Chapter 5
Prepared by : Dr. N. Ait Messaoudene Based on: El-Wakil, Power Plant Technology, McGraw-Hill, 1984. nd
2
semester 2012-2013
1 INTRODUCTION
The gas turbine obtains its power by utilizing the energy of burnt gases and a ir, which is at high temperature and pressure by expanding through the several stages of fixed and moving blades (stator and rotor). To get a high pressure (of the order of 4 to 10 bar) of working fluid, which is essential for expansion a compressor, is required. A simple gas turbine cycle consists of (1) a compressor, (2) a combustion chamber and (3) a turbine. Since the compressor is coupled with the turbine t urbine shaft, shaft, it absorbs some of the power produced by the turbine and hence lowers the eefficiency. fficiency. The network is therefore the difference between the turbine work and work required by the compressor to drive it. Gas t urbines are constructed to work mainly on oil and/or natural gas. CLASSIFICATION OF GAS TURBINE POWER PLANT
The gas turbine power plants which are used in electric power industry are classified into two groups as per the cycle of operation: Open cycle c ycle gas turbine; Closed cycle gas turbine. Open cycle gas turbine
Advantages 1. No warm-up time. Once the turbine is brought up to the t he rated speed by the starting motor and the t he fuel is ignited, the gas turbine will be accelerated from cold st art to full load without warm-up time. t ime. The stipulation of a quick start and take-up of load frequently are the points in favor of open cycle plant when the plant is used as peak load plant. 2. Low weight and size. The weight in kg per kW developed is less. 3. Fuels. Almost any hydrocarbon fuel from high-octane gasoline to heavy diesel o ils can be used in the combustion chamber. 4. Open cycle plants o ccupy comparatively little space. 5. Open-cycle gas turbine power plant, except those having an intercooler, does not require cooling water. Therefore, the plant is independent of o f cooling medium and becomes self-contained.
1 INTRODUCTION
The gas turbine obtains its power by utilizing the energy of burnt gases and a ir, which is at high temperature and pressure by expanding through the several stages of fixed and moving blades (stator and rotor). To get a high pressure (of the order of 4 to 10 bar) of working fluid, which is essential for expansion a compressor, is required. A simple gas turbine cycle consists of (1) a compressor, (2) a combustion chamber and (3) a turbine. Since the compressor is coupled with the turbine t urbine shaft, shaft, it absorbs some of the power produced by the turbine and hence lowers the eefficiency. fficiency. The network is therefore the difference between the turbine work and work required by the compressor to drive it. Gas t urbines are constructed to work mainly on oil and/or natural gas. CLASSIFICATION OF GAS TURBINE POWER PLANT
The gas turbine power plants which are used in electric power industry are classified into two groups as per the cycle of operation: Open cycle c ycle gas turbine; Closed cycle gas turbine. Open cycle gas turbine
Advantages 1. No warm-up time. Once the turbine is brought up to the t he rated speed by the starting motor and the t he fuel is ignited, the gas turbine will be accelerated from cold st art to full load without warm-up time. t ime. The stipulation of a quick start and take-up of load frequently are the points in favor of open cycle plant when the plant is used as peak load plant. 2. Low weight and size. The weight in kg per kW developed is less. 3. Fuels. Almost any hydrocarbon fuel from high-octane gasoline to heavy diesel o ils can be used in the combustion chamber. 4. Open cycle plants o ccupy comparatively little space. 5. Open-cycle gas turbine power plant, except those having an intercooler, does not require cooling water. Therefore, the plant is independent of o f cooling medium and becomes self-contained.
Disadvantages 1. The part load efficiency of the open cycle plant decreases rapidly as the considerable percentage of power developed by the turbine is used to drive the compressor. 2. The system is sensitive to the component efficiency; particularly that o f compressor. The open o pen cycle plant is sensitive to changes in the atmospheric air temperature, pressure and humidity. 3. The open-cycle gas t urbine plant has high air rate compared to the t he other cycles, therefore, it results in increased loss of heat in the exhaust gases a nd large diameter d iameter ductwork is necessary. 4. It is essential that the dust should be prevented from entering into the compressor in order to minimize erosion and depositions on the blades and passages o f the compressor and turbine and so impairing their profile and efficiency. The deposition of the carbon a nd ash on the turbine turb ine blades is not at all desirable as it also reduces the efficiency of the turbine. Therefore, T herefore, air filters and fuel treatment are necessary; combustion must also be handled with a lot of care. Closed cycle gas turbine
Advantages 1. The inherent disadvantage of open cycle gas turbine is the atmospheric backpressure at the turbine exhaust. With closed cycle gas turbine plants, the backpressure can be increased, thus increasing power rating. The density of the working medium can be maintained high by increasing internal pressure range. The high density of the working fluid also increases the heat transfer properties in the heat exchanger. Therefore the machine can be smaller and cheaper than the machine used to develop the same power using open cycle plant. 2. The closed cycle avoids erosion of the turbine blades due to the contaminated gases and fouling of compressor blades due to dust. Therefore, it is practically free from deterioration of efficiency in service. 3. The need for filtration filtration of the incoming air which is a severe problem in open c ycle plant is co mpletely eliminated.
4. Load variation is usually obtained by vary ing the absolute pressure and mass flow of the circulating medium, while the pressure ratio, the temperatures and the air velocities remain almost constant. This result in velocity ratio in the compressor and turbine independent o f the load and full load thermal efficiency maintained over the full range of operating loads. 5. As indirect heating is used in closed cycle plant, the inferior oil or solid fuel can be used in the furnace and these fuels can be used more economically because these are available in abundance. Even more encouraging prospects can be predicted with the possibility of using renewable energy sources (solar towers for example). 6. Finally the closed cycle opens the new field for the use of working medium (other than air as argon, CO2, helium) having more desirable properties. As we are going to see, the ratio γ of the working fluid plays an important role in determining the performance of the gas turbine plant. 7. The maintenance cost is low and reliability is high due to longer useful life. 8. The thermal efficiency increases as the pressure ratio ( R p) decreases. Therefore, appreciable higher thermal efficiencies are obtainable with closed cycle for the same maximum and minimum temperature limits as with the open cycle plant. 9. Starting of plane is simplified by reducing the pressure to atmospheric or even below atmosphere so that the power required for starting purposes is reduced considerably. Disadvantages 1. The system is dependent on external means as considerable quantity of cooling water is required in the precooler. 2. Higher internal pressures involve complicated design of all co mponents and high quality material is required which increases the cost of the plant. 3. The response to the load variations is poor compared to the open-cycle plant, 4. It requires very big heat-exchangers as t he heating of workings fluid is done indirectly. The space required for the heat exchanger is considerably large. The full heat of the fuel is also not used in this plant. The closed cycle is only preferable over o pen cycle where the inferior type of fuel or solid fuel is to be used and ample cooling water is available at the proposed site of the plant. However, closed cycle gas turbine plants have not been used for electricity production, excep t in very limited cases. APPLICATIONS OF GAS TURBINE IN POWER GENERATION 1. Gas turbine plants are used as standby p lants for the hydro-electric power plants. 2. Gas turbine power plants may be used as peak loads plant and standby plants for smaller power units. ADVANTAGES OF GAS TURBINE POWER PLANT 1. It is smaller in size and weight as compared to an equivalent steam power plant 2. The initial cost and operating cost of the plant is lower than an equ ivalent steam power plant. 3. The plant requires less water as compared to a condensing steam power plant. 4. The plant can be started quickly, and can be put on load in a very short time. 5. There are no standby losses in the gas turbine power plant whereas in steam power plant these losses occur because boiler is kept in operation even when the turbine is not supplying any load. 6. The maintenance of the p lant is easier and maintenance cost is low. 7. The lubrication of the plant is easy. In this plant lubrication is needed mainly in compressor, turbine main bearing and bearings of auxiliary equipment. 8. The plant does not require heavy foundations and building. 9. There is great simplification of the plant over a steam plant due to the absence of boilers with their feed water evaporator and condensing system.
DISADVANTAGES 1. Major part of the work developed in the turbine is used to derive the compressor. Therefore, network outp ut of the plant is low. 2. Since the temperature of the pro ducts of combustion becomes too high so service conditions become complicated even at moderate pressures. 3. Proper air filtering and combustion control to prevent corrosion problems.
2 THE IDEAL BRAYTON CYCLE The ideal cycle for gas turbine work is the Brayton cycle. It is composed of two adiabatic-reversible (and hence isentropic) and two constant-pressure processes.
P·V and T·s diagrams of an ideal Brayton cycle. The gas is compressed isentropically from point Ito 2, heated at constant pressure from 2 to 3, and then expanded isentropically through the turbine from point 3 to 4. Cooling occurs from point 4 to point 1, either in a heat exchanger (closed cycle) or in the open atmosphere (open cycle). The work done in the t urbine (a steady-flow machine) per unit time (power) is given by:
It should be noted that specific heats for monatomic gases such as helium and argon are essentially constant and independent of temperature. Specific heats increase with temperature for diatomic gases such as air and N2 and increase even faster with temperature for triatomic gases suchas CO2 (Fig. 8-6).
Variation of c p with temperature for various gases. The following analysis, therefore, is exact for monatomic gases and only approximate for others (the airstandard assumption: c p cte and supposed to the value for air; called cold air-standard assumption if properties are taken at standard co nditions):
Processes 1-2 and 3-4 are isentropic, and P 2 = P 3 and P 4 = P 1. Thus,
Also recalling:
c p-cv = R and ; k = c p /cv
And the definition of the pressure ratio across the turbine r pT = P 3 /P 4 ; which can related to the temperature ratio by:
The work of the turbine can t herefore be written as:
For the compressor; the pressure ratio across is r pC = P 2 /P 1 ; which can be which can related to the temperature ratio by:
The absolute value of the co mpressor work (a consumed work) can also be wr itten as:
Assuming no pressure losses in the cycle ( P 2 = P 3 and P 4 = P 1), the pressure ratio is the same for the turbine and compressor r pT = r pC = r p. The net cycle work rate is then given by:
On the other hand, the by heat added to the cycle is given by: The efficiency of the cycle is then:
Equation 8 – 10 shows that under the co ld-air-standard assumptions, the thermal efficiency of an ideal Brayton cycle depends on the pressure ratio of the gas turbine and the spe cific heat ratio of the working fluid.
Although the above equations pertain to constant specific-heat gases, the trends t hey predict apply to all gases. Rewriting Eq. (8-8) in terms of T 1 and T 3 , using Eq. (8-6). Again for r pT = r pC = r p we can write:
Examination of Eqs (8-8) and (8-11) shows the following: For the same T 1 , T 3 , r p , and k, the work per unit mass of gas is a d irect function of c p. Hence, helium can produce more than five times wnet (work per unit mass) than air (at low temperatures).
For the same T 1 , T 3 , r p , and c p , the work per unit mass of gas increases with k. Again, this shows an advantage for He over air (k for air decreases with temperature). For any one gas, an increase in r p from its lowest value of 1.0 (where the work is zero) decreases one part of Eq. (8-11) and increases the other. The net work thus goes through a maximum at an optimum value of r p. This can be shown graphically by the three ideal cycles of Fig. 8-7. These operate between the same temperatures T 1 , and T 3 (limited by metallurgical considerations), and have the same inlet exhaust pressure but different values of r p. The net work in each case is represented by the enclosed area of the cycle.
Figure 8-7 Effect of pressure ratio on ideal Brayton cycle ( T 1 and T 3 are fixed for the 3 cycles).
The optimum pressure ratio can be evaluated for ideal cycles by differentiating the net work in Eq. (8-11) with respect to r p and equating the derivative to zero. Temperature T 2 is the obtained:
Note that the quantity k/2(k - I) decreases as k increases. Thus, for fixed initial and maximum cycle temperatures, the optimum pressure ratio for monatomic gases (He) is, in general, lower than for diatomic gases (air, N2). These in turn have lower ratios than the triatomic gases (CO2). This means an operation with lower maximum pressure under optimal conditions.
3 THE NONIDEAL BRAYTON CYCLE
The actual gas-turbine cycle differs from the ideal Brayton c ycle on several accounts. For one thing, so me pressure drop during the heat-addition and heat rejection processes is inevitable. But this effect can be neglected as a first analysis. More importantly, the actual work input to the compressor is more, and the actual work output from the turbine is less because of irreversibilities.
p· V and T-s diagrams of ideal and nonideal Brayton cycle.
The deviation of actual compressor and turbine behavior from the idealized isentropic behavior can be accurately accounted for by utilizing the isentropic efficiencies of the turbine and compressor as
If c p is assumed constant:
For the turbine:
If c p is assumed constant:
And the net power for the cycle (work rate) is given as:
This equation can be written in terms of the initial temperature T 1, a chosen metallurgical limit T 3 , and the compressor and turbine efficiencies (above) to give:
The second quantity in parentheses can be recognized as the efficiency of the corresponding ideal cycle, i.e., one having the same pressure ratio and using the same fluid. As in the case of the ideal cycle, the specific power of the nonideal cycle attains a maximum value at some optimum pressure ratio and is a d irect function of the specific heat of the gas used. The heat added in the cycle, Q., is given by
The efficiency of the nonideal cycle can then be obtained by dividing Eq. (8-19) by Eq. (8-20). Although the efficiency of the ideal cycle is independent o f cycle temperatures, except as they may affect k, and increases asymptotically with c p , the efficiency of the nonideal cycle is very much a function of the cycle temperatures. It also assumes a maximum value at an o ptimum pressure ratio for each set of temperatures T 3 and T1. The two optimum pressure ratios, for specific power and for efficiency, are not the same, and this necessitates a compromise in design. Figures 8-10 and 8-11 show results of calculations for η and the work per unit mass of a simple air-combustion Brayton cycle (solid lines) and of one with a regenerator (dashed lines; explained below). For the simple cycle, the following data were assumed and act ual variable properties were used for air and co mbustion gases.
It can be seen that both the efficiency and the work depend strongly on T3. This means that the cycle must be operated as closely as possible to the maximum tolerable te mperature. The effect of r Pc is also very strong, with optimum r Pc increasing with T 3 for both efficiency and work. It can also be seen that the optimum value of r Pc is greater for the efficiency than for work.
Figure 8.10 Efficiency versus compressor pressure ratio of a nonideal Brayton cycle, showing effects of maximum temperature and regeneration.
Figure 8.11 Specific power versus compressor pressure ratio of a nonideal Brayton cycle, showing effects of maximum temperature and regeneration.
Figure 8-12: 33.75 MW direct cycle gas turbine powerplant.
Overview of a gas turbine-set (Nordström, 2005)
4 MODIFICATIONS OF THE BRAYTON CYCLE Modifications that are made to the basic cycle to improve the output and efficiency (and hence t he heat rate):
Regeneration Compressor intercooling Turbine reheat Water injection
Regeneration
Regeneration, therefore, is used to preheat the compressed gas at 2 by the exhaust gases at 4 in a surface-type heat exchanger called the regenerato r or, sometimes, the recuperator. Figure 8-13 shows such a n arrangement for a closed cycle, suitable for He, but also used equally effectively for open cycles with air.
Figure 8-13 Flow and T-s diagrams of a closed nonideal Brayton cycle with regeneration. If the regenerator were 100 percent effective, the temperature of the gas entering t he combustion chamber or nuclear reactor would be raised from T 2 to T 2”. The heat added would be reduced from H 3 – H 2 to H 3 – H 2”, with corresponding increase in cycle efficiency. Actually, the regenerator effectiveness is never 100 percent, and the compressed gases are heated instead to a lower temperature T 2’ . Regenerator-effectiveness , R, is defined as the ratio of the actual to maximum possible temperature change. In other words :
and since temperature at points 4” and 2 are the same (max heating potential of exhaust gases):
For the real cycle computations shown in Figures 8-10 and 8-11, the effects of adding a regenerator with R = 0.75, are included and shown by the dashed lines. It can be seen that the effect of adding a regenerator on efficiency is remarkable and shifts the optimum pressure ratio for efficiency to lower values. The efficiency curves for a cycle with regenerator cross those for the simple cycle at points such as a, beyond w hich the effect of a regenerator on efficiency is negative. These points represent pressure ratios at which t he exhaust gases are cooler than those after compression. The effect of the regenerator on the specific power curves (when taking into account pressure losses) is only to reduce them somewhat because of the added pressure losses in the regenerator. Because regenerative gas-turbine cycles are more efficient than simple gas-turbine cycles, thus reducing fuel consumption by 30 percent or more, they are now used by utilities for meeting cycling duty as well as base- load assistance in driving pumps, compressors, and ot her auxiliary equipment. Compressor Intercooling
Since compressor work is negative (consumed), it is advantageous to keep a low temperature while reaching the desired pressure P2. This can theoretically be done by continuous cooling of the compressed gas to keep it atT1 as shown by the lower horizontal dashed line of Fig.8-14. However, this is not physically possible, and cooling, instead, is done in stages. Intercoolers can be a ir-cooled heat exchangers but are more commonly watercooled. Figure 8-13, drawn for simplicity for ideal (isentropic) compression and expansion, shows two stages of intercooling. Ideally T 1 = T 1’ = T 1” and T 2 = T 2’ = T 2”. In that case we have three compressor sections operating in tandem with equal work because for anyone compressor section (replacing c p by nR/(n-1))
where n is the polytropic exponent for compression (equal to k for ideal compression). When the temperature rises are equal, the pressure ratios are equal because
and the pressure ratio per stage is given by
Where Nc is the number of compressor sections. Thus for an o verall compressor pressure ratio of 10 and 3 sections, the pressure ratio per stage is 3 10 = 2.154 (not 10/3=3.33). The improvement in the cycle is in increased net work (due to a decrease in compression work) and efficiency: The heat added is also increased by H x – H 2”; but this is offset by the net work increase and efficiency is improved.
Figure 8-14 Flow and T-s diagrams of a closed ideal Brayton cycle with two stages of intercooling, one stage of reheat and regeneration. Turbine Reheat
Turbine work can be increased by keeping the gas temperatures in the turbine high. This can also be done theoretically by continuous heating of the gas as it expands through the turbine, as shown by the upper horizontal dashed line of Fig. 8-14. Note that if cooling and heating were at constant temperatures, and if the rest of the cycle were ideal, we would have an ideal Ericsson cycle, which has the same efficiency as a Carnot cycle operating between the same temperature limits T 1 and T 3. Again, continuous heating is not practical and reheat is done in steps or stages. Figure 8-14 shows two turbine sections and one stage of reheat. For T 3 = T 3’ and T 4 = T 4’ , the pressure ratio per turbine stage is
The effect of reheat is an increase in turbine work output with an increase in heat input. But, the net effect is an increase in both work and e fficiency. Intercooling, reheat, and regeneration can all be combined in one cycle as shown in Fig. 8-14. General equations for the specific power and heat added for a composite cycle as the one discussed above, for the case of constant specific heat, but with nonidealities taken into account, are:
The efficiency of the cycle may now be obtained by dividing Eq. (8-25) by Eq.(8-26).The greater the number of reheat and intercooling stages there are, the higher the efficiency. However, this is attained at the co st of the capital investment and size of the p lant. The design of the plant should be optimized, with consideration given to capital versus-operating (fuel, etc.) expenses and to size.
Water Injection Water injection is. a method by which the power output of a gas-turbine cycle is, materially increased and the efficiency is only marginally increased. In gas-turbine cycles that have regenerators, water injection is more beneficial if it is injected between the compressor and regenerator. The method can be used on both single- and two-shaft units. Figure 8-15 shows a schematic of a t wo shaft unit with water injection between compressor and regenerator.
Figure 8-15: Flow and T·s diagrams of a two-shaft gas-turbine cycle with water injection and egeneration
The quantity of water vapor to be injected is that which would saturate the compressed air at T 3. A greater amount of water results in liquid carry through which, although it results in somewhat increased work, also results in reduced efficiency compared with that o f saturated air and in fouling of the regenerato r, local severe temperature differences, and associated t hermal stresses. The increase in work of a t urbine plant with water injection is, in part, a result o f increased turbine work due to the increased mass-flow rate of air and water vapor without a corresponding increase in co mpressor work. The increased mass stems from the saturated vapor at point 3 (F ig. 8-15) minus the water vapor originally in the air at point 1. Using Eq. (7-4), this is given by
The temperature at point 3 can be obtained by an energy balance on the dry air and water vapor
The exhaust emissions are also favorably affected by water injection. Emissions of CO and unburned hydrocarbons in gas-turbine powerplants are not significant because o f the high air-to-fuel ratios used in them. They become of concern only at very high loads when the air-to-fuel ratios are reduced. The oxides of nitrogen (NOx), however, are becoming a problem in gas-turbine combustion because of the steadily increasing combustion temperatures in modem units. It has been found that water injection reduces NOx by at least half
5. DESIGN FOR HIGH TEMPERATURE From the previous sections, it is clear that gas-turbine po werplants need to be operated at high turbine inlet temperature to achieve higher efficiencies and po wer output. This also means higher pressure ratios because optimum pressures increase with increasing turbine inlet temperatures for both efficiency and power. High pressure ratio units have higher capital costs than lower-pressure ones, but the decrease in fuel consumption rapidly pays back for this capital cost differential. Another co ncern that goes with higher temperatures is increased potential for corrosion, which has to be dealt with. As indicated earlier, research and development is underway to raise turbine inlet temperatures from the present 20 00 to 2300°F (1090 to l260°C) to near 2800°F (1540°C). Such temperatures are well above those t hat modem steam turbines have to cope with, which are around 1000 to l200°F (540 to 650°C).
The present range is suitable for peaking service, and with regeneration, for cyclic and some base -load service. It is also competitive with steam plants when used in a co mbined cycle. Future ranges would make them competitive on their own. There are several approaches to the problems associated with high gas temperatures. In general they can be categorized as developing suitable (1) materials, (2) cooling, and (3) fuels. Materials The components that suffer most from a combination of high t emperatures, high stresses, and chemical attack are those of the turbine first-stage fixed blades (nozzles) and moving blades. They must be weldable and castable and must resist corrosion, oxidation, and thermal fatigue. Heat resistant materials and precision casting are two recent advances largely attributable to aircraft engine developments. Cobalt-based alloys have been used for the first-stage fixed blades (which are subjected to t he highest temperatures but not the high stress of the moving blades). These alloys are now being supplemented by vacuum-east nickel-base alloys that are strengthened through solution and precipitation-hardened heat treatment. For the moving blades, cobalt-based alloys with high chromium content are now used.
Ceramic materials are also being developed, especially for the turbine inlet fixed blades. Developmental problems here are inherent brittleness, which causes fabrication problems and raises uncertainties about the mechanical properties of ceramic materials.
Cooling Early turbines operated uncooled, as do many present-day ones. The increases in temperatures we are witnessing require cooling, however. The t hermal stresses in high-temperature turbine moving blades are caused by the high rotational speeds, uneven temperature distributions in the different blade cro ss sections, and static and pulsating gas forces that may give rise to dangero us vibrational stresses. Other thermal stresses occur during start-up, shutdown, and load changes. Thermal stresses are thus caused by steady-state as well as transient operation. The latter give rise to low-cycle fatigue, which reduces blade life. In addition there are problems of creep rupture, high-temperature corrosion, and oxidation. It is generally agreed that blade surfaces should be kept below about l650"F (900°C>, to reduce corrosion to a tolerable degree. A blade is cooled by being made hollow so that a coolant can circulate through it. A hollow blade is lighter than a solid blade and has a much lower Biot number; and hence a fairly uniform temperature distribution.
The coolants that have been used and/or are under consideration are air and water (and steam). The ranges for these are air for gas temperatures up to about 2100 °F (1l50°C), water for gas temperatures above 24 00 °F (l315°C), and a hybrid systems for the intermediate range. In the hybrid system, water cooling is used for the highest temperature components, mainly the inlet fixed b lades, and air for the remaining blades a nd rotor [71].
Fig 8-17 Air-cooled GT fixed blade
Figure 8-20 A water cooled GT moving blade.
Fuels Residual liquid fuels, the residue left after the profitable light fractions have been extracted from the crude, have been used in gas turbines to some extent [74]. They are (I) viscous and (2) tend to polymerize (form sludge or tar) when overheated. (3) Their high carbon content leads to excessive carbon deposits in the combustion chamber. (4) Their contents of alkali metals- such as so dium- combine with sulfur to form Sulfates that are corrosive. (5) They have other metals like vanadium with co mpounds that form during combustion also being corrosive. (6) They have relatively high ash content that deposits mostly on the inlet fixed blades, thus reducing gas flow and power output.
The rate of corrosion increases with increasing gas temperatures. Early turbines de signed for residual fuel use operated at temperatures below 1650'F (900 K) to avoid the problem. Ash deposition is not a problem with intermittent operation because of successive expansions and contract ions, but it is a serious problem with steady operation.
6. ELEMENTS OF GAS TURBINE POWER PLANT COMPRESSORS The type of compressor which is commonly used is the axial flow type. The axial flow compressor consists of a series of rotor and stator stages with decreasing diameters along the flow of air. A satisfactory air filter is absolutely necessary for cleaning the air before it enters the compressor because it is essential to maintain the designed profile of the aerofo il blades. The deposition of dust particles on the blade surfaces reduces the efficiency rapidly.
Axial Flow Air Compressor
INTERCOOLERS AND HEAT EXCHANGERS
The intercooler is generally used in gas turbine plant when the pressure ratio used is sufficiently large and the compression is completed with two or more stages. The cooling of co mpressed air is generally done with the use of cooling water. A cross-flow type intercooler is generally preferred for effective heat transfer. The regenerators, which are commonly used in gas turbine plant, are of two types, recuperator and regenerator. In a recuperative type of heat exchanger, the a ir and hot gases are made to flow in counter direction as the effect of counterflow gives high average temperature difference causing the higher heat flow.
The regenerator type heat exchanger consists of a heat-conducting member that is exposed alternately to the hot exhaust gases and the cooler compressed air. The heat-conducting member is made of a metallic mesh or matrix, which is rotated slowly (40-60 r.p.m.) and continuously exposed to hot and cold air. The major disadvantage of this heat exchanger is, there will be always a tendency for air leakage to the exhaust gases as the compressed air is at a much higher pressure t han exhaust gases.
COMBUSTION CHAMBERS
One of the vital problems associated with the design o f gas turbine combustion system is to secure a steady and stable flame inside the combustion chamber. The gas t urbine combustion system has to function under certain different operating conditions which are not usually met with the combustion systems of IC engines. A few of them are listed below: Combustion in the gas turbine takes place in a continuous flow system. High rate of mass flow results in high velocities at various points throughout the cycle (300 m/sec). On the other hand, the chemical reaction takes place relatively slowly thus requiring large residence t ime in the combustion chamber in order to achieve complete combustion.
The gas turbine requires about 100:1 air-fuel mass ratio (for comparison, the air-fuel ratio required for the combustion in diesel engine is approximately 15:1) and it is impossible to ignite and maintain a continuous combustion with such weak mixture. It is therefore necessary to allow required air in the combustion zone (usually a rich mixture) and the remaining air is added a fter complete combustion to reduce the gas temperature before passing into the turbine. The solution is to create a pilot or recirculated zone in the main flow to establish a stable flame that helps to ignite the combustible mixture continuously. A stable continuous flame can be maintained inside the combustion chamber when the stream velocity and fuel burning velocity are equa l. Unfortunately most of the fuels have low burning velocities of the order of a few meters per second; therefore, flame stabilization is not possible unless some technique is employed to anchor the flame in the combustion chamber. The common methods of flame stabilization used in practice are bluff body method and swirl flow method.
Combustion Chamber with Upstream Injection with Bluff-body Flame Holder.
Combustion Chamber with Downstream Injection and Swirl Holder.
GAS TURBINES The common types of turbines, which are in use, are axial flow type. The basic requirements of the turbines are lightweight, high efficiency; reliability in operation and long working life. Large work output can be o btained per stage with high blade speeds when the blades are designed to sustain higher stresses. More stages of the turbine are always preferred in gas turbine power plant because it helps to reduce the stresses in the blades and increases the overall life of the turbine. More stages are further preferred with stationary power plants because weight is not the major consideration in the design which is essential in aircraft turbine-plant.
Compressor detailed internal blade view (Rolls-Royce, 1992)
SGT-750 35 MW Siemens gas turbine
The cooling of the gas turbine blades is essential for long life as it is continuously subjected to high temperature gases. There are different methods of cooling the b lades. The common method used is the air-cooling. The air is passed through the holes provided through the blade.
AUXILIARY SYSTEMS
Auxiliary systems are the backbone of the gas t urbine plant. Without auxiliary system, the very existence of the gas turbine is impossible. It permits the safe working of the gas turbine. The auxiliary system includes starting, ignition, lubrication, air filtering and fuel system and control. STARTING SYSTEMS Two separate systems-starting and ignition are required to ensure a g as turbine engine will start satisfactorily. During engine starting the two systems must operate simultaneously.
Air starting (pneumatic) is used mostly as it is light, simple and economical to o perate. The starter turbine is rotated by air pressure taken from an external supply, from an aux iliary power unit or from an engine that is running. The starter turbine rotor transmits power through a reduct ion gear and clutch to the starter output shaft that is connected to the powerplant turbine. The clutch automatically disengages as the engine accelerates to a predetermined starter speed. Other starting systems are: electrical; combustion (a small auxiliary GT); hydraulic. IGNITION SYSTEMS Ignition system is utilized to initiate spark during the starting. Once it starts, the combustion is continuous and the working of ignition system is cut-off automatically. A flame detector is used to ensure combustion and automatically activate the ignition system if necessary.
LUBRICATION SYSTEM The following are the elements of lubrication system of a gas t urbine 1. Oil tank, 2. Oil pump, 3. Filter and strainer, 4. Relief valve, 5. Oil cooler, 6. Oil and pipe line, 7. Magnetic drain plug, 8. By-pass, valve, and 9. Warning devices.
FUEL SYSTEM Basically there are two sub-systems: low pressure and high pressure. The low-pressure system consists of a fuel tank, boost pump, strainer, dual fine oil filters, Δ p indicator and ordinary valves. The high-pressure system consists of main fuel pump, relief valve, strainer, fuel control, shut-o ff valve, flow distributor (to ensure that the fuel is well distributed in the various fuel injectors) and injectors (nozzles).
Fuel System for Gas Turbine.
Splash Plate Type Injector and Duplex Nozzle Injector Air Filtering
Filtration’s role in power generation is critical. Air intake filters protect the most valuable equipment from degradation caused by exposure to outdoor air pollutants. Proper inlet filtration is designed based on the required app lication in order to prevent a decrease in gas t urbine performance and even destruction of the gas turbine. In order to understand the importance of inlet filtration, the effects of poor filtration on the gas turbine should be u nderstood. The gas turbine is affected by various substances in the inlet air depending upon their composition and their particle size. There are six common consequences of poor inlet air filtration: foreign object damage, erosion, fouling, turbine blade cooling passage plugging, particle fusion, and corrosion (hot and cold).
CONTROL OF GAS TURBINES The purpose of gas turbine controls is to meet the specific control requirements of users and safe operation of the turbine. There are basically two types o f controls: Prime control and Protection control.
The objective of the prime control is to ensure the proper application of the turbine po wer to the load. In power generation applications, the main requirement is to co ntrol the frequency of the a.c. generator at any load. This is achieved by selecting the pr imary controller as a speed governor (a speed sensing device), which maintains the constant electrical loads. The objective of the protective control is to ensure adequate protection for the turbine in preventing its operation under adverse conditions. Whenever, unsafe operating conditions are approached, the prime co ntrol is overtaken by the protective control to protect the turbine or driven equipment. Basically, the protective control is of two types: Shutdown control and Modu lating control. The shut down type control detects a condition which can cause a serious malfunction and actuate the shut-off valve to stop the turbine. Following are the various types of shut down controls. (a) Turbine over temperature, (b) Turbine over speed, (c) Low lube oil pressure, (d ) High lube oil temperature, and (e) Excess vibration. The purpose of modulating control is to sense an impending malfunction or a condition, which could adversely affect turbine life and make some modification to the op erating condition of the t urbine in order to alleviate the undesired conditions. An example of this control may be maximum turbine inlet temperature and maximum speed. The modulating control is more co mplex and more costly than the shutdown control. But it offers and advantage in allowing a turbine or turbine driven plant to continue o perating when normally a shut down occurs. There are some conditions such as high vibrations and low lube oil pressure which can not be taken care by corrective measures as in this case shut down of the turbine is essential.
Additional solved example Example 1: In a constant pressure open cycle gas turbine air enters at 1 bar and 20°C and leaves the compressor at 5 bar. Using the following data: Temperature of gases entering the turbine = 68 0°C Pressure loss in the combustion chamber = 0.1 bar ηcompressor = 85%, ηturbine = 80%, ηcombustion = 85%, γ = 1.4 and c p = 1.024 kJ/kgK for air and gas.
Find: (1) The quantity of air circulation if the plant develops 1065 kW. (2) Heat supplied per hg of air circulation. (3) The thermal efficiency of the cycle. Mass of the fuel may be neglected. Solution T-s Diagram of the cycle:
Example 2: A gas turbine plant of 800 kW capacities takes the air at 1.01 bar and 15°C. The pressure ratio of the cycle is 6 and maximum temperature is limited to 700°C. A regenerator of 75% effectiveness is added in the plant to increase the overall efficiency of the plant. T he pressure drop in the combustion chamber is 0.15 bars as well as in the regenerator is also 0.15 bars. Assuming the isentropic efficiency of the compressor 80% and o f the turbine 85%, determine the plant thermal efficiency. Neglect the mass of the fuel. Solution: The arrangement of the components and the processes are represented on T-s diagram as sho wn in the figure:
The given data is
Example 3: In an open cycle regenerative gas turbine plant, the air enters the compressor at 1 bar abs 32°C and leaves at 6.9 bar abs. The temperature at the end of combustion chamber is 816°C. The isentropic efficiencies of compressor and turbine are respectively 0.84 and 0.85. Combustion efficiency is 90% and the regenerator effectiveness is 60 percent, determine: (a) Thermal efficiency, (b) Air rate, (c) Work ratio. Solution: T-s Diagram of the cycle:
Given data: P1 = 1.0 bar, T1 = 273 + 32 = 305 K P2 = P2a = 6.9 bar T4 = 816 + 273 = 1089 K
Or for the turbine:
