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1. Introduction
1.1 Motivation
The use of gas turbine in the petrochemical, power generation and offshore industry has mushroomed in past few years. The growth of the gas turbine in recent years has been brought about most significantly by metallurgical advances to employ high temperatures in the combustor and turbine components, and the cumulative background of aerodynamic and thermodynamic knowledge, and the utilization of computer technology in the design and simulation of turbine blade and combustor. Therefore this is very interesting field and motivated one. A lot of research is going on to improve the efficiency of the gas turbine and its operating temperature.
1.2 Objective
The project objectives were as 1 Design calculation of axial flow compressor 2 components drawing 3 Manufacturing of axial flow compressor
After going through initial design calculation of axial flow compressor, the manufacturing drawing of lab scale axial flow compressor is to be drawn along with assembly drawing of axial flow compressor.. In this thesis the design methodology of axial flow compressor was discussed first and then attention was focused on manufacturing of lab scale axial flow compressor.
1.3 Scope of project
The significance of turbojet engine due to their use in military aircrafts makes this project tremendously important, since compressor is one of the fundamental components of engine and the performance of engine is directly dependent on the performance of compressor.
1.4 Compressor
Compressors are machines that compress air or gas. Compression is achieved through the reduction of the volume that the gas (or air) occupies. As a side effect of the minimization of volume, the temperature of air or gas increase. Compressors are used in refrigeration and air conditioner equipment to move heat from one place to another in refrigerant cycles [1]. Compressor could be of axial flow, centrifugal, or a combination of both types in order to produce highly efficient compressed air which can be used for useful power output. Kinetic energy is imparted to compressor in impeller part of compressor and then this kinetic energy is converted to pressure energy by passing air through diverging section called diffuser.
1.5 Design Calculation
Figure 1-1 Types of compressor Types of compressor are shown in fig 1-1.positive displacement compressors confine successive volume of gas in an enclosed space where pressure increases as the volume of enclosed space decreases. They can be thought as constant volume-variable pressure machine. That is, they move a certain volume of gas with each stroke, and the pressure is that of the system into which they discharge. With dynamic compressors, the mechanical section of rotating impellers imparts pressure and velocities to the gas. They are constant pressure-variable volume machines [1].
in 1678 BC. Then in 1791 BC, John Barber took out a patent for „A method for rising inflammable air for the purposes of producing motion and facilitating metallurgical operations [3].
Today, gas turbines are used widely in various industries to produce mechanical power and are employed to drive various loads such as jet propulsion, generators, pumps and process compressors, varying from civil and military aviation to power generation and process industries. The gas turbine began as a relatively simple engine and evolved into a complex but reliable and high efficiency prime mover. In the quest to perfect the gas turbine, compressor pressure ratios have increased from about 4:1 to over 40:1 together with high operating temperature of compressor (about 1800 K), resulting in thermal efficiencies exceeding 40%. These features make the gas turbine a formidable competitor to other types of prime movers. Clearly, the power output from a gas turbine depends on the efficiency of the compressor, turbine and the combustor. The higher the efficiency of these components, the better will be the performance of the gas turbine, resulting in increased power output and thermal efficiency [4].
Due to these fast growing improvements in the field and wide applications of gas turbines, there is need to work on the design and manufacturing of efficient gas turbines indigenously. Beside the principles and main components, the main emphasize should be given to the thermodynamic, aerodynamic and mechanical design study of a gas turbine.
The next several decades natural gas and syngas fired combined cycle plants will make up as much as 20% of the new electric generating plant additions. Power plants utilizing techniques such as the atmospheric fluidized bed combustion systems, advanced pressurized fluidized bed combustion systems, integrated gasification- combined-cycle systems and integrated gasification-fuel cell combined-cycle systems will make up the new generation of gas turbine applications [4].
With the rapid improvements in gas turbine technology brought about by the war effort fifty years ago, these power and efficiency improvement techniques were set aside. Advances achieved in aircraft engine technology such as; airfoil loading, single crystal airfoils and thermal barrier coatings are being transferred to the
industrial gas turbine. Improvement in power output and efficiency rests primarily on increases in turbine inlet temperature. The union of ceramics and super alloys will provide the material strength and temperature resistance necessary to facilitate increased turbine firing temperature. However, increasing firing temperature will also increase emissions. To insure continued acceptance of the gas turbine, emissions must be eliminated or significantly reduced.
To reduce emissions, catalytic combustors are being developed. The catalytic combustors will reduce NOx formation within the combustion chamber. They will also reduce combustion temperatures and extend combustor and turbine parts life. While increased power (in excess of 200+ megawatts) will be provided without increasing the size of the gas turbine unit, the balance of plant equipment (such as; pre- and inter-coolers, regenerators, combined cycles, gasifiers, etc.) will increase the overhaul size of the facility. Designers and engineers must address the total plant not just the gas turbine. The size of the various process components must be optimized to match each component‟s cycle with the gas turbine cycle, as a function of ambient conditions and load requirements. Computer (microcomputer or programmable logic controller) control systems must be designed to interface and control these various processes during steady-state and transient operation [4].
Researchers will intensify their efforts to supply hydrogen, processed from non-fossil resources, for use in petroleum based equipment. The objective is to produce recoverable, cost effective, and environmentally benign energy. Ultimately the gas turbine will be required to burn hydrogen, even with a plentiful supply of fossil fuel. The use of hydrogen eliminates the fuel bound nitrogen that is found in fossil fuels. With combustion systems currently capable of reducing the dry NOx level to 25 ppmv; and catalytic combustion systems demonstrating their ability to reduce emissions to “single digits”; the use of hydrogen fuel promises to reduce the emission levels to less than 1 ppmv. The gas turbine will need to achieve this low level in order to compete with the fuel cell. Today the largest fuel cell manufactured is rated at only 200 kilowatts. This technology is available, it will be developed, and it will compete with the gas turbine [4].
Hospitals, office complexes, and shopping malls will become the prime sales targets for 1 to 5 megawatt power plants. Operating these small plants, on site, may
2. Axial Flow Compressor
Axial flow compressors are dynamic type of compressors in which air flow is parallel to the axis of rotation. Cut way sketch of axial flow compressor is shown is figure 2-1. The axial flow compressor can achieve higher pressures at a higher level of efficiency. There are two important characteristics of the axial flow compressor high pressure ratios at good efficiency and thrust per unit frontal area [5]. Compressor usually works in two steps 1 Increase the momentum of air by doing work on it(using rotor blades) 2 Decelerate the air to increase static pressure(using stator blades)
Figure 2-1 Axial flow compressor
2.1 Basic Operation
Axial flow compressors are rotating, airfoil based compressors in which the working fluid principally flows parallel to the axis of rotation. This is in contrast with other rotating compressors such as centrifugal, axi-centrifugal and mixed-flow compressors where the air may enter axially but will have a significant radial component on exit.
Axial flow compressors produce a continuous flow of compressed gas, and have the benefits of high efficiencies and large mass flow capacity, particularly in relation to their cross-section. They do, however, require several rows of airfoils to achieve large
pressure rises making them complex and expensive relative to other designs (e.g. centrifugal compressor).
Axial flow compressors are widely used in gas turbines, such as jet engines, high speed ship engines, and small scale power stations. They are also used in industrial applications such as large volume air separation plants, blast furnace air, fluid catalytic cracking air, and propane dehydrogenation. Axial flow compressors, known as superchargers, have also been used to boost the power of automotive reciprocating engines by compressing the intake air, though these are very rare [5].
Axial flow compressors consist of rotating and stationary components. A shaft drives a central drum, retained by bearings, which has a number of annular airfoil rows attached. These rotate between a similar numbers of stationary airfoil rows attached to a stationary tubular casing. The rows alternate between the rotating airfoils (rotors) and stationary airfoils (stators), with the rotors imparting energy into the fluid, and the stators converting the increased rotational kinetic energy into static pressure through diffusion. A pair of rotating and stationary airfoils is called a stage. The cross- sectional area between rotor drum and casing is reduced in the flow direction to maintain axial velocity as the fluid is compressed.
Figure 2-2 Velocity triangle The basic principle of axial flow compressor operation is that kinetic energy is imparted to the air in the rotating blade row, and then diffused through passages of both rotating and stationary blades. The process is carried out over multiple numbers of stages. As mentioned earlier, diffusion is a deceleration process. It is efficient only when the pressure rise per stage is very small. The balding diagram and the velocity
Compressor cascade is shown in figure 2-4.The stagger angle (^) is the angle between
the chord line and the axial direction and represents the angle at which the blade is set in the cascade. (^) 1 / and (^) / are the camber angles of the entry and exit tangents the
camber line makes with the axial direction. The blade camber angle is 1 /^ /.
The chord c is the length of the perpendicular of the blade profile onto the chord line. It is approximately equal to the linear distance between the leading edge and the trailing edge. The pitch s is the distance in the direction of rotation between corresponding points on adjacent blades. The incidence angle i is the difference between the air inlet angle and the blade inlet angle. That is i= (^) 1 /- /.The air
deflection angle 1 is the difference between the entry and exit air angles.
For any particular test, the blade camber angle its chord c and the pitch (or space) s will be fixed and the blade inlet and outlet angles 1 /and / are determined by
stagger angle which is calculated from appendix A-1. [5].
Figure: 2-4 Compressor cascade and blade notation
3. Designing of Axial Flow Compressor
3.1 Symbols
There are three state points within a compressor that are important when analyzing the flow. They are located at the inlet guide vanes, the rotor entrance, and at the stator exit. Rotor blade is located between 1 and 3 as shown in figure 3-
Figure 3-1 Compressor blading arrays Following is the nomenclature and symbols of different variables used which will be used to explain the basic theory. C 1 Absolute velocity of fluid at inlet to stator C 2 Absolute velocity of fluid at exit from stator C 3 Absolute velocity of fluid at exit from rotor C a Axial velocity of fluid U Blade speed v 2 Relative velocity of fluid at rotor inlet v 3 Relative velocity of fluid at rotor exit α 1 Angle of C 1 with axial direction α 2 Angle of C 2 with axial direction β 2 Angle of v 2 with axial direction (blade inlet angle) β 3 Angle of v 3 with axial direction (blade exit angle) p 01 / p 03 Stagnation pressure ratio per stage ηs Isentropic efficiency of stage Λ Degree of reaction
3.4 Flow Coefficient (Φ)
Flow coefficient is the ratio between the axial velocity of fluid and the blade velocity U.
Ca
U
(3-6)
3.5 Aspect Ratio (h/c)
The value of aspect ratio is very much important in designing of axial flow compressor. Low value of aspect ratio leads to secondary flow and tip clearance effects On the other hand, too high a value of h/ c will increase the likelihood of vibration trouble: vibration characteristics are difficult to predict and they depend on the damping provided by the method of attaching the blades to the turbine disc. A value of h/c between 3 and 4 would certainly be very satisfactory, and it would be unwise to use a value below 1. Number of blades can be calculated by using the following equation
(3-7)
Figure 3-2 rotor disk with blade pitch
n^2 r
s
^ (3-8) Where: r :is the root radius s: is the pitch of blade. It is usual to avoid numbers with common multiples to reduce the probability of introducing resonant forcing frequencies. A common practice is to use an even number for the nozzle blades and an odd number for the rotor blades [5].
4. Design Calculations
To design the single stage compressor the given initial data is shown in table 4-1. Table 4-1: Initial Data Axial velocity 45 m/s Inlet temperature 300K Inlet Pressure 1bar Mass flow rate, m 1 kg/sec Rotational speed, N 6000 rpm Polytropic efficiency. Pressure ratio, (^) 2. Hub-tip ratio 0.
4.1 Design Process
In the following section design calculations are given step by step.
4.1.1 Calculation of Temperature, Pressure and Density at
Inlet So first of all we have to calculate static temperature and pressure at inlet of single stage axial flow compressor. Static temperature is calculated by using equation (4-1)
(4-1)
Substituting the given values of stagnation temperature, axial velocity in equation (4-1) 2 1 300 45 T 2 C (^) p
=299k Now static pressure at inlet can be calculated by using equation (4-2)
After calculating tip radius tip velocity can be calculated by using the equation (4- 5).but first of all tip radius can be calculated by using the equation (4-8)
r t ^ A^1 rr^2 Substituting the values of area at inlet and root radius in equation (4-8) tip radius is given as
rt .0258^ .06^2
=0.1m Now tip velocity can be calculated by using the formula
t t 60 U r
Substituting the values of rotational speed which is fixed N=6000 r.p.m and tip radius in equation (4-9)
U t .1 60 =62.832 m/sec Axial velocity is constant along the annulus. Tip velocity can be obtained by using equation (4-10)
V t^2 Ut Ca^2
Substituting the value of tip velocity and axial velocity which is fixed in equation (4- 9). (^2) 62.8 2 452 Vt Vt 77.26 m / sec
4.1.3 Calculation of Mach Number at Inlet Now at this stage it is approximate to check Mach number relative to rotor tip at inlet to compressor
Speed of sound= a RT 1
Substituting the values of static temperature and R in equation (4-11) gives the value of speed of sound.
Speed of sound= a 1.4 287 299 =346.6 m/sec Mach number is calculated by the ratio of tip velocity and speed of sound.
Mach number = 1^ t M V a
Substituting the values of tip velocity and speed of sound in equation (4-12) gives value of Mach number at inlet
1
M 346.
4.1.4 Calculation of Temperature, Pressure and Density at Exit It is instructive to estimate the annulus dimension at exit from the compressor. The compressor delivery pressure is 2.02.To estimate the compressor delivery temperature it is assumed that polytropic efficiency is 0.90.Thus 1 02 01 02 01
n P n T T P
^ as 1 1 1 p
n n
Substituting the values of polytrophic efficiency in equation (4-14) 1 1 1.4 1 0.9 1.
n n
Substituting the values of stagnation temperature at inlet, stagnation pressure at inlet and outlet in equation (4-13) gives value of stagnation temperature at outlet.
T 02 (^) 300 2 0. =373.85K Now static pressure, temperature and density at exit can be calculated by following relationships (4-15)
r te ^ A^2 rr^2
Substituting the values of area at exit and root radius in equation (4-19)
.012 (^) .06 2 rte (^)
=0.086 m/sec Blade height at exit can be calculated by the difference between tip radius at exit and root radius.
he 0.086 0. =0.026 m =2.6 cm
4.1.5 Flow Coefficient Flow coefficient is ratio between axial velocity and blade velocity. Now flow coefficient can be calculated by the following relationship (4-20) Cx U
Velocity triangle for single stage compressor is shown in figure 4-1. velocity triangle is used to calculate blade angles at inlet and exit
Figure 4-1 Velocity triangle
4.1.6 Blade Angles Now blade angles can be calculated from the velocity triangle shown in the figure (4- 1).blade inlet angle can be calculated by using relationship (4-20)
tan 1 a
U
C
Substituting the values of blade speed and axial velocity in equation (4-21)
1 tan 62. 45
1 tan^1 62. 45
^
1 54.38^0
Velocity of air at inlet can be calculated by using equation (4-22)
1 cos 1
V C^ a
Substituting the values of axial velocity and blade angle at inlet in equation (4-22)
1
45 cos54.
V
V 1 77.
In order to estimate maximum possible deflection we will apply the haller criterion (^2) 0. 1
V
V ^ .On the basis of allowable value of V^1^^ 77. V 2 (^) 77.26
V 2 55.
Now we can calculate balde angle at outlet by using the relationship (4-23)
cos 2 2 a
C
V
Substituting the values in equation (4-23)
2 cos 45 55.
