specific humidity, it is the specific humidity that affects engine performance. Figure 4.6 shows variation of specific humidity with ambient temperature The increase in specific humidity of air increases its gas constant (R) and specific heat of constant pressure (cp) while reducing the isentropic index γ (ratio of specific heats (cp/cv. cv is the specific heat at constant volume). However, the decrease in γ is smaller than the increase in R and cp. Therefore an increase in humidity would decrease the compressor non-dimensional speed (N1/ γrt1) primarily due to the increase in the gas constant of air (R). (Note we are now including the gas property terms present in the non-dimensional parameters). The decrease in compressor nondimensional speed will also decrease the compressor non-dimensional flow (W1 (γrt1)/p1). The decrease in the non-dimensional flow and R will decrease the airflow rate through the compressor, which will reduce the power output of the gas turbine. W 3 R3T 3/ γ 3 P3 W1 = R1T 1/ γ1 P1 P2 P1 P2 P3 T3 W 3 T1 W1 R3 R1 γ γ 3 where R1 and γ1 are the gas constant and isentropic index of air at inlet to the compressor respectively. R3 and γ3 are the gas constant and isentropic index at inlet 39

41 to the turbine and are usually different from R1 and γ1 due to the change in gas composition due to the presence of products of combustion. (Note: any change in R1 and γ1 due to humidity changes will be carried through the other engine components, namely the turbine, and resulting in the ratios R3/R1 and γ3/ γ1 to remain approximately constant. Equation 4.2 is effectively equation 4.1 but includes all the gas property terms.) If we continue to operate at a constant turbine entry temperature while the humidity changes (i.e. by adjusting the fuel flow into the combustor), it is evident from equation 4.2 that any decrease in compressor non-dimensional flow (W1 (γrt1)/p1) will decrease the compressor pressure ratio (P2/P1). The decrease in compressor pressure ratio will also decrease the turbine pressure ratio. As the turbine entry temperature (T3) is constant, the decrease in turbine pressure ratio will result in an increase in the exhaust gas temperature (EGT), which corresponds to the temperature at salient point 4 in Figure 4.1. The gas turbine power output is usually limited by the EGT. Therefore any increase in T4 (EGT) will lead to a decrease in fuel flow to maintain the EGT. This will reduce the turbine entry temperature (T3). Although the increase in specific heat (cp), due to the increase in humidity, will increase the specific work (see equation 2.8, in Chapter 2), the decrease in T3 hence T3/T1, compressor pressure ratio (P2/P1) and airflow rate (W1) will reduce the power output and thermal efficiency of the gas turbine. Thus for a single shaft gas turbine an increase in humidity has a negative impact on engine performance. However, the impact of humidity on engine performance is small but will be noticeable at high ambient temperatures because the change in specific humidity with relative humidity is most significant at high ambient temperatures. This effect can be seen in Figure 4.6, which shows a simplified psychrometric chart. For a given ambient temperature and relative humidity, the specific humidity is dependent on the ambient pressure and increases with the decrease in ambient pressure. Although we stated in Section that the ambient pressure does not affect the thermal efficiency of the gas turbine, in humid air there will be a very small decrease in thermal efficiency with ambient pressure. This is due to the increase in specific humidity as discussed above. Also, this increase in specific humidity will slightly add to the power loss due to the decrease in ambient pressure. 4.2 Off-design behaviour of a two-shaft gas turbine operating with a free power turbine The application of two-shaft gas turbines operating with a free power turbine, as illustrated in Figure 4.2 above, is widely found in mechanical drive, such as process compressors and pumps, and naval applications. In these applications the speed of the drive equipment will vary with load which is determined by the system resistance of the process system. For example, in oil and gas exploration and production, the driven process compressor speed may be low during dense phase operation (i.e. high suction pressures at the process compressor inlet) but the power demand will be high in order to maximise oil and gas production. In such applications a free power turbine arrangement is better suited, as the gas generator can operate at the maximum rating while the power turbine will be constrained to operate at the speed of the driven 40

42 equipment. If we attempt to achieve such a task using a single shaft gas turbine, the gas turbine speed will be low and consequently its performance will be poor. Figure 4.7 shows the matching of turbines operating in series As a result of the engine component arrangement of a two-shaft gas turbine, namely the splitting of the turbine section into two (gas generator and power turbine, as illustrated in Figure 4.2), the off-design behaviour of a two-shaft gas turbine operating with a free power turbine is very different to that of a single shaft. Although there is no mechanical link between the gas generator and the power turbine, there exists a strong fluid or aerodynamic coupling between these turbines. It is this coupling or matching of these turbines, as shown in Figure 4.7, that result in a different off-design behaviour of the two-shaft gas turbine. In the normal (high) power operating range the power turbine is most likely to remain choked and its pressure ratio will vary between points (2) and (3), as shown in Figure 4.7. Choked power turbine operation means that its non-dimensional flow remains constant and must equal the exit or outlet nondimensional flow from the gas generator turbine. In other words, the swallowing capacity of the power turbine controls the gas turbine pressure ratio. Thus, during choked power turbine operation the gas generator pressure ratio remains constant, as indicated by point (2,3) in Figure 4.7. At low power operation, when the power turbine becomes un-choked, as indicated by point (1) on the power turbine characteristic (Figure 4.7), the gas generator pressure ratio decreases to point (1) on the gas generator turbine characteristic (Figure 4.7). This reduction in gas generator pressure ratio is necessary to maintain the flow compatibility between the gas generator and power turbine. Referring to Figure 4.2, we can write the flow compatibility equation between the gas generator and power turbine as follows: W 4 T 4 P4 W 3 T3 P3 T 4 W 4 = P3 P4 T3 W 3 41

43 where W3 T3/P3 is the non-dimensional flow at entry to the gas generator turbine P3/P4 is the gas generator turbine pressure ratio T3/T4 is the gas generator turbine temperature ratio W4 T4/P4 is the power turbine inlet non-dimensional flow And from the turbine isentropic efficiency equation: γg 1 T 4 P4 γg = 1 ηt T3 P3 Where γg and η t are the isentropic index (cp/cv) and isentropic efficiency respectively. From Figure 3.12, at normal turbine operating speeds (80% to 100% speed) the turbine isentropic efficiency remains essentially constant with pressure ratio (at high pressure ratios). Thus the gas generator temperature ratio (T4/T3) is constant as the power turbine pressure ratio varies with load. This also implies that the gas generator turbine non-dimensional temperature drop ( T34/T3, where T34 = T3-T4) is also constant for the choked power turbine operating range. In the steady state, the (non-dimensional) power balance between the compressor and the gas generator turbine is: T 21 = T1 T34 T3 T3 T1 cpg cpa W W1 (Note that all the compressor power is provided by the gas generator turbine and the power turbine power output drives the load) We have stated on Section 4.1 that for a given compressor non-dimensional speed the compressor non-dimensional temperature rise ( T21/T1) is approximately constant. Therefore (from equation 4.5), the maximum to minimum cycle temperature ratio (T3/T1) is also constant. We have also stated in Section 4.1 that the steepness of the compressor speed line makes the compressor inlet non-dimensional flow (W1 T1/P1) approximately constant with compressor pressure ratio. Thus, from the flow compatibility equation between the compressor and gas generator turbine (which is the same as equation 4.1) the compressor pressure ratio is fixed for a given compressor non-dimensional speed. Hence, for each compressor non-dimensional speed, there is a unique compressor pressure ratio on the compressor characteristics. Joining these pressure ratios gives a unique equilibrium or steady state running line on the compressor characteristics, as shown in Figure 4.8. It should be noted that this is only true for a single line turbine characteristic where the turbine non-dimensional flow capacity is independent of the turbine non-dimensional speed. This is generally the case when the number of turbine stages is small (no more than three stages). If the number of stages is large (e.g. six stages) then there is a decrease in turbine nondimensional flow with increasing turbine non-dimensional speed and the running line on the compressor characteristic would not be unique. However, this deviation in turbine flow with speed is quite small and even in this case the running line on the turbine characteristic is, to a good approximation, unique. 42

44 Figure 4.8 shows the steady state running line on the compressor characteristics for a two-shaft gas turbine operating with a free power turbine Effects of ambient condition on engine performance In Section 4.1.2, we discussed the impact of ambient conditions on engine performance for a single shaft gas turbine. Two-shaft gas turbines operating with a free power turbine is also affected by ambient conditions but some of the performance characteristics are different and discussed as follows Ambient temperature If we consider the two-shaft gas turbine operating at a constant maximum cycle temperature (i.e. T3 is constant), any increase in ambient temperature (T1) will result in a decrease in the temperature ratio T3/T1. From gas generator power balance, equation 4.5, the compressor non-dimensional temperature rise ( T21/T1) will decrease. The decrease in T21/T1 will result in a decrease in compressor nondimensional speed (N1/ T1), as indicated by Figure 4.4. From the above analysis for a two-shaft gas turbine operating with a free power turbine, the operating point on the compressor characteristic will move down the (unique) steady state running line. This differs from a single shaft gas turbine, which produces a family of constant power lines, as shown in Figure 4.5. The decrease in compressor non-dimensional flow and thus the compressor inlet airflow rate, pressure ratio and maximum to minimum cycle temperature ratio (T3/T1) results in a decrease in gas turbine power output and thermal efficiency. This is similar to that discussed above for the single shaft gas turbine case. Unlike the case for a single shaft gas turbine where the gas turbine speed remains constant, in this case the gas generator speed will change. As to whether there is an 43

45 increase or decrease in gas generator speed depends on the shape of the compressor characteristic (i.e. how close the speed lines cluster together on the compressor characteristic) and the matching of this characteristic with other component characteristics such as the turbine. Generally, there is an increase in gas generator speed with decrease in ambient temperature Ambient pressure The effect of ambient pressure on engine performance is similar to that discussed for the single shaft gas turbine case (Section ). If we consider operating the twoshaft gas turbine at constant T3/T1 while the ambient pressure changes, from equation 4.5 the compressor non-dimensional temperature rise ( T21/T1) will also be constant (assuming turbines are choked). From Figure 4.4, a constant T21/T1 implies a constant compressor non-dimensional speed (N1/ T1). From the flow compatibility equation 4.1, the compressor pressure ratio will also remain constant as the ambient pressure decreases. This is due to the compressor non-dimensional flow (W1 T1/P1) being approximately constant due to the steep speed line on the compressor characteristic. However, the decrease in ambient pressure will decrease (P1), which will decrease the compressor airflow rate (W1). Thus the power output will decrease with the reduction in ambient pressure. The operating point on the compressor characteristic will remain largely constant. Since the compressor pressure ratio and T3/T1 are not affected by the decrease in ambient pressure, the thermal efficiency remains constant Humidity As stated in Section , the effect of humidity alters the gas property terms, which in turn affect engine performance. For a two-shaft gas turbine operation with a free power turbine, the impact of humidity on performance is somewhat different compared with that on a single shaft gas turbine. If we consider operating at a constant T3/T1, then the compressor non-dimensional temperature rise ( T21/T1) will also be constant. (note:- T21/T1 is affected by the ratio of specific heats, γ, but this change is small compared with the change in gas constant (R) and specific heat at constant pressure (cp) of air with the change in humidity). Thus the compressor nondimensional speed (N1/ γrt1) is constant (note:-we have included the gas property terms to explain the effect of humidity on engine performance). However, the increase in humidity will increase the gas constant (R) of air. Thus the compressor/gas generator speed N1 increases to maintain the constant compressor non-dimensional speed, as observed in practice. Unlike a single shaft gas turbine, a two-shaft engine operating with a free power turbine can operate at constant temperature ratio (T3/T1) because of the gas generator speed can change. This differs from a single shaft gas turbine where T3/T1 decreases with increase in humidity and this is due to the decrease in compressor nondimensional speed. The increase in specific heat with humidity results in an increase in specific work (equation 2.8). Although the compressor inlet airflow rate decreases because of the increase in humidity, the increase in specific work in this case results in a noticeable increase in power output of a two-shaft gas turbine operating with a free power turbine, particularly at high ambient temperatures. However, the thermal efficiency decreases with increase in humidity. 44

46 We have stated that the specific humidity increases as the ambient pressure decreases. Thus the increase in humidity will partly compensate for the losses in power as the ambient pressure decreases, as discussed in Section This is different to the case of the single shaft gas turbine where the increase in humidity augments the loss in power output with the decrease in ambient pressure. As stated above (Section ), the effects of humidity on engine performance is small and largely noticeable at high ambient temperatures. 4.3 Performance deterioration We have shown that the performance of gas turbines is dependent on the component characteristics and the interaction of these characteristics. In fact, the measurable parameters such as pressure, temperatures flow and speeds are determined by the performance and the interaction of these components, which in turn determines the power output and thermal efficiency. Figure 4.9 shows the change in the compressor blade profile due to fouling When performance deterioration occur these component characteristics alter. The interaction of these deteriorated characteristics generally results in a loss of power and thermal efficiency. They also give rise to changes in measurable parameters. All gas turbine deteriorate during operation and the following are some causes of performance deterioration: 1) Fouling 2) Hot End Damage 3) Tip Rubs 4) Vibration 5) Seal Wear & Damage 6) Variable Guide Vane Schedule 7) Foreign and domestic object damage 8) Erosion 9) Corrosion Compressor fouling is the most common form of performance deterioration and occurs due to the build-up of dirt on the compressor stages. This reduces the flow capacity of the compressor and decreases its efficiency. Compressor fouling affects the flow capacity to a greater extent than efficiency. For example, a moderately fouled 45

47 compressor results in the compressor flow capacity decreasing by about 3% while the efficiency decreases by about 1% to 1.5%. Figure 4.9 shows the build-up of dirt on a compressor stage and Figure 4.10 shows the change in the compressor characteristic due to fouling. The shift in the running line and operating point on the compressor characteristic is also shown in Figure Figure 4.10 shows the change in the compressor characteristics, running line and operating point due to fouling The rate of compressor fouling and thus the loss in engine performance due to fouling decreases with time. This is primarily due to the decrease in the rate of build-up of dirt and debris on the compressor blades with time. The larger the build-up of dirt on the compressor blades increases the tendency to blow away the excess dirt by the flow of air through the compressor. Hot end damage occurs due to prolonged exposure of the turbine to high temperatures. The main effect is the burning of the turbine nozzle guide vanes, particularly the trailing edge which controls the flow capacity of the turbine. Also, combustion problems can cause hot streaks of gases, damaging the turbine. Therefore, hot end damage results in increased turbine non-dimensional flow and reduced turbine efficiencies. For a two-shaft gas turbine operating at constant exhaust gas temperature, an increase in turbine capacity due to hot end damage reduces the compressor pressure ratio and increases turbine entry temperature, resulting in a loss in power out and thermal efficiency. At constant gas generator speed operation, hot end damage will increase the power output and turbine creep life usage but decrease the thermal efficiency. These issues will be discussed later when we use simulators to illustrate these types of faults. Performance deterioration results in increased life cycle cost. Unlike compressor fouling which can be recovered by washing the compressor, to recover the performance loss due to hot end damage will require repair or replacement of the damaged turbine. Rubs and seal wear also affect engine performance by largely reducing the component efficiencies, but these are usually long term deteriorations and their effects are normally observed after tens of thousands of operating hours. Major engine overalls 46

48 are required to recover the gas turbine performance in these cases. Since the low pressure (LP) stages of the compressor control its flow capacity at normal operating speeds, any increase in tip clearances between the compressor LP rotor tip and the casing due to rubs can affect both the compressor flow capacity and efficiency. Rubs in the high pressure compressor stages usually affect the compressor efficiency rather than its flow capacity. 47

49 5.0 Engine control systems The gas turbine power output is primarily determined by the heat input (thermal input), which is usually achieved by burning fuel in the combustor. The higher the fuel flow, the higher is the power output. However, the power output of the gas turbine cannot be increased indefinitely and is limited by the design limitations of the engine components. The limitations include turbine temperature and gas turbine speed. These limits are imposed to achieve suitable turbine creep life, as discussed in Section above, and limit the stress in the rotating members of the engine components, such as disks and blades present in the compressor and turbine to their design values. It is the function of the control system to achieve the desired power demand but also subjecting the engine to the temperature and speed limitations. 5.1 Proportional, Integral and Derivative control The control system achieves the control functions by the use of a PID control, which stands for Proportional, Integral and Derivative control. The PID algorithm is usually defined as: Kc d( err) OP = err Kc + err. dt + Kc Td..5.1 Ti dt Where OP is controller output, Kc is the proportional gain, Kc/Ti is the reset rate or integral gain and Kc x Td is the derivative gain. The err term is the controller error and represents the percentage deviation between the power output from the gas turbine and the set point. The set point is the desired or required power output. The output from the controller acts on the fuel valve either opening or closing the fuel valve until the desired power output is achieved (i.e. power output specified by the set point). Sometimes the proportional gain is replaced by the proportional band, which is defined as the change in the input to cause a change in the output from 0 to 100%. Thus the proportional gain Kc is give by Kc = 100/PB where PB is the proportional band. A proportional-only controller will leave an offset and a manual adjustment to the fuel valve position is required to achieve the required set point. The addition of the Kc integral action err. dt eliminates the proportional offset and the action of integral Ti control is often referred to as automatic reset. Integral action also has drawbacks such as integral windup. Windup occurs when the set point is not achieved by the control system due to the control variable such as the fuel valve becoming fully opened. Since no change in fuel flow can occur, the error remains unchanged. Under this situation the integral output, which is effectively a summation process of the error, continues to increase hence winding up the output of the control system. If we reduce the set point (i.e. reduce the power demand) to such a level that the fuel valve will normally be partly opened the fuel valve will continue to remain fully opened because of the windup of the control system. Thus the power output of the engine does not change until the control system becomes fully unwound. This unexpected behaviour of the control system is called integral windup. To fix this problem we need to reset the output to 100% when the fuel valve is fully open. Now a reduction in set point would 48

50 result in the fuel valve closing and the gas turbine responding to any change in power demand. Derivative control augments the control system output during transients and is usually employed when the transient response or performance of the process is slow. Examples of such processes are blast furnaces. The transient response of gas turbines is much more rapid and derivative control is often omitted in their control. It should be noted that derivative control will produce no response due to steady state errors such as proportional offsets or integral windups. 5.2 Low signal selection The function of the control system should not only enable the engine to develop the power demanded as specified by the set point, but it should also protect the engine from damage. As stated above, exhaust gas temperature (EGT) limits are necessary to prevent the turbine from overheating in order to achieve the required turbine creep life. Also, speed limits for compressors and turbines are necessary to prevent rotating members from exceeding design stress levels. Speed limits may also be imposed as determined from rating curves, which describe the variation of gas turbine power output with ambient temperature. Rating curves find their origins in aero-gas turbines. In hot ambient temperature and/or low ambient pressure environments the thrust/power output from the engine at take-off will decrease, as discussed in Sections and above. In these situations the EGT can be increased sufficiently to achieve the required take-off thrust at the expense of increase turbine creep life usage. However, in environments where the ambient temperatures are low, the required take-off thrust will be achieved at a lower EGT, thus the creep life usage is low. Such trade-offs in turbine creep life usage results in the overall creep life of the turbine remaining unaffected. Manufacturers of industrial gas turbines, particularly aero-derivatives may use such rating curves to increase the power output of the gas turbine at high ambient temperatures. At low ambient temperatures, the compressor/gas generator speed limit may be used to reduce the EGT, thus reducing the turbine creep life usage. This will compensate for the increased turbine creep life usage at high ambient temperatures. As the ambient temperature decreases the compressor non-dimensional speed (N1/ T1) increases. The increase in compressor non-dimensional speed can force the LP stages of the compressor to choke, resulting in the stalling of HP stages. The stalling of the HP stages can lead to compressor surge. This was discussed in Section Therefore it is necessary to impose a compressor non-dimensional speed limit to avoid surge at low ambient temperatures. To prevent the engine from exceeding these limiting values, errors are calculated using these limiting values (EGT and speeds) as set points, including the power required. They are compared with each other and the lowest values are used by the control system via equation 5.1. This method of signal selection is referred to as low signal selection. As an example, the error (err) is calculated as follows: 49

51 P SetP err = 100,Where P is the power output and Setp is the required power SetP output (set point). EGT EGTl err = 100,Where EGT is the exhaust gas temperature and EGTl is the EGTl maximum exhaust gas temperature limit. err = limit. N1 N1max 100,Where N1 is the speed and N1Max is the maximum speed N1max N1 N1max = T1 T1 N1 err 100,Where is the compressor non-dimensional speed N1max T1 T1 N1max limit and is the limiting compressor non-dimensional speed. T1 Similarly, other limiting values such as pressures or powers can also be included in the low signal selection process. A typical gas turbine control strategy is illustrated in Figure 5.1. Figure 5.1 shows a simplified schematic representation of a control system applicable to a two-shaft gas turbine operating with a free power turbine 50

52 The above discusses the control of the gas turbine operating under steady state conditions. However, gas turbines have to operate under transient conditions and these transients arise due to the change in power demand. The changes in ambient conditions, such as ambient temperature, pressure and humidity also introduce transients, but these are usually slow transients. Increasing the power output of the gas turbine requires an increases the fuel flow. A rapid increase in fuel flow can lead to flame out, thus tripping the engine due to the fuel air ratio becoming too rich. Such increases in fuel flow can also overheat the turbine resulting in rapid usage of turbine creep life. There is also the possibility of compressor surge due to the operating point leaves the steady state operating line during transient operation, where the compressor pressure ratio increases more rapidly compared with the compressor airflow rate. Similarly during deceleration, the fuel-air ratio decreases and the weak or lower extinction limit may be exceeded resulting in flameout. Figure 5.2 shows the acceleration-deceleration limit lines and the expected transient running line for a two-shaft gas turbine operating with a free power turbine To prevent such damage and trips, acceleration and deceleration limit lines are introduced. For example, at any give compressor non-dimensional speed the acceleration and deceleration limit lines are used to determine the limiting compressor pressure ratios. Errors are calculated using these limiting values for pressure ratios and the current (operating) compressor pressure ratio. These errors are used in the signal selection process described above to determine the error value for use in the control system. It should be noted that during deceleration, high signal selection is used. Figure 5.2 shows an example of transient running lines on the compressor characteristic. The steady state running line, acceleration and deceleration limit lines are also shown in the figure. Since the airflow measurement is often unavailable, the compressor non-dimensional flow is replaced by the compressor non-dimensional speed, as shown in Figure 5.3. The compressor pressure ratio shown in Figure 5.2 may be replaced by the non-dimensional fuel flow. 51

53 Figure 5.3 shows the acceleration and deceleration lines on a pressure ratio versus speed basis 52

54 Part 2 The use of simulators can eloquently illustrate the performance and operability of a gas turbine. The gas turbine simulators are based on the quasi-steady state model using time constants to simulate the transient effects. Although such simulators are only strictly valid under steady state conditions, much useful insight into engine operation can be achieved during transients. They are excellent as training simulators. The concept of component matching, as described in [Walsh and Fletcher 2000, Saravanamutoo, Rogers and Cohen 2001, Razak 2007] and discussed briefly in Section 4, has been extensively used in building the simulator. It is therefore capable of illustrating what has been discussed in this course in detail including turbine creep life usage, engine emissions, as discussed in [Boyce ] and [Lefebvra ] respectively. As stated, component matching is the interaction of the various gas turbine component characteristics, namely compressors, combustors and turbines. It is the interaction of these component characteristics that is responsible for determining engine parameters such as power output, pressures, temperatures, flows and speeds. These parameters indeed affect turbine creep life and engine emissions which can therefore be simulated. This includes the impact of humidity, which has a profound effect on gas turbine emissions, namely NOx. The simulator is also capable of illustrating the impact of performance deterioration on engine performance in detail. The deterioration changes the component characteristics and normally results in a reduction in power out and thermal efficiency. Of course, any simulator would require a control system and a simple PID (proportional, integral and derivative) controller has been used in the simulator. Thus control system performance faults such as proportion offset, integral windup and engine trips can also be simulated. Details about control systems may be found in [Shaw ]. 6.0 Single shaft gas turbine performance and operation The use of single shaft gas turbines is widespread in power generation and is therefore assumed to drive an electrical generator. Since the generator operates at a constant speed which corresponds to the synchronous speed, a single shaft gas turbine also has to operate at a constant speed. Thus, the gas turbine speed remains constant with load and ambient condition changes. The single shaft gas turbine simulator corresponds to an advanced gas turbine having an ISO rating of about 40 MW and incorporates a variable inlet guide vane (VIGV). The reader is encouraged to run the necessary simulations discussed and produce the graphical displays describing the engine performance, turbine creep life and emissions under various operating conditions. The single shaft gas turbine simulator user guide gives details on how to operate the simulator. Note:- the simulator should be set to operate using fixed exhaust gas temperature (EGT) control unless stated otherwise, (i.e. the EGT limit remains constant with the change in ambient temperature). See user guide for details using fixed and variable temperature control. Although some generalisation is possible regarding the performance and operability of single shaft gas turbines using the simulator, however, the discussion in this course is strictly applicable to the engine described by the simulator. 53

55 6.1 Effects of ambient temperature on gas turbine performance From the discussion in Section , the ambient temperature has a significant impact on gas turbine performance. The higher the ambient temperature the lower is the power output and thermal efficiency and conversely a lower the ambient temperature results in better engine performance. The better gas turbine performance at low ambient temperatures is due to the increase in the maximum to minimum cycle temperature ratio (T3/T1), compressor pressure ratio and airflow rate. However, at low ambient temperatures, the single shaft gas turbine performance may be limited by the power output rather than the exhaust gas temperature (EGT), which limits the performance of the gas turbine at high ambient temperatures. The running line on the compressor characteristic is steep and may result in the running line intersecting the compressor surge line at low ambient temperatures, a condition which must be avoided. Adequate margin between the surge line and the operating point must be maintained to allow for transient conditions. This is achieved by limiting the maximum power output from the gas turbine, which in this case is set to 45 MW. Limiting the power output as the ambient temperature decreases will result in a decrease in turbine entry temperature (T3). This is necessary to reduce the specific work in order to maintain the power output from the gas turbine. The reduction in T3 reduces the turbine creep life usage at these ambient temperatures. Thus manufacturers may impose the maximum power limit to ensure satisfactory turbine creep life at high ambient temperatures referred to as rating curves, as discussed in Section 5.2. Figure 6.1 shows the variation of gas turbine power output when EGT or power limited The simulator may be used to illustrate these effects on gas turbine performance due to ambient temperature change. The ambient temperature may be varied from 30 degrees Celsius down to -30 degrees Celsius over a large enough time period, say, one hour (ramp time=3600 seconds). The power demand from the generator may be maintained at 60 MW thus ensuring that the engine is always on a limiting condition during the ambient temperature transient. 54

56 The change in gas turbine/generator power output may be plotted using the output data from the simulator. This plot (Figures 6.1), which the user should reproduce, shows the power output increasing as the ambient temperature decreases when the EGT limits the performance of the gas turbine. Although the decrease in ambient temperature results in a decrease in the specific work of the gas turbine, the increase in maximum to minimum temperature ratio (T3/T1) increases the specific work. For a given value of T3/T1 the increase in compressor ratio has little effect on the specific work due to the relatively high compressor design pressure ratios of this engine (at lower design compressor pressure ratios the power output increases with increase in compressor pressure). The increase in compressor non-dimensional speed, and therefore the increase in compressor non-dimensional flow, increases the air flow through the engine. The effect of these factors increases the power output of the gas turbine during the period of engine operation when the engine performance is limited by the exhaust gas temperature (EGT). The thermal efficiency of the gas turbine increases as shown in Figure 6.2 and this is primarily due to the increase in both compressor pressure ratio and T3/T1. The parameters that affect the engine performance can also be seen as trends on the simulator outputs. Figure 6.2 shows the variation of gas turbine thermal efficiency when EGT or power limited As the ambient temperature decreases below 1 degree Celsius the gas turbine performance is power limited and the power output from the gas turbine remains constant. During this period of operation the rate of increase in T3/T1 decreases. The decrease in ambient temperature and thus the compressor inlet temperature is sufficient to decrease the specific work as shown in Figure 6.3. Although the air flow rate increases during this period of operation, the decrease in specific work maintains a constant power output from the gas turbine. The compressor pressure ratio also increases during this period of operation. The increase in the compressor pressure ratio and T3/T1 increases the thermal efficiency of the gas turbine during constant power operation as shown in Figure

57 Figures 6.1 and 6.2 have been produced for three different values for relative humidity (0%, 60% and 100%). It should be noted that the specific humidity which is the amount (mass) of water vapour present in the dry air, rather that the relative humidity which is the amount of water required to saturate the air, affects gas turbine performance. The effect of specific humidity on power and thermal efficiency is small. However, the increase in specific humidity reduces both power output and thermal efficiency. As discussed in Section , the increase in specific humidity increases the gas constant of air and this reduces the compressor non-dimensional speed, which also reduces the compressor non-dimensional flow. Thus the increase in humidity reduces mass flow rate, compressor pressure ratio and turbine entry temperature, and therefore decrease the power output and thermal efficiency. Figure 6.3 shows the variation of specific work when EGT or power limited Figure 6.4 shows the creep life usage as time to next overhaul when EGT or power limited 56

58 It is observed that the effect of humidity is more pronounced at high ambient temperatures. For a given increase in relative humidity, the increase in specific humidity is greater at high ambient temperatures as illustrated in Figure 4.6 above. Thus the impact of humidity on gas turbine performance decreases with ambient temperature. As the ambient temperature decreases, the turbine creep usage increases during the period of constant EGT operation. This is primarily due to the increase in compressor pressure ratio and thus turbine pressure ratio while operating at a constant EGT. The turbine entry temperature (TET or T3) and hence the turbine blade metal temperature increase, as the ambient temperature decreases. At ambient temperatures when the power output from the gas turbine is constant the turbine entry temperature (T3) decreases. Thus the turbine metal temperature decreases. This results in a significant decrease in turbine creep life usage when operating at these ambient temperatures. The variation of turbine creep life usage with ambient temperature and humidity is illustrated in Figure 6.4. At high humidity the decrease in turbine entry temperature reduces the creep life usage. Figure 6.5 shows NOx and CO emissions when EGT or power limited Gas turbine emissions are affected primarily by combustion pressure, temperature, fuel-air ratio and humidity. An increase in combustion pressure, temperature and fuelair ratio results in increased Oxides of Nitrogen (NOx) emissions but reduce the emissions of Carbon Monoxide (CO). The effect of humidity has a profound impact on NOx emissions where NOx decreases exponentially with increase in specific humidity. This is due to humidity suppressing of the peak combustion temperatures. The impact of humidity on CO emissions is much smaller. Therefore NOx emissions increase with the decrease in ambient temperature during the period of constant EGT. Also, NOx emissions increase with the decrease in humidity. During constant power operation NOx emissions decreases with ambient temperature and is due to the decrease in combustion temperature during this period of operation. CO emissions decrease during constant EGT operation but increase during constant power operation. Carbon Dioxide (CO 2 ) emissions on a mass basis, follows the fuel flow 57

59 changes and increase during constant EGT operation while decreasing during constant power operation. CO 2 emissions on an index basis decrease as the thermal efficiency increases therefore decreasing during this transient. The CO 2 emissions index is defined as the mass of CO 2 produced per unit of gas turbine power output. Thus for a given fuel, it is directly proportional to the specific fuel consumption or heat rate. Since humidity decreases thermal efficiency the increase in humidity also increases the CO 2 emissions index, although this effect is small. Emission profiles are illustrated in Figure 6.5 and 6.6. Figure 6.6 shows CO 2 emissions when EGT or power limited The generator power demand of 60MW is well above the capacity of the gas turbine and would normally trip the engine due to the frequency shift in the electrical generator. Thus such simulators are invaluable as an operator tool in predicting maximum capacity of the gas turbine under different ambient conditions and degrees of engine degradation, hence optimising revenue and profit. We may elect to run the simulator at lower generator power demand, say, 30 MW, while subjecting the engine to the same ambient temperature transient discussed above. Some of the differences are that the fuel flow and temperatures decrease throughout the transient for this low power case provided the VIGV remains fully opened. This is because the increase in compressor airflow rate at lower ambient temperatures results in a decrease in specific work so maintaining the power demand of the generator at 30 MW. During this (low power setting) ambient temperature transient, and at ambient temperatures below -11 degrees Celsius the VIGV starts to close This is because the EGT falls below the set point (650K) for VIGV control and the VIGV closes to maintain the EGT at the set point. VIGV control is used to reduce starting power requirements by closing the VIGV at low power outputs thus reducing the mass flow rates through the compressor. This change results in reducing starting power requirements. The control of the VIGV is such that the EGT is maintained at a certain 58

60 value, in this case 650K. As the EGT falls below this set point, the VIGV closes, and when the EGT is above this value the VIGV is fully opened thus ensuring that the maximum power output of the gas turbine may be achieved. Thus there is a power output range where the EGT is maintained at the set point by modulating the VIGV. The decrease in specific work decreases the turbine entry temperature (T3) thus reducing the turbine metal temperature. This decreases the creep life usage significantly. The decrease in turbine entry temperature, and therefore the combustion temperature also results in a decrease in NOx emissions but an increase in the CO emissions. In the period when the VIGV closes to maintain the EGT, there is a slight increase in NOx but the CO emissions remain essentially constant. This is primarily due to a small increase in combustion temperature during the period when the VIGV is operating. 59

61 7.0 Effects of ambient pressure on gas turbine performance Another factor that affects engine performance is ambient pressure. A high ambient pressure increases the maximum power output of the gas turbine. This is primarily due to the increase in airflow through the gas turbine. The thermal efficiency is unaffected by the change in ambient pressure. Conversely, gas turbines operating at high elevations produce less power output because of the reduced ambient pressure. However, the ambient temperature at high elevations is usually lower than sea level, particularly in tropical countries and this partly compensates for the loss in the power output due to reduced ambient pressure. The lower ambient temperatures encountered at higher elevations improves the thermal efficiency of the gas turbine. At sea level, the ambient pressure can vary between 1.05 and 0.97 bar corresponding to a high pressure and low pressure day respectively. This presents about an 8% loss in power output. At high elevations the ambient pressure can be even lower. For example, at an elevation of 1000m the ambient pressure on an ISA (International Standard Atmosphere) day will be about 0.9 bar whereas at sea level the ambient pressure is bar. Therefore for a given ambient temperature, there is about a 12.5% lower power output will result when operating at an elevation of 1000m. When the engine is operating on a control limit such as EGT or power (e.g. high power output condition) the engine parameters such as pressures are affected by changes in the ambient pressure. However, the temperatures remain largely unaffected by ambient pressure. Furthermore, the pressure and temperature ratios and the nondimensional parameters such as mass flows are also unaffected by the change in ambient pressure at such high power output conditions. Hence the operating point on the compressor characteristic remains unaffected by the change in ambient pressure. The reduced power output from the gas turbine due to a decrease in ambient pressure results in reduced stress in the turbine components. Although the temperatures remain largely unaffected, the decrease in stresses in the turbine component results in a small decrease in creep life usage. The decrease in ambient pressure decreases the combustion pressure. This reduces the NOx emissions while increasing CO emissions. The decrease in fuel flow, and therefore the loss in power output, decreases the emissions of CO 2 on a mass basis. Since the thermal efficiency remains largely unaffected by ambient pressure changes the CO 2 emissions as an index also remains unaffected. The above discusses the impact of the change of ambient pressure on gas turbine performance when operating at the maximum power condition determined by an engine limit such as EGT or gas turbine power output. Gas turbines often operate at lower power outputs so these engine operating limits are never reached. We may simulate this case by setting the generator power demand to about 33MW. Under these operating conditions the performance of the gas turbine is very different. For example, the power demand from the generator is always met. However, the specific work increases to compensate for the decrease in airflow rate through the engine due to the decrease in ambient pressure. This results in an increase in the maximum to minimum cycle temperature ratio (T3/T1) and compressor pressure ratio. The increase 60

62 in these two parameters improves the thermal efficiency. Thus a lower ambient pressure is quite desirable when the power demand from the gas turbine is low and not constrained by the engine control system limiting parameters, such as EGT or power. In fact, this is indeed the principle of a closed cycle gas turbine, where we change the system pressure to change the load of the gas turbine so that the thermal efficiency remains constant with the decrease in engine load. The increase in the temperature ratio T3/T1 increases the turbine entry temperature (T3). This increases the turbine metal temperature and therefore the turbine creep life usage. The increase in combustion temperature when operating at lower power outputs will increase NOx emissions. However, decrease in ambient pressure and hence the decrease in combustion pressure will reduce NOx emissions. The increase in compressor pressure ratio suppresses the decrease in combustion pressure due to the lower ambient pressure. The net effect of these changes in combustion pressure and temperature is an increase in NOx emissions. However, the increase in combustion temperature decreases CO emissions as CO emissions are more sensitive to combustion temperature than pressure. The production of CO 2 decreases on both a mass and emission index basis and this is due to the increased thermal efficiency and operation at a constant power output. For a given ambient temperature, a decrease in ambient pressure will increase the specific humidity. However, the increase in specific humidity is too small to affect the results discussed above and we leave it to the reader to investigate the effects of humidity on gas turbine performance and emissions due to ambient pressure changes. The impact of auxiliary loads on engine performance due to ambient pressure changes can also be investigated using the simulator. Again, we leave it to the reader to run the necessary simulations to illustrate the impact of auxiliary loads on engine performance. This can be investigated by switching on the vapour compression chillers to augment the power output (due to turbine inlet cooling) and varying the ambient pressure while the chillers are active. 61

63 8.0 Simulating the effects of component deterioration on engine performance We have discussed in Chapter 4 that the performance of the gas turbine is primarily determined by the performance of the individual engine components such as the compressor, turbine, combustor, and the interaction between these components. The performance of these components and their interaction determine power output and thermal efficiency including parameters such as pressures, temperatures, flows and speeds. The performance of the individual components is represented by their component characteristics. For compressors and turbines these characteristics are represented by plotting the pressure ratio with the non-dimensional flow for a series of non-dimensional speeds. The component efficiencies, normally isentropic efficiencies, are also displayed on their characteristics as contours or plots. A typical compressor characteristic is shown in Figure 3.5 and typical turbine characteristics are shown in Figures 3.11 and As stated in Section 4.3, performance deterioration of the gas turbine results from the deterioration of one or more of the engine components, usually the compressors and turbines. Deterioration of these engine components result in changes in their characteristics. The interaction of these deteriorated component characteristics results in the change in power output, thermal efficiency and engine parameter such as pressure, temperatures, speeds and flows. Many factors affect the engine performance and listed in Section Compressor fouling Figure 8.1a shows the impact of compressor fouling on the compressor flow characteristic This is probably the most common form of performance deterioration and manifests itself by the ingestion of dirt, airborne debris, pollen, plant sap etc. Filtration is normally provided to trap them but can only arrest fouling, not prevent it. In certain environments, ingestion of sand can damage the filtration system and also result in severe erosion of the compressor. The effect of fouling on the compressor performance is to reduce the capacity of the compressor and its efficiency. Thus for a 62

64 given non-dimensional speed, compressor fouling decreases the non-dimensional flow and efficiency. The effect of compressor fouling on flow is usually greater than on efficiency. In effect, these lines of constant non-dimensional speed are shifted to the left on the compressor characteristic, as shown in Figure 8.1a. The effect of compressor fouling on compressor efficiency for a given speed is shown in Figure 8.1b. Figure 8.1b shows the impact of compressor fouling on compressor efficiency for a given speed We simulate compressor fouling by the use of fault indices. Fault indices represent the change in component characteristics from their design condition. To simulate moderate compressor fouling we set the compressor fouling fault index to -3% and its efficiency fault index to -1%. The fouling fault index effectively reduces the compressor flow capacity by 3% and simultaneously reduces the isentropic efficiency of the compressor by 1% for a given compressor non-dimensional speed line. Fouling occurs over a long period time, typically weeks, but we simulate fouling over a period of one hour by setting the ramp time to 3600 seconds. The use of these indices is described in the single shaft gas turbine simulator user guide. It should be noted that the simulator assumes fouling to increase linearly with time. In practice this is not the case, where the rate of fouling effectively decreases with time, as discussed in Section 4.3. When operating at high power demand the engine remains on an EGT operating limit at high ambient temperatures and power output limit at low ambient temperatures. The compressor discharge pressure and flow decrease during compressor fouling. The power loss due to moderate fouling is typically about 5% and the thermal efficiency decreases by about 1.5%. The turbine creep life usage also decreases during fouling. But this is somewhat misleading because there is also a loss in power output. It is when we run the simulator at this reduced power output due to fouling, but when fouling is absent, that we see an increase creep life usage in real terms. 63

65 A similar picture exists regarding gas turbine emissions where NOx decreases during compressor fouling. However, it is when we run the simulator at the reduced power level, but when no fouling is present, that we see an increase in NOx and CO 2 due to compressor fouling. CO emissions decrease during compressor fouling and this is primarily due to the higher combustion temperature in real terms during compressor fouling. We may plot the compressor inlet non-dimensional flow varying with its nondimensional speed with, and without, fouling. A downward shift in this running line during compressor fouling is observed and this is generally a good indicator of compressor fouling, as shown in Figure 8.2. It should be noted that the shift in the running line is not affected a great deal by other faults in the engine and will be discussed later in this section. We can therefore use such a curve to detect compressor fouling. Figure 8.2 shows the effect of fouling on the variation of compressor non-dimensional flow with its speed The correct definition of non-dimensional flow and speed are W( RT/γ)/P and N/( γrt) as they allow for the change in the gas properties of air with ambient temperature and humidity. R and γ correspond to the gas constant and the ratio of specific heats (cp/cv) respectively. These issues are discussed in detail in [Walsh and Fletcher -200, and Razak -2007]. However, for a first order approximation we can use the quasi non-dimensional parameters as shown in Figure 8.2. The reader should plot Figure 8.2 using the correct non-dimensional parameters and compare it with the case when using quasi non-dimensional flows and speeds. The above considered the impact of compressor fouling on engine performance at high power outputs. We can run the simulator with the same level of compressor fouling (i.e. the same compressor fault indices as above) but at a low enough power (say 35MW) such that no engine operating limits are reached during fouling. The change in air flow is similar and this is due to the constant gas turbine speed which is 64

66 determined by the synchronous speed of the generator. As a result the movement of the operating point on the compressor characteristic due to fouling is similar to the high power operating case discussed above. An increase in turbine entry temperature is necessary to compensate for the loss in power output due to the loss in compressor efficiency and airflow rate through the compressor due to fouling. We may plot the operating point on the curves describing the variation of compressor non-dimensional flow with its non-dimensional speed. Since the level of compressor fouling is the same for the high and low power output cases, the shift in this running line is unchanged. Due to the constant gas turbine speed, the operating points for these respective cases are also similar. Compressor washing mitigates compressor fouling and the shift in these running lines may be used to determine compressor washing frequency. Optimising the frequency of compressor washes is more complex and has to take into account many factors such as downtime for washing, fuel cost and revenue. Generally if the power demand is below the maximum available from the engine, the compressor wash frequency can be decreased. However, if the power demand is high and limited by the engine operating limits such as EGT or power, the wash frequency should be increased. This is because the loss in revenue due to fouling is significant in the high power case, while the low power case, the increase in cost of fuel due to the loss is thermal efficiency is only marginal, compared with the lost revenue due to downtime required for washing. If standby power is available or on-line wash facilities are available then higher wash frequencies may be considered as the lost revenue, due to downtime, will be small. At low generator power demand (35MW) the increase in turbine entry temperature results in a significant increase in creep life usage of the turbine. Similarly, there is also an increase in NOx and CO 2 emissions. However, the higher combustion temperatures result in a slight decrease in CO emissions. For the low power operating case, we may elect to operate the gas turbine where we maintain the EGT limit at lower power outputs. This is achieved by setting the EGT set point for VIGV control to correspond to the maximum EGT limit (i.e. increase the EGT set point from 650K to 825K - use control system option 2). The decrease in airflow through the compressor due to fouling results in an increase in the EGT. The VIGV control system responds to the high EGT by opening the VIGV thus maintaining the EGT on the set point. The opening of the VIGV tends to compensate for the decrease in compressor flow capacity due to fouling and also improves the compressor efficiency. Thus, when we operate the gas turbine at low power conditions with the VIGV controlling the EGT at maximum or limiting value, compressor fouling has only a minimum effect on performance deterioration. It must be pointed out that the compressor characteristic changes significantly with the change in VIGV position. Thus, it is not possible to detect compressor fouling by using the displacement of the curve describing the compressor non-dimensional flow with its non-dimensional speed for this case. 65

67 8.2 Turbine damage Turbines are exposed to very high temperatures and often employ turbine cooling in order to achieve satisfactory life, as discussed in Section However, prolonged exposure of the turbine to high temperature combustion gases can cause damage in this component. The damage can manifest itself over a short period of time, particularly if the engine suffers from combustion problems thus producing very hot streaks of gases or carbon due to incomplete combustion which can easily destroy the turbine. Such damage is usually referred to as hot engine damage and results in an increase in turbine capacity, as shown in Figure 8.3. A reduction in turbine isentropic efficiency also occurs due to hot end damage. Turbine fouling is unusual, but burning low grade fuels which have a high content of ash can foul turbines. During starting, high vibration can occur as the engine passes through critical speed resulting in rubs of the rotating members of the turbine with the casing. Over a period of time the clearances will increase thus increasing the leakages and hence affecting the engine performance adversely. Blade rubs normally affect the turbine efficiency rather than the flow capacity. It must be pointed out that such damage also occurs in compressors. The impact of rubs on flow and efficiency depends on which compressor stages have rubbed. As stated in Section 4.3, if the LP stage has rubbed then the flow capacity and efficiency of the compressor will be affected, but if the HP stages have rubbed, only efficiency that is affected. The reason for this is that the LP or the front stage of the compressor controls the flow through the compressor at normal operating speeds. Figure 8.3 shows the increased turbine capacity due to hot end damage It has been stated that hot end damage results in an increase in the turbine (nondimensional flow) capacity and a decrease in the turbine isentropic efficiency. Hot end damage to the turbine may be simulated by increasing its fouling fault index to 3% while decreasing its efficiency fault index by 2%. At high ambient temperatures and power demand, the gas turbine power output is limited by the EGT rather than the power limit. The increase in turbine capacity decreases the compressor pressure and thus the turbine pressure ratio. This results in a decrease in the turbine entry temperature (T3) due to a constant EGT operation. The decrease in turbine entry 66

68 temperature decreases in gas turbine specific work, thus power output. The decrease in compressor pressure ratio decreases the thermal efficiency. The turbine creep life usage decreases and this is due to the decrease in turbine entry temperature. Hot end damage decreases the turbine entry temperature and compressor pressure ratio. This results in a decrease in the combustion temperature and pressure. Thus the NOx emissions decrease while CO emissions increase. Although the thermal efficiency decreases, the decrease in power output is greater than the decrease in thermal efficiency. This results in a decrease in fuel flow and therefore CO 2 emissions on a mass basis. Again, these changes in creep life usage and emissions are somewhat misleading. When the simulator is run at the reduced power conditions due to turbine damage but when no damage is present, then an increase in creep life usage and emissions in real terms is seen. At low ambient temperatures, the power output from the gas turbine is limited by the maximum power limit. Provided that the increase in EGT due to hot end damage does not reach the EGT limit no loss in power output will occur. However, the decrease in the compressor pressure ratio, turbine entry temperature and turbine efficiency results in a loss in thermal efficiency. The increase in EGT and thus turbine entry temperature results in an increase in turbine creep life usage. The decrease in compressor pressure ratio and thus combustion pressure due to hot end damage will normally reduce NOx emissions. But the increase in turbine entry temperature and thus combustion temperature is sufficient to increase NOx emissions due to hot end damage. CO emissions remain essentially constant. CO 2 emissions on a mass flow and emissions index basis increase and this is due to the decrease in thermal efficiency. We have not altered the compressor characteristics during this simulation. Thus there is no shift in the curve describing the variation of compressor inlet non-dimensional flow with its speed. Any change in the operating point on this curve due to ambient temperature changes will be along this line corresponding to the clean compressor, as shown in Figure 8.2. This implies that there is no compressor fouling is present during this fault simulation. Turbine blade rubs on engine performance can also be simulated. This fault is simulated by only decreasing the turbine efficiency fault index. The behaviour of the gas turbine with respect to its performance, turbine creep life usage and emissions due to turbine blade rubs is very similar to that discussed above. When the VIGV is active during the normal operating power range, the impact of hot end damage on engine performance at high ambient temperatures is no different to that discussed above. It is only when the ambient temperature is low and the performance of the gas turbine is limited by the power output rather than the EGT that the performance is different. Under these conditions the VIGV will not be fully opened. The effect of hot end damage will increase the EGT and therefore open the VIGV in order to maintain the EGT on its set point. The opening of VIGV increases the compressor airflow rate and thus the specific work has to decrease to maintain the power output of the gas turbine at its limiting value. The decrease in specific work is achieved by decreasing the maximum to minimum cycle temperature ratio (T3/T1) which results in a decrease in turbine entry temperature (T3). The decrease in T3/T1 67

69 and turbine efficiency is sufficient to decrease the thermal efficiency of the gas turbine. The decrease in the turbine entry temperature results in a decrease in turbine metal temperature and therefore reduces the turbine creep life usage. In this case hot end damage results in a noticeable decrease in creep life usage in real terms. Emissions of NOx and CO emissions remain essentially constant and this is due to the air-fuel ratio remaining approximately constant. Such VIGV control is quite common in single shaft gas turbines employing dry low emission (DLE) combustion systems. In this case the control of the VIGV system maintains the fuel-air ratio thus keeping the emissions of NOx and CO approximately constant at different loads and ambient conditions. However, the emissions of NOx levels are significantly lower and this is due to the DLE combustion process maintaining lower maximum combustion temperature at about 1800K. This results in NOx emissions in the order of about 10 to 25 ppmv. Although the DLE combustion process operates at low temperatures, the system is also designed to maintain low CO emissions. It should be pointed out that the simulator assumes a non-dle combustion system (diffusion flame) and therefore gives higher values for NOx emissions. The decrease in thermal efficiency results in the increase in CO 2 emissions on a mass and index basis. 68

70 9.0 Power augmentation For a given engine design the maximum power output of the gas turbine is limited. This is necessary to prevent the engine from overheating thus protecting the engine from damage and achieving the required creep life from the turbine. As discussed in Chapter 5, various engine parameters limit the power output of the gas turbine depending on the ambient temperature. At high ambient temperatures, the exhaust gas temperature (EGT) limits the power output, while at low ambient temperatures a power limit is imposed on the gas turbine. However, manufacturers may allow operators to increase the power output from their gas turbines for limited periods of operation where the engine power output can be increased by a certain amount. However, the engine life will be reduced due to the increased turbine operating temperature. 9.1 Peak rating At high ambient temperatures the EGT limit can be raised to increase the power output of the gas turbine. Increasing the EGT limit by about 20 to 40 degrees Celsius achieves about a 5% to 10% increase in power output. A worthwhile increase in thermal efficiency will also result. The increase in EGT limit increases the turbine entry temperature (T3) and compressor pressure ratio. In a single shaft gas turbine no increase in compressor airflow rate occurs due to the constant speed of the gas turbine. Thus all the increase in power output due to peak rating results from the increase in specific work. Figure 9.1 shows the increased in power output under peak rating conditions A graph describing the maximum power output and thermal efficiency of the gas turbine at different ambient conditions for both base and peak rating is shown in Figures 9.1 and 9.2 respectively. We observe that peak rating is most effective at high 69

71 ambient temperatures and will have no effect at low ambient temperatures because at these ambient temperatures the power output limits the performance of the gas turbine. Figure 9.2 shows the increase in thermal efficiency under peak rating conditions Turbine creep life usage increases and is primarily due to the higher turbine entry temperature, leading to higher turbine blade temperatures. Peak rating could double the turbine creep life usage as shown in Figure 9.3 where the EGT limit is increased by 20 degrees Celsius. Figure 9.3 shows the increase in turbine creep life usage due to the effect of peak rating 70

72 Gas turbine emissions such as NOx increase and this is due to the higher combustion temperatures and pressures. CO decreases because the increase in these parameters reduces CO emissions. CO 2 emissions increase on a mass basis because the increase in power output is greater than the increase in thermal efficiency. The higher thermal efficiency reduces the CO 2 emissions index, thus in real terms peak rating reduces CO 2 emissions. Changes in emissions due to peak rating are illustrated graphically in Figures 9.4 and 9.5. Figure 9.4 shows the effect of peak rating on emissions Figure 9.5 shows the effect of peak rating on CO2 emissions 71

73 9.2 Power augmentation by water injection Gas turbine power output may also be augmented by water injection. Water may be injected into the inlet of the compressor thereby reducing the compressor inlet temperature and so increasing the power output of the gas turbine. Alternatively, the power output of the gas turbine may also be augmented by directly injecting the water into the primary zone of the combustion chamber. Augmenting the power by injecting water into the engine inlet is often referred to as turbine inlet cooling. Turbine inlet cooling is most suitable when the humidity is low because the amount of water that can be evaporated is determined by the humidity of the inlet air. Technologies suitable for turbine inlet cooling, including chillers will be discussed in Section 9.3. Direct water injection into the combustor is not limited by the humidity of the air and the amount of water injected is often determined by the increase in CO emissions due to the suppression of the flame temperature in the primary zone. Figure 9.6 shows the increase in power output due to the effect of water injection We have stated that direct water injection increases the power output of the gas turbine and this is due to the increased gas flow through the turbine. The increase in gas flow through the turbine also increases the compressor pressure ratio. When operating at the EGT limit, the increase in the compressor ratio increases the turbine entry temperature which also contributes to the increase in power output of the gas turbine due to water injection. The thermal efficiency on the other hand decreases quite substantially and this is due to the additional heating required, thus increased fuel flow, to evaporate the water and heat it to the turbine entry temperature. As the latent heat needed to evaporate the water cannot be recovered by doing work the thermal efficiency decreases. Power augmentation due to direct water injection in to the combustor is most effective at high ambient temperatures where the power output of the gas turbine is limited by the EGT. A plot of the power output varying with ambient temperature with and without water injection is shown in Figure 9.6. The 72

74 water fuel-ratio of unity is used in this example. It will also be noticed that the maximum power limit will be reached at a high ambient temperature when using water direct injection. The variation of thermal efficiency with ambient temperature due to water injection is shown in Figure 9.7. Figure 9.7 shows the decrease in thermal efficiency due to the effect of water injection Figure 9.8 shows the changes in turbine creep life usage due to water injection At ambient temperatures when the EGT limits the power output of the gas turbine, water injection increases the turbine creep life usage. This is due to the increase in turbine power output increasing the stresses in the turbine blades, thereby increasing the turbine creep life usage. The increase in turbine entry temperature also increases the turbine blade temperature and therefore contributes to the increase in turbine creep 73

75 life usage. The change in turbine creep life usage due to water injection is shown in Figure 9.8. Figure 9.9 shows the change in NOx and CO emissions due to water injection Figure 9.10 shows the increase in CO2 due to water injection Direct water injection suppresses the flame temperature in the primary zone and a significant reduction in NOx emissions occurs. However, the suppression of the flame temperature also increases CO emissions and is often a limiting factor on how much water injection that can be used to augment the gas turbine power output. CO 2 emissions on a mass and index basis increase and this is due to the decrease in the gas turbine thermal efficiency. Changes in emissions due to water injection are illustrated in Figures 9.9 and Figure

76 9.3 Turbine inlet cooling We have observed the adverse effect high ambient temperature has on power output and thermal efficiency. We have also discussed means to improve gas turbine power output at high ambient temperatures using peak rating and water injection. However, they invariably have some disadvantages such as increased creep life usage and lower thermal efficiencies when direct water injection is employed. As stated in Section 9.2, we can cool the turbine inlet thereby reducing the compressor inlet temperature to augment the power output at high ambient temperatures. This is referred to as turbine inlet cooling or TIC. There are two main technologies available to reduce the compressor inlet temperature and they are known as evaporative cooling (wetted media and inlet fogging) and chilling. Wetted media cooling and fogging operate on the same principles where the evaporation of water absorbs latent heat of evaporation thus cooling the turbine inlet air. With wetted media, the media is saturated with water and is exposed to the compressor inlet air and the resultant evaporation reduces the compressor inlet temperature thus increasing the gas turbine power output. Alternatively, we can introduce the water into the inlet as a very fine spray. The resultant evaporation cools the compressor inlet air. The requirements for water quantity for evaporative cooling, is discussed in Another useful website for further information is Unlike media evaporative cooling, which can operate with raw water, fogging systems require demineralised water. It should be noted that demineralised water is quite aggressive and will attack certain metal and the inlet systems should use materials such as stainless steel which are resistant to attack from demineralised water. With media cooling using raw water, sufficient amount of water recirculation is necessary to prevent the concentration of minerals and salts in the evaporative media. Increased concentration of such minerals and salts will damage the wetted media resulting in loss of cooling effectiveness. The presence of the wetted media also increases the pressure loss in the inlet system, which does not occur with fogging systems. The amount of water that can be evaporated depends on the relative humidity of the air. The lower the humidity more water can be evaporated and results in a greater degree of turbine inlet temperature cooling. Another factor that limits the amount of cooling is the effectiveness of the cooling system. The more efficient the cooler (i.e. greater the effectiveness) closer does the dry bulb temperature, which is effectively the ambient temperature, approaches the web bulb temperature. For evaporative cooling systems using wetted media the effectiveness can vary from 0.85 to 0.95, which is measure of the difference between the final dry bulb and wet bulb temperature. The effectiveness of evaporative cooling is given by: Ta T1 ε =..9.1 T 1 Tw Where Ta is the ambient temperature (also known as the dry bulb temperature), T1 is the compressor inlet temperature and Tw is the wet bulb temperature. For fogging system the effectiveness can approach unity. In this case the dry bulb temperature approaches the wet bulb temperature and compressor inlet air will be 75

77 saturated. In fact evaporative cooling occurs at a constant wet bulb temperature. Therefore if the ambient wet bulb temperature is close to the dry bulb temperature (i.e. high relative humidity) only little turbine inlet cooling is possible. Another evaporative cooling technique is wet compression or overspray. Here additional water is added as a fine spray directly into the inlet of the compressor. This water evaporates in the compressor due to the high temperatures that occur during adiabatic compression, thus cools the air within the compressor. This is similar to isothermal compression but to a less agree and reduces the compression power demand therefore increasing the power output of the gas turbine. Hence this method of power augmentation is also referred to as fog intercooling. Since water is added directly into the compressor as a spray there is an increased risk of compressor damage due to erosion which can result in compressor surge and damage the engine severely. As with fogging system the demineralised water should be used for wet compression. Figure 9.11 shows the amount of turbine inlet cooling with ambient temperature using different cooling technologies The cooling process using evaporative cooling is adiabatic. With inlet chilling we remove heat from the inlet air using some form of refrigeration. Thus inlet chillers are not limited by the wet bulb temperature and the compressor inlet air can be cooled down to any desired temperature provided the cooling capacity is available. However, when the inlet temperature decreases below 10 degrees Celsius there is an increased risk of ice formation in the inlet, which can break away and enter the engine thereby damaging the engine. Thus turbine inlet cooling, whether evaporative or chilling is limited to compressor inlet temperatures about 10 degrees Celsius. Refrigeration systems for chillers can be either vapour compression or vapour absorption. The power demand from vapour compressor systems is significant and referred to as parasitic losses. In spite of such losses there is still a significant gain in engine performance, particularly at very high ambient temperature and low humidity. Absorption refrigeration systems require a heat source, which can be provided from 76

78 waste heat. Thus their parasitic losses are very small; however their performance is much less that vapour compression systems. If waste heat is readily available the poor performance of vapour absorption refrigeration systems is of little consequence. Other sources for chilling include LNG evaporation systems where the turbine inlet air is used as a heat source for evaporation of LNG. Figure 9.11 shows the turbine inlet cooling using three types of cooling technologies. The most significant cooling is achieved using chillers followed by fogging and wetted media cooling. This figure has been produced assuming a constant relative humidity of 60% and we observe that the amount of cooling increases with ambient temperature. With wetted media and fogging the potential to cool the compressor inlet decreases with ambient temperature due to the divergence of the lines on constant relative humidity on the psychrometric chart, as shown in Figure 4.6 above. With chillers, this decrease is more acute as seen in Figure We have assumed the maximum cooling capacity of 9MW, which is sufficient to maintain the compressor inlet temperature at 10 degrees Celsius for the ambient temperature range considered in Figure The humidity level has a significant impact on the level of cooling that can be achieved, particularly with wetted media and fogging. Lower the humidity greater is the cooling that can be achieved. The reader should reproduce this figure for various levels of relative humidity. Figure 9.12 shows the generator power output with ambient temperatures using various turbine inlet cooling technologies The decrease in compressor inlet temperature, due to turbine inlet cooling, increases the power output of the gas turbine and thus the generator power output. This is shown in Figure 9.12 for each of the above mentioned cooling technologies. It can be seen that chillers produces the largest gain in power output followed by fogging and wetted media. The figure also shows the effect of the cooling technology employed by chillers on increased power output. With vapour compression chillers the impact of power demand by the chiller on generator output is significant and appears as a parasitic loss. Nevertheless, there is a significant increase in power output at the 77

79 generator terminals. With vapour absorption chillers this parasitic loss is small and is ignored by the simulator as the simulator represents an open cycle gas turbine and the heating required by the absorption refrigeration system can be provided from the gas turbine exhaust heat. The decrease in compressor inlet temperature due to turbine inlet cooling also increases the thermal efficiency of the gas turbine, which the user should produce. Increase in compressor airflow and therefore increased exhaust flow rate also increases the heat rejection from the gas turbine and is beneficial to combined cycle power plants. Turbine inlet cooling using wetted media and fogging requires water flow for evaporation and shown in Figure The simulator is a useful calculator to determine the cooling water requirements for a given level of turbine inlet cooling. The information can be scaled to suit any gas turbine engine see simulator user guide for further details. With chillers, condensation can occur when the relative humidity reaches 100% due to cooling, as shown in Figure At high ambient temperatures and humidity the condensation can be about three times that required by evaporative cooling systems. For example, at an ambient temperature of 30 degrees Celsius the condensation flow rates can be as much as 100 tonnes per day. Thus means to remove this water must be provided when chillers are employed. The cooling loads and power demand for chillers using vapour compression systems is shown in Figure We note the cooling loads and power demand increases exponentially when condensation occurs. This is primarily due to the absorption of the latent heat of evaporation of water by the chillers. As an illustration, when turbine inlet cooling of about 8 degrees Celsius is require the cooling loads and power demand from the vapour compression chillers is about 1MW and 0.23MW respectively (assuming a coefficient of performance of refrigeration of 5.0). If we double the turbine inlet cooling to 16 degrees Celsius, the cooling load and power demand from the chillers increases to nearly 6MW and 1.2 MW respectively and this is due to the formation of condensation at the higher levels of turbine inlet cooling. Evaporative cooling increases the relative and specific humidity at the compressor inlet, as illustrated in Figure It should be noted that the wet bulb temperature remains constant. The ambient temperature and relative humidity are held constant at 30 degrees Celsius and 60% respectively. The increase in specific humidity suppresses the increase in NOx emissions which normally occurs due to the increase in combustion temperature and pressure as the compressor inlet temperature decreases. 78

80 Figure 9.13 shows the water requirements for turbine inlet cooling systems Figure 9.14 shows the cooling loads and power demand for vapour compression refrigeration system due to turbine inlet cooling 79

81 Figure 9.15 shows the running line on the psychometric chart for various turbine inlet cooling technologies With chillers, the relative humidity increases and the wet bulb temperature decrease while the specific humidity remains constant provided no condensation occurs. When relative humidity reaches 100%, any further cooling will produce condensation and this decreases the specific humidity, as shown in Figure Thus the decrease in specific humidity and the increase in combustion temperature and pressure result in an increase in NOx emissions. These effects can be simulated and the reader should use the simulator to illustrate these issues. As with wetted media cooling and fogging, it should be noted that the simulator is a useful calculator to determine the chiller cooling load requirements and produced water (condensation) for a given level of turbine inlet cooling. The information can be scaled to suit any gas turbine both single and multi-shaft engines, as discussed in the simulator user guide. 80

82 10.0 Simulation of engine control system performance As stated in Section 5.0, the gas turbine power output is determined by the thermal input which is achieved when burning fuel in the combustion system. The higher the fuel flow, the higher is the power output from the gas turbine. However, high turbine temperatures must be avoided as this will result in a significant loss in turbine creep life or even engine failure. It is therefore the function of the engine control system to ensure that the required power from the gas turbine is achieved safely. In general there are two groups of control systems that can be employed and they correspond to the open loop and closed loop control systems. In an open loop control system, the input (fuel flow) to the control system is independent of the output (gas turbine power output). The input usually acts for a period of time after which the output is expected to have reached the required set point. This corresponds to the required power demand from the load (generator). In such a control system the output seldom reaches the set point and the control system usually leaves an offset between the output and the set point. In a closed loop control system the offset left by the open loop control system is used as the input to the closed loop controller to generate the output. This is referred to as negative feed back. By such means it is possible to eliminate the offset and the control system output will then correspond to the set point. In a closed loop control system the offset is converted to an error which is calculated as the percentage deviation of the output from the set point and used as the input to the controller PID Loop A closed loop control system normally achieves the output using a proportional (P), integral (I) and derivative (D) action or a PID loop or controller, also known as a three-term controller, and discussed in Section 5.0. Further details are available in [Shaw ] Proportional action The proportional action generates an output from the control system which is proportional to the error plus a bias. Therefore, proportional-only control systems leave a difference between the power output from the generator and the generator power demand (set point). Proportional only action can be simulated by switching off the integral action, as described in the user guide (single shaft gas turbine simulator). This action will leave an offset between the generator power output and the generator set point. The generator power demand should be set to 39MW before starting this simulation. Reducing the proportional band (i.e. increasing the proportional gain) will reduce the proportional off-set Proportional and integral action The integral action occurs as a result of the error being continuously integrated or summed up. Thus the proportional offset is eliminated when both proportional and integral control is employed and there is no need for manually resetting the proportional offset. Integral action is also referred to as automatic reset. This can be 81

83 illustrated by repeating the simulation exercise discussed in Section and switching the integral output button on and off. The addition of integral action can also result in drawbacks and this is referred to as integral windup. Windup can occur when the conditions are such that the output from the process (in this case the gas turbine power output) is unaffected by the controller action. For example, it occurs if the fuel valve becomes fully opened before the power demand from the gas turbine is reached. This can be due to insufficient fuel valve size, as discussed in Section 5.1. We can simulate integral windup using the simulator by switching off the reset windup and setting the fuel valve capacity to half its design value. This will ensure that the fuel valve will open fully at high gas turbine power demand. A similar exercise may be carried out for the VIGV control system using either control strategy, as described in Section 10.5 below and the simulator user guide. The user may also alter control setting (proportional band and integral gain) to investigate the transient response of the gas turbine. For example, increasing the proportional band and integral gain will result in an oscillatory response from the gas turbine. Conversely, decreasing these values will result in a sluggish response from the gas turbine. (Note: - A low proportional band (high gain) may result in the control system becoming unstable. The model may suffer from numerical instabilities) Proportional, integral and derivative action The addition of derivative action enhances the controller output during transients. It is normally used when the response of the system is very slow (e.g. furnaces), but is often omitted in gas turbine control systems. As stated in Section 5.1, derivative control produces no action when a steady state error occurs due to the proportional offset or integral windup as the rate of change of the error under these conditions will be zero Signal selection We have discussed above in Section 6.1 that the engine performance is limited by various parameters such as exhaust gas temperature (EGT) at high ambient temperatures and power output at low ambient temperatures. The control system achieves this by using signal selection where further errors are calculated using the limiting conditions such as EGT and maximum power limit as set points. These errors are compared with each other including the error using the generator power output and the required generator power demand (set point). The lowest or smallest error is used by the PID controller, thus ensuring that none of the limiting parameters (EGT and powers) are exceeded. This approach is referred to as low signal selection. High signal selection is also employed by engine control systems and is often used to ensure power reduction without encountering trips due to flame-out, which can result if the air-fuel ratio exceeding the weak extinction limit. 82

84 10.3 Optimising EGT We have assumed that the EGT limit remains constant with the variation of the ambient temperature (fixed temperature control). With single shaft gas turbines, the EGT is usually measured at the exit of the turbine rather than at some intermediate point between the inlet and exit of the turbine. The increase in ambient temperature results in a decrease in the compressor pressure ratio and thus turbine pressure ratio. For a fixed EGT limit, this results in a decrease in the turbine entry temperature and hence a decrease in turbine creep life usage as shown in Figure It is therefore possible to increase the EGT limit at higher ambient temperatures to improve the power output but decrease the EGT limit at low ambient temperatures to trade-off some of the life from the low ambient temperature operation to high ambient temperature operation such that the overall turbine creep life is unaffected. Such trade-off is often referred to as rating curves and each operating site can be optimised for maximum performance. The increase in the EGT limit at high ambient temperatures will increase the power output of the gas turbine at these ambient temperatures, as shown in Figure 10.2 (variable temperature control). At an ambient temperature of 30 degrees Celsius, the gas turbine power output increases by about 4% compared with the case when the EGT limit remains constant with ambient temperature changes. There is also a useful increase in the thermal efficiency of the gas turbine due to the higher EGT limit. Figure 10.1 shows the effect of fixed and variable exhaust gas temperature (EGT) control on turbine creep life usage with ambient temperature 83

85 Figure 10.2 shows the change in gas turbine power output for a fixed and variable EGT control with ambient temperature 10.4 Trips We have stated that the engine power output is limited by the EGT or power during steady state operation. However, during transient operation these limiting values may be exceeded resulting in excursions in temperatures and powers. Such excursions may occur due to poor engine control setting or improper operation of the gas turbine. Thus, trip values are provided to protect the engine under such extreme transient conditions. Trip values are usually set at about 5% to 10% above the limiting values for temperatures and powers. We may simulate such trips due to either poor control system settings or improper operation using the simulator. With respect to improper operation, we may consider inject water to augment the power output of the gas turbine as discussed in Section 9.2 above using a water-fuel ratio of 2. If water injection is cut off rapidly, (cease water injection using a ramp time of 1 second) the EGT exceeds the trip level as the control system cannot effectively respond to such a rapid transient, thus tripping the engine (the generator power demand should be set to 43MW) VIGV Control Single shaft gas turbines often employ variable inlet guide vanes (VIGV) or variable stator vanes (VSV) to reduce starting powers and also to maintain exhaust gas temperature at low powers. Maintaining exhaust gas temperatures at low powers improves the off-design thermal efficiency when used in a combined cycle plant or a regenerative cycle. The use of VIGV and VSV also maintain the air-fuel ratios at low power conditions and therefore is employed in dry low emission (DLE) combustion systems. 84

86 Two possible strategies for the controlling of the VGV system exist. In the first approach, the VIGV is modulated to maintain the exhaust gas temperature, while the fuel flow is modulated to maintain the power output of the gas turbine (control system option 1 in the user guide). Such a strategy is suitable when VIGV modulation is used to reduce starting powers. In this case the EGT set point for VIGV control is below the maximum EGT limit. Thus at starting conditions and low power operation the EGT will be below the EGT set point for VIGV control and the VIGV will be fully closed. This will reduce compressor airflow rate so reducing starting power requirements. As the power output from the gas turbine increases the EGT will exceed the EGT set point for VIGV control and the VIGV will be fully opened, thus helping to achieve the gas turbine design power output. When the gas turbine is part of the combined cycle plant, maintaining the EGT at its maximum value at part power conditions will improve the part-load thermal efficiency of the combined cycle plant. It may seem that this can be achieved using the VIGV control strategy described above by setting the EGT set point for VIGV control to correspond to the maximum EGT limit. However, as the power output increases from idle the VIGV remains closed until the EGT exceeds the EGT set point for VIGV control. As the EGT reaches the maximum EGT value, which is now the EGT set point for VIGV control, low signal selection will prevent the fuel valve from opening thus preventing the turbine from overheating. This, in effect, prevents the VIGV from opening fully, thus preventing the engine from developing the designed power output. An open loop response can be included to improve the situation. This can be achieved by opening the VIGV fully for a given period of time when ever the power demand from the generator is increased. The fully opened VIGV will now reduce the EGT and therefore allowing the fuel flow to increase, thus helping to achieve the required power demand from the gas turbine. After the period of open loop control the VIGV control system switches to closed loop control and VIGV now closes to maintain the EGT at the control set point, which now corresponds to the maximum EGT limit. Normally, the EGT set point for VIGV control is set slightly lower than the maximum EGT limit to ensure that the VIGV is fully opened during maximum power operation is required. Although the above VIGV control strategy works well, the air-fuel ratio increases during the period of open loop response. When such a VIGV control strategy is used in DLE gas turbines an increase in air-fuel ratio may exceed the weak extinction limit resulting in flame out and therefore will trip the engine. An alternative strategy for VIGV control (control system option 2 in the user guide) is to modulate the VIGV to maintain the power demand from the gas turbine and to modulate the fuel flow to maintain the EGT on the set point. Such a control strategy eliminates the need for open loop response from the VIGV during transients and maintains the air-fuel ratio without resulting in trips. 85

87 Part Two-shaft gas turbine operating with a free power turbine performance and operation In this section we shall use the two-shaft gas turbine simulator to illustrate the performance and operation of a two-shaft gas turbine operating with a free power turbine. Their use is widespread in mechanical drive applications and examples are gas compression and gas transmissions. They are also used in industrial power generation and for peak lopping. They are also used in naval and marine propulsion. The two-shaft gas turbine simulator simulates an advanced aero-derived industrial gas turbine having an ISO rating of about 20 MW. The simulator assumes that the driven load is an electrical generator. Again, the reader is encouraged to run the necessary simulations discussed and produce the graphical displays describing the engine performance, turbine creep life and emissions under various operating conditions. The two-shaft gas turbine simulator user guide gives details on how to operate the simulator. Although some generalisation is possible regarding the performance and operability of two-shaft gas turbines operating with a free power turbine using the simulator, however, the discussion is strictly applicable to the engine described by the simulator Effects of ambient temperature on gas turbine performance As discussed in Section , ambient temperature has a significant impact on gas turbine performance. Generally, higher the ambient temperature, the lower is the power output and thermal efficiency. Conversely, the lower the ambient temperature the better is the engine performance. The better gas turbine performance at low ambient temperatures is due to the increase in the maximum to minimum cycle temperature ratio (T3/T1), compressor pressure ratio and airflow rate. However, at low ambient temperatures, the engine performance of a two-shaft gas turbine operating with a free power turbine may be limited by the gas generator speed or compressor non-dimensional speed (N1/ T1) rather than the exhaust gas temperature (EGT), which limits the performance of the gas turbine at high ambient temperatures. This differs from a single shaft gas turbine, as discussed in Part 2, where a maximum power limit is usually imposed on the engine. The simulator may be used to illustrate these effects due to ambient temperature changes on gas turbine performance. The ambient temperature may be varied from 30 degrees Celsius down to -30 degrees Celsius over a large enough time period, say, one hour (ramp time=3600 seconds). The power demand from the generator should be maintained at 25 MW thus ensuring that the engine is always on a limiting condition during the ambient temperature transient. The change in gas turbine/generator power output and thermal efficiency may be plotted using the output data from the simulator. These plots (Figures 11.1 and 11.2), which the user should produce, shows the power output increasing as the ambient 86

88 temperature decreases while the EGT limits the performance of the gas turbine (Figure 11.1). Although the decrease in ambient temperature results in a decrease in the specific work of the gas turbine, the increase in maximum to minimum temperature ratio (T3/T1) increases the specific work. For a given value of T3/T1, the increase in compressor ratio has little effect on the specific work due to the very high compressor design pressure ratios of this engine (at lower design compressor pressure ratios the power output increases with increase in compressor pressure). The increase in compressor non-dimensional flow increases the airflow through the engine. The effect of these factors increases the power output of the gas turbine during the period of engine operation when the engine performance is limited by the EGT. The thermal efficiency of the gas turbine increases and this is primarily due to the increase in both compressor pressure ratio and T3/T1, as shown in Figure Figure 11.1 shows the impact of ambient temperature on gas turbine power output As the ambient temperature decreases below 15 degrees Celsius the gas turbine performance is limited by the gas generator speed and the power output from the gas turbine remains relatively flat. During this period of operation the rate of increase in T3/T1 and the compressor airflow decreases. The decrease in ambient temperature and thus the compressor inlet temperature is sufficient to decrease the specific work. Although the airflow rate increases during this period of operation, the increase in the power output is smaller. The compressor pressure ratio also increases during constant gas generator speed operation but at a decreasing rate. The increase in the compressor pressure ratio and T3/T1 increases the thermal efficiency of the gas turbine during constant gas generator speed operation. At ambient temperatures below -11 degrees Celsius the gas turbine performance is limited by the compressor non-dimensional speed (N1/ T1). During this period of operation T3/T1 and compressor pressure ratio remains approximately constant. Thus the fall in specific work is more acute as the ambient temperature decreases below -11 degrees Celsius. Although the compressor non-dimensional speed and thus the compressor non-dimensional flow remain constant, the decrease in ambient temperature during this period of operation increases the compressor airflow rate. 87

89 However, the loss in specific work is large enough to decrease the power output from the gas turbine as seen in Figure Figure 11.2 shows the impact of ambient temperature on gas turbine thermal efficiency Figures 11.1 and 11.2 have been produced for three different values for relative humidity (0%, 60% and 100%). It should be noted that the specific humidity which is the amount (mass) of water vapour present in the dry air, rather that the relative humidity which is the amount of water required to saturate the air, affects gas turbine performance. The effect of humidity on power and thermal efficiency is small. However, the increase in specific humidity increases power output but decreases thermal efficiency during constant exhaust gas temperature operation. At constant gas generator speed operation both power output and thermal efficiency decreases. This decrease in performance at constant gas generator speed is very small because such operation normally occurs at low ambient temperatures where the effects of specific humidity are also small. This is observed in Figure 4.6 above, which shows the effect of humidity is more pronounced at high ambient temperatures. For a given value of relative humidity the specific humidity increases with ambient temperature as illustrated in Figure 4.6. Thus the impact of humidity on gas turbine performance decreases with ambient temperature. The variation of gas generator and power turbine creep life usage during the ambient temperature transient is shown in Figure The gas generator turbine creep usage decreases slightly during the period of constant EGT operation. This is primarily due to a slight decrease in cooling air temperature. At ambient temperatures when the gas generator or the compressor non-dimensional speed limits the engine performance, the turbine entry temperature (T3) decreases. Thus the gas generator turbine metal temperature decreases. This results in a significant decrease in gas generator turbine creep life usage when operating at these ambient temperatures. The power turbine creep life usage increases during constant EGT operation and decreases during constant gas generator speed and compressor non-dimensional speed operation. The 88

90 power turbine is assumed to be un-cooled and during constant EGT operation the increased power output as the ambient temperature decreases increases the torque on the power turbine blades and thus stress. The increase in stress in the power turbine blades results in an increased power turbine creep life usage. The effect of humidity turbine creep life usage is very small. Figure 11.3 shows the turbines lives changing with ambient temperature As stated above, gas turbine emissions are affected primarily by combustion pressure, temperature, fuel-air ratio and humidity. An increase in combustion pressure, temperature and fuel-air ratio results in increased Oxides of Nitrogen (NOx) emissions but reduce the emissions of Carbon Monoxide (CO). The effect of humidity has a profound impact on NOx emissions where NOx decreases exponentially with increase in specific humidity. This is due to humidity suppressing of the peak combustion temperatures. The impact of humidity on CO emissions is much smaller. Therefore at low relative humidity, NOx emissions increase during the period of constant EGT and decrease during constant gas generator and compressor nondimensional speed operation. At high relative humidity there is a small increase in NOx during constant gas generator speed operation, but decreases at constant compressor non-dimensional speed. This is due to the decrease in specific humidity as the ambient temperature decreases during constant gas generator speed operation. At constant compressor non-dimensional speed operation the combustion temperature decreases sufficiently resulting in a decrease in NOx during this period of engine operation. CO emissions decrease during constant EGT operation but increase during constant gas generator and compressor non-dimensional speed operation. This is primarily due to the decrease in combustion temperature during constant speed operation. Carbon Dioxide (CO 2 ) emissions on a mass basis, follows the fuel flow changes and increase during constant EGT operation while decreasing during constant speed operation. CO 2 emissions on an index basis decrease as the thermal efficiency increases therefore decreasing during this transient. Emission profiles are illustrated in Figure 11.4 and

91 Figure 11.4 shows the impact of ambient temperature on NOx and CO emissions Figure 11.5 shows the impact of ambient temperature on CO2 emissions We may elect to run a simulator at lower generator power demand, say, 17.5 MW while subjecting the engine to the same ambient temperature transient discussed above. Some of the differences are that the gas generator speed and the fuel flow decrease throughout the transient for this low power case. The gas generator and power turbine creep life usage decrease. The emissions of NOx increase at high ambient temperatures while decreasing at low ambient temperatures due to the continuous decrease is specific humidity. At very low relative humidity NOx emissions will continuously decrease for this low power case and this is primarily due to the continuous decrease in combustion temperature. Therefore, CO emissions 90