Chapter 5
Prepared by : Dr. N. Ait Messaoudene
Based on:
El-Wakil, Power Plant Technology, McGraw-Hill, 1984.
2nd semester 2012-2013
1 INTRODUCTION
The gas turbine obtains its power by utilizing the energy of burnt gases and air, 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 shaft, it absorbs some of the power produced by the turbine
and hence lowers the efficiency. The network is therefore the difference between the turbine work and work
required by the compressor to drive it. Gas turbines 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 gas turbine; Closed cycle gas turbine.
Open cycle gas turbine
Advantages
1. No warm-up time. Once the turbine is brought up to the rated speed by the starting motor and the fuel is
ignited, the gas turbine will be accelerated from cold start to full load without warm-up time. 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 oils can be used in the
combustion chamber.
4. Open cycle plants occupy 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 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 of compressor. The open cycle plant is
sensitive to changes in the atmospheric air temperature, pressure and humidity.
3. The open-cycle gas turbine plant has high air rate compared to the other cycles, therefore, it results in
increased loss of heat in the exhaust gases and large diameter 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 of the compressor and turbine and so impairing their profile
and efficiency. The deposition of the carbon and ash on the turbine blades is not at all desirable as it also
reduces the efficiency of the turbine. Therefore, 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 of the incoming air which is a severe problem in open cycle plant is completely
eliminated.
4. Load variation is usually obtained by varying 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 of 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 (Rp) 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 components 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 the 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 open 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, except in very limited cases.
APPLICATIONS OF GAS TURBINE IN POWER GENERATION
1. Gas turbine plants are used as standby plants 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 equivalent 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 plant 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 output
of the plant is low.
2. Since the temperature of the products 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 turbine (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 cp with temperature for various gases.
The following analysis, therefore, is exact for monatomic gases and only approximate for others (the airstandard assumption: cp cte and supposed to the value for air; called cold air-standard assumption if properties
are taken at standard conditions):
Processes 1-2 and 3-4 are isentropic, and P2 = P3 and P4 = P1. Thus,
Also recalling:
cp-cv = R and ; k = cp/cv
And the definition of the pressure ratio across the turbine rpT = P3/P4 ; which can related to the temperature
ratio by:
The work of the turbine can therefore be written as:
For the compressor; the pressure ratio across is rpC = P2/P1 ; which can be which can related to the temperature
ratio by:
The absolute value of the compressor work (a consumed work) can also be written as:
Assuming no pressure losses in the cycle (P2 = P3 and P4 = P1), the pressure ratio is the same for the turbine and
compressor rpT = rpC = rp. 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 cold-air-standard assumptions, the thermal efficiency of an ideal Brayton
cycle depends on the pressure ratio of the gas turbine and the specific heat ratio of the working fluid.
Although the above equations pertain to constant specific-heat gases, the trends they predict apply to all gases.
Rewriting Eq. (8-8) in terms of T1 and T3 , using Eq. (8-6). Again for rpT = rpC = rp we can write:
Examination of Eqs (8-8) and (8-11) shows the following:
For the same T1, T3, rp , and k, the work per unit mass of gas is a direct function of cp. Hence, helium can
produce more than five times wnet (work per unit mass) than air (at low temperatures).
For the same T1, T3, rp , and cp, 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 rp 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 rp. This can
be shown graphically by the three ideal cycles of Fig. 8-7. These operate between the same temperatures T1, and
T3 (limited by metallurgical considerations), and have the same inlet exhaust pressure but different values of rp.
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 (T1 and T3 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 rp and equating the derivative to zero. Temperature T2 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 cycle on several accounts. For one thing, some
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·Vand 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 cp is assumed constant:
For the turbine:
If cp 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 T1, a chosen metallurgical limit T3, 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 direct 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 of cycle temperatures, except as they may affect k, and increases
asymptotically with cp, the efficiency of the nonideal cycle is very much a function of the cycle temperatures. It
also assumes a maximum value at an optimum pressure ratio for each set of temperatures T3 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 actual variable properties were used for air and combustion gases.
It can be seen that both the efficiency and the work depend strongly on T 3. This means that the cycle must be
operated as closely as possible to the maximum tolerable temperature. The effect of rPc is also very strong, with
optimum rPc increasing with T3 for both efficiency and work. It can also be seen that the optimum value of rPc 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 the 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 regenerator or, sometimes, the recuperator. Figure 8-13 shows such an 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 the combustion chamber or
nuclear reactor would be raised from T2 to T2”. The heat added would be reduced from H3 – H2 to H3 – H2”,
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 T2’. 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 which the effect
of a regenerator on efficiency is negative. These points represent pressure ratios at which the 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 other 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 at T1
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 air-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 T1 = T1’ = T1” and T2 = T2’ = T2”. In that case we have three compressor sections operating
in tandem with equal work because for anyone compressor section (replacing cp 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 overall 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 Hx – H2”; 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 T1 and T3.
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 T3 = T3’ and T4 = T4’, 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 efficiency.
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 cost of the
capital investment and size of the plant. 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 two 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 T3. 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 of saturated air and in fouling of the regenerator, local severe
temperature differences, and associated thermal stresses.
The increase in work of a turbine plant with water injection is, in part, a result of increased turbine work due to
the increased mass-flow rate of air and water vapor without a corresponding increase in compressor work. The
increased mass stems from the saturated vapor at point 3 (Fig. 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 of 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 powerplants need to be operated at high turbine inlet
temperature to achieve higher efficiencies and power output. This also means higher pressure ratios because
optimum pressures increase with increasing turbine inlet temperatures for both efficiency and power. Highpressure 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 concern 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 2000 to 2300°F (1090 to l260°C) to near 2800°F
(1540°C). Such temperatures are well above those that 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 combined 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 temperatures, 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 the 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 thermal stresses in high-temperature turbine moving blades are
caused by the high rotational speeds, uneven temperature distributions in the different blade cross sections, and
static and pulsating gas forces that may give rise to dangerous 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 2400 °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 blades, and air for the remaining blades and 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 sodium- combine with sulfur to form Sulfates that are
corrosive. (5) They have other metals like vanadium with compounds 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 designed 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 contractions, 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 aerofoil 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 compressed 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 air 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 than exhaust gases.
COMBUSTION CHAMBERS
One of the vital problems associated with the design of gas turbine combustion system is to secure a steady and
stable flame inside the combustion chamber. The gas turbine 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 time 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 after 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 equal. 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 obtained
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 blades. 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 turbine 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 gas 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 operate. The starter turbine is
rotated by air pressure taken from an external supply, from an auxiliary power unit or from an engine that is
running. The starter turbine rotor transmits power through a reduction 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 turbine
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, Δp indicator and ordinary valves. The high-pressure system
consists of main fuel pump, relief valve, strainer, fuel control, shut-off 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 application in order to prevent a decrease in gas turbine
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 understood. 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 of controls: Prime control and Protection control.
The objective of the prime control is to ensure the proper application of the turbine power to the load. In power
generation applications, the main requirement is to control the frequency of the a.c. generator at any load. This
is achieved by selecting the primary 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 control is
overtaken by the protective control to protect the turbine or driven equipment. Basically, the protective control
is of two types: Shutdown control and Modulating 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 operating condition of the turbine 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 complex and more costly than the shutdown control. But it offers and
advantage in allowing a turbine or turbine driven plant to continue operating when normally a shut down
occurs. There are some conditions such as high vibrations and low lube oil pressure which cannot 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 = 680°C
Pressure loss in the combustion chamber = 0.1 bar
ηcompressor = 85%, ηturbine = 80%, ηcombustion = 85%, γ = 1.4 and cp = 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. The 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 of 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 shown 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:
Example 4:
A gas turbine power plant is operated between 1 bar and 9 bar pressures and minimum and maximum cycle
temperatures are 25°C and 1250°C. Compression is carried out in two stages with perfect intercooling. The
gases coming out from H.P turbine are heated to 1250°C before entering into L.P turbine. The expansions in
both turbines are arranged in such a way that each stage develops same power. Compressors and turbines
isentropic efficiencies are assumed to be 83%.
(1) Determine the cycle efficiency assuming ideal regenerator. Neglect the mass of fuel.
(2) Find the power developed by the cycle in kW if the airflow through the power plant is 16.5 kg/sec.
Solution: The arrangement of the components and the processes are shown in the following figure:
9
And the temperature variation in the regenerator is given in as:
The given data is