FIELD OF INVENTION:
The invention generally relates to multistage axial flow compressor with high flow coefficient handling higher air flow with moderate discharge pressure in the range of 5 to 6.5 bar for process applications. The compressed air is to be discharged radially to accommodate the compressor drive which necessitates use of a radial stage and volute at the end configuration. Air is compressed gradually from inlet to exit of the axial compressor and the volume of air to be handled reduces at the last stages. Inlet guide vane mechanism is used for guiding the flow at inlet for better flow control and efficiency improvement. Introduction of 2D impeller radial stage as last stage for axial compressor will help in achieving radial discharge reducing overall axial length with improved stall margins and increased discharge pressure for the same operating speed.
The present invention more particularly relates to multistage axial flow compressor for process applications such as blast furnace for smelting iron, catalytic cracking of petroleum, air separation plants, power generation in gas turbine and cooling of turbo- generator where higher efficiencies with lower compressor cost and better surge margins are demanded.
BACKGROUND OF THE INVENTION AND PRIOR ART:
Multistage axial flow compressors for process applications such as blast furnace application for smelting of iron, catalytic cracking of petroleum, air separation plants, power generation in gas turbine and cooling of turbo- generator handles air flow rates ranging from 1,00,000 to 5,00,000 NM3 / hr with relative humidity ranging from 20 % to 90 %. These compressors operate at speed of 3800 – 5000 rpm to compress ambient air to moderate discharge pressure of 5 to 6.5 bar.
Multistage axial flow compressor consists of moving and stationary blades, rotor shaft, inlet guide vanes, inner casing, outer casing, sealing arrangements in horizontal parting plane, final radial stage, discharge volute. When the compressor is driven by external source, air enters axially through the bell

mouth and inlet guide vanes and passes successively through moving blades and stationary blades, that are arranged alternatively so that each stage provides an incremental increase in the pressure of the air. The static pressure of air increases axially across each row of stationary blades. Finally, gets compressed to the desired pressure in radial stage and is discharged radially through the volute.
The present process industrial trend demands compact axial compressors with high efficiency more than 90% with good stall margin of 10 – 16 % without altering the desired discharge pressure and performance. It is also demanded that the overall cost of the compressor should not be higher. None of the prior arts discusses about the usage of suitable blade profiles, thickness distribution, axial gap between the stages, axial gap between the rows and usage of last radial stage for axial compressor which can give higher efficiency, better stall margin and reduced axial length and there by overall compressor cost. It is required to develop a high efficiency low cost multistage axial compressor with a last radial stage for handling high volume flow rates with better surge margins for process applications.
Patents related to the axial flow compressor are given below,
[1] US2726806 13/12/1955 F.H. Keast, Canada;
[2] US2006/0177302 10/08/2006 Henry Mitchell Berry, Blue Bell, PA 19422, US;
[3] US2012/0070267 13/11/2012 Yasuo Takahashi, Hitachi, Japan;
[4] US 8,308,429 B2 13/11/2012 Mark O Wakker, Rolls-Royce, London;
In Ref [1] an axial compressor is depicted with repeating stage concept and equal work in each stage. The diameter of rotor is progressively increased from first to 10th stage. The prior art Ref. [2] is related to axial compressor for use in locomotives driven by IC engine. It increases power output of IC engine by pressurizing and decreasing the volume of air at intake of the engine. The prior

art Ref. [3] is related to an axial compressor for gas turbine engine. A method to reduce leakage flow underneath the stationary blades of the machine and increased efficiency has been shown. But the method requires varying rotor diameter, complex shaft design and manufacture of small sized blades. The prior art Ref. [4] depicts use of 3 exit guide vanes for removing the swirl component of flow in a gas turbine engine. The machine shown in figure 15 has varying rotor diameter and axial discharge.
The present invention differs from the prior arts on several counts and the salient feature of the present invention vis-a-vis the prior art [1], [2], [3] and [4] are described below.
1) Prior art Ref. [1], [3] and [4] refers to axial compressor having variable diameter of rotor which progressively increases from first to last stage. Variable rotor diameter in design gives rise to complex shaft design, manufacturing complications, rotor dynamic unbalance and higher overall compressor cost. In present invention rotor hub diameter is constant from first to last stage and suitable solidity based on hub diameter is used keeping in view that number of blades, root strength and stage loading.
2) The prior art Ref. [1], [2], [3] and [4] no information is revealed about the blade profile which influences the compressor performance whereas the present invention has used the optimum blade profiles for all stages which gives surge margin up to 16% and efficiency of 92 % without increasing the axial length.
3) The prior art Ref. [1], [2], [3] and [4] no information is revealed on the solidity adapted whereas in present invention optimal solidity has been adapted for all the stages for both rotor and stators based on the hub diameter, blade length, centrifugal stress, mechanical strength, no. of blades, vibrations and rotor stability.

4) The prior art Ref. [1], [2] and [4] the air is discharged axially whereas in present invention the air is discharged radially direction after getting collected by a volute chamber to reduce the overall length of the compressor.
5) The prior art Ref. [1], [2] and [4] there is no radial stage in the machine whereas in present invention there is a radial stage at last which makes the discharged flow radial.
6) The prior art Ref. [3] makes use of 3 stator stages at exit to remove the whirl velocity from the flow where the present invention makes use of a centrifugal stage to remove the whirl velocity from discharge flow.
7) The prior art Ref. [1], [2], [3] and [4] there is no information on the axial gap between the blade rows whereas in present invention optimal axial gap has been provided between the blade rows to control the trailing edge wake zones before air enters the next stage.
8) The prior art Ref. [1], [2] and [3] there is no provision to accommodate different flow path sizes in same casing, whereas in present invention can accommodate different flow path size by using a modular inner. Same outer casing can be used for different inner casing sizes.
9) The prior art Ref. [1], [2] and [3] does not include a mechanism for reducing the thrust forces generated during compression whereas in present invention balancing drum is provided for that purpose.
10) The prior art Ref. [3] an inlet guide vane disposed upstream of an initial
rotor vane row is used for controlling an inlet flow rate where as in present
invention the inlet guide vane is used as a deflecting plate to add swirl of 20 deg,
in order to make the inlet relative flow velocity below speed of sound. Thus in the
present invention higher speeds and pressure rise is possible with the same rotor
tip diameter compared to prior art. The overall size of machine is reduced in
present invention.

SUMMARY OF THE INVENTION:
Multistage axial compressors for process applications handles high volume flow rates ranging from 1,00,000 to 5,00,000 NM3 / hr with moderate discharge pressure range of 5 to 6.5 bar with minimum 10 % of surge margin. Surge margin requirement, higher efficiency, reduced pressure losses, reduced flow separation and reduced overall cost of compressor can be achieved by appropriate blade profile geometry, solidity ratio and axial gap between stages of axial compressor blades. Axial compressors also need to have radial discharge for accommodating compressor drive and this can be achieved by inter connecting ducting between the last stage and discharge ducting. However, adapting a centrifugal stage as last stage of axial compressor eliminates the need of separate inter connecting ducting as the diffuser of the last stage serves this purpose. This also reduces the overall axial length of compressor and enable us to achieve better surge and chock margins and slightly higher discharge pressure for the same operating speed. A multistage axial compressor is developed by considering the requirements of process industries applications.
OBJECTS OF THE INVENTION:
An object of the invention is to provide a multistage axial compressor stator blades with variation in blade angle with respect to axial cord that can remove the swirl or tangential component of flow added by the rotor blades.
Another object of the invention is to provide rotor blade distribution with respect to axial cord that enables to align the blades in the relative flow direction which leads to maximum pressure rise with no flow separations from blade surfaces.
Another object of the invention is to provide optimal variable axial gap between the blade rows for all stage of compressor which removes trailing edge wakes from the flow before reaching the downstream blade resulting into better performance without increasing the overall axial length and there by overall coat of machine.

Another object of the invention is to provide optimal solidity for both rotors and stators to have sufficient flow guiding without increasing the frictional losses and without increasing the overall cost of the machine.
Another object of the invention is to provide rotor and stator blade thickness distribution along the axial chord length to have better diffusion control, better lift co-efficient and to minimize the overall pressure loss.
Another object of the invention is to provide rotor and stator blade thickness distribution along the axial chord length to meet the strength and vibration requirement.
Another object of the invention is to provide a multi stage axial compressor a tapered inner casing, outer casing, having axial inlet opening and a radial discharge forming a fluid flow path there between; a shaft, disposed axially within the tapered inner casing from the inlet to discharge; a plurality of rotor blades, mounted to the rotor within the tapered inner casing for compressing air within the tapered housing; and a plurality of stators, mounted within the tapered housing to handle large amount of volume flow in the range of 1,00,000 to 5,00,000 NM3 / hr with relative humidity ranging from 30 % to 90 % at speed of 3800 – 5000 rpm with compression ratio from 4 – 6.5 with 13 to 18 stages with compressor for process applications.
Another object of the invention is to provide multistage axial compressor with constant hub diameter for ease of manufacturing and cost reduction.
Another object of the invention is to provide multistage axial compressor with variable inner casing diameter to provide continues reduced cross sectional area from inlet to discharge to accommodate gas compression.
Another object of the invention is to provide multistage axial compressor with intermediate blow off system after half of the total number of axial stages.

Another object of the invention is to provide a multistage axial compressor with a modular outer casing where variable sizes of inner casings can be accommodated within the same outer casing.
Another object of the invention is to provide a multistage axial compressor with inlet guide vanes and speed control mechanism for precise surge control.
Another object of the invention is to provide a multistage axial compressor with a minimum surge and chock margin of 10% with appropriate compressor blade profiles for process applications.
Another object of the invention is to provide a multistage axial compressor with a centrifugal stage as last stage to improve surge and chock margins.
Another object of the invention is to provide a multistage axial compressor with a centrifugal stage as last stage to minimize the overall axial length of compressor for same pressure ration.
Another objective of the invention is to have higher pressure ratio in the last stage of axial compressor for achieving desired delivery pressure.
Another object of the invention is to provide a multistage axial compressor with inlet guide vanes and speed control mechanism for precise surge control.
BRIEF DESCRIPTION OF THE ACCOMPANYING DRAWINGS:
It is to be noted, however, that the appended drawings illustrate only typical embodiments of the present subject matter and are therefore not to be considered for limiting of its scope, for the invention may admit to other equally effective embodiments. The detailed description is described with reference to the accompanying figures. In the figures, the left-most digit (s) of a reference number identifies the figure in which the reference number first appears. The same numbers are used throughout the figures to reference like features and components. Some embodiments of system or methods in accordance with

embodiments of the present subject matter are now described, by way of example, and with reference to the accompanying figures, in which:
Figure 1 shows variation in blade angle with respect to axial cord for stators of stages 1-8 of present invention multistage axial compressor.
Figure 2 shows variation in blade angle with respect to axial cord for stators of stage no. 9 to stage no.15 of present invention multistage axial compressor.
Figure3 shows variation in blade angle with respect to axial cord for rotors of stage no.1 to stage no.8 of present invention multistage axial compressor.
Figure 4 shows variation in blade angle with respect to axial cord for rotors of stage no.9 to stage no.16 of present invention multistage axial compressor.
Figure 5 shows variation of axial gap for stage no.1 to stage no. 16 in terms of axial cord of present invention multistage axial compressor.
Figure 6 shows variation of rotors solidity for stage no.1 to stage no. 16 of present invention multistage axial compressor.
Figure 7 shows variation of stators solidity for stage no.1 to stage no. 15 of present invention multistage axial compressor.
Figure 8 shows rotor blade thickness distribution of stage no.1 to stage no. 8 of present invention multistage axial compressor.
Figure 9 shows rotor blade thickness distribution of stage no.9 to stage no. 16 of present invention multistage axial compressor.
Figure 10 shows stator blade thickness distribution of stage no.1 to stage no. 8 of present invention multistage axial compressor.
Figure 11 shows stator blade thickness distribution of stage no.9 to stage no. 15 of present invention multistage axial compressor.

Figure 12 shows velocity triangles of inlet guide vane and first rotor of prior art Ref [3] where the inlet guide vane is used as deflecting plate for flow control.
Figure 13 shows the present invention inlet guide vane used for inlet relative velocity control and as well as deflecting plate for flow control.
Figure 14 shows comparison of present invention and prior art. Flow velocity W2 at inlet to rotor blade is limited to subsonic region by the use of IGV in present invention.
Figure 15 shows layout of axial compressor for prior art Ref. [1] with variable hub and tip diameter which is difficult for manufacturing.
Figure 16 shows layout of axial compressor for prior art Ref. [2] With only axial stages used for IC engines for combustion air supply.
Figure 17 shows layout of axial compressor for prior art Ref. [3] with exit guide vanes for removing whirl velocity at compressor exit.
Figure 18 shows sealing arrangement of prior art Ref. [4] with complex rotor and non-uniform hub which requires additional power to drive additional blade at sealing.
Figure 19 shows configuration of present invention with constant hub diameter, last radial stage and simple rotor for better manufacturability and reduction axial length.
Figure 20 shows layout and details of present invention axial compressor for process application.
The figures depict embodiments of the present subject matter for the purposes of illustration only. A person skilled in the art will easily recognize from the following description that alternative embodiments of the structures and methods illustrated herein may be employed without departing from the principles of the disclosure described herein.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS:
Geometry of the invention is defined in Bezier curve by a parametric function for both the x and y variables. Two-dimensional curve is defined not by a single ordinary function like y=f(x) but by a set of two functions x=x(t) and y=y(t) of a parameter “t”. Any solid, closed object has some tangent lines or parallel lines to the principal axis, which give results of infinity in slope with the use of equation y=f(x). Thus for defining engineering curves which have arbitrary lines parallel to axes, curvatures of infinity and curvatures tending to zero, parametric curves like Bezier curves are required. The curve so generated is independent of any reference coordinate system. The Bezier curve is defined by a set of control points, the number of which determines the degree of polynomial. Changing the position of any control point produces shape change throughout the curve. The first derivative and second derivative with respect to parameter “t” of Bezier curve equation is continuous and the variation is smooth. With this curve the rate of diffusion across the blade passages is controlled and scalability of blade-profiles are better than other type of design curve. Also these curves provide smoother shape, which is easier to manufacture. Because of these properties the design of axial compressor blades has been carried out with the use of Bezier curves.
Multistage axial compressor provided with stator blades of blade angle variation with axial cord length as per the Equation-1 and Equation-2. Variation of stator blade angle with axial cord length for stage no.1 to stage no.8 of present invention is shown in Figure.1. This pattern of blade angle variation has helped in removing the swirl or tangential component of flow that is added by the rotor blades. Blade angle for stators stage no.1 to stage no.8 follows a linear pattern from inlet to exit and the rate of change in blade angle is constant.
X: - 0.32 * t^4 + 0.56 * t^3 - 0.21 * t^2 + 0.049 * t – 2 * 10^ (-4) Equation 1
Y: -t * (17 * t^3 – 34 * t^2 + 16 * t - 24) Equation 2

Multistage axial compressor provided with stator blades with blade angle variation with axial cord length as per the Equation-3 and Equation-4. Variation of stator blade angle with axial cord length for stage no.9 to stage no.15 of invention is shown in Figure.2. This pattern of blade angle variation has helped in removing the swirl or tangential component of flow that is added by the rotor blades. Blade angle for stators stage no. 9 to stage no. 15 follows a linear pattern from inlet to exit and the rate of change in blade angle is constant.
X: - 0.35 * t^4 + 0.64 * t^3 - 0.26 * t^2 + 0.04 * t + 2.7 Equation 3
Y: 122.0 * t^4 - 199.0 * t^3 + 43.0 * t^2 - 11.0 * t + 27.0 Equation 4
Multistage axial compressor provided with rotor blades of blade angle variation as per the Equation-5 and Equation-6 with axial cord length. The inlet angle of rotor blade is aligned to the relative flow direction. The flow has to be turned by the blades to transfer energy to the fluid by increasing the tangential velocity. Very fast turning of flow leads to flow separations and decreases efficiency of compressor. Blade angle variation of rotors of stage no.1 to stage no. 8 has been shown in Fig.3 which has helped to achieve maximum pressure rise with no flow separation from blade surfaces.
X: - 0.25 * t^4 + 0.48 * t^3 - 0.18 * t^2 + 0.043 * t – 1 * 10^ (-4) Equation 5
Y: 28.0 * t^4 - 48.0 * t^3 + 39.0 * t^2 + 10.0 * t + 0.094 Equation 6
Multistage axial compressor provided with rotor blades of blade angle variation as per the Equation-7 and Equation-8 with axial cord length. The inlet angle of rotor blade is aligned to the relative flow direction. The flow has to be turned by the blades to transfer energy to the fluid by increasing the tangential velocity. Very fast turning of flow leads to flow separations and decreases efficiency of compressor. Blade angle variation of rotors of stage no.9 to stage no. 16 has been shown in Fig.4 which has helped to achieve maximum pressure rise with no flow separation from blade surfaces.
X: - 0.1 * t^4 + 0.22 * t^3 - 0.085 * t^2 + 0.016 * t + 5 * 10^ (-4) Equation 7

Y: 36.0 * t^4 - 77.0 * t^3 + 46.0 * t^2 + 5.8 * t + 3.5 Equation 8
Multistage axial compressor provided with appropriate axial gap between the stages which is governed by Equation-9 and Equation-10. The variation in axial gap in % of chord length for stage no. 1 to stage no.16 has been shown in Fig.5. Axial gap is required for removing trailing edge wakes from the flow. Increase in axial gap helps in killing the trailing edge wakes before reaching the downstream of blade. However, increase in axial gap increases the overall rotor length and the overall cost of the machine. The axial gap in % of chord length of blades from stage no.1 to stage no.8 is kept constant at 15% of chord. The chord length of stages no.9 to stage no.16 decreases and a minimum axial gap is required for flow stabilization hence axial gap in % of chord is increased progressively.
X: - 0.86 * t^4 + 12 * t^3 – 20 * t^2 + 24 * t + 1 Equation 9
Y: - 0.19 * t^4 + 0.37 * t^3 - 0.15 * t^2 + 0.016 * t + 0.15 Equation 10
Multistage axial compressor rotors provided with appropriate solidity for improving efficiency. Solidity is defined as the ratio of chord length to spacing between blades. More solidity provides more flow guidance by the compressor blades but the frictional losses increases and cost of the machine increases due to more no. of blades. Increase in solidity at hub decreases the blade root width, which reduces the life and operational reliability of machine. Solidity for rotors of the invention is governed by Equation-11 and Equation-12 and variation of solidity is shown in Fig.6 for rotors of stage no.1 to stage no.16. Solidity for rotor blades are lower than stator blades due to lower perimeter of rotor boundary than casing boundary.
X: 23.0 * t^4 - 59.0 * t^3 + 49.0 * t^2 + 0.84 * t + 1.3 Equation 11
Y: 0.37 * t^4 + 0.31 * t^3 - 1.0 * t^2 - 0.19 * t + 1.6 Equation 12
Multistage axial compressor stators provided with appropriate solidity for improving efficiency. Solidity is defined as the ratio of chord length to spacing between blades. More solidity provides more flow guidance by the compressor

blades but the frictional losses increases and cost of the machine increases due to more no. of blades. Increase in solidity at hub decreases the blade root width, which reduces the life and operational reliability of machine. Solidity for stators of the present invention is governed by Equation-13 and Equation-14 and variation of solidity is shown in Fig.7 for stators of stage no.1 to stage no. 15. Solidity for rotor blades are lower than stator blades due to lower perimeter of rotor boundary than casing boundary.
X: -39 * t^4 + 68 * t^3 – 34 * t^2 + 19 * t + 1 Equation 13
Y: 0.13 * t^4 + 0.8 * t^3 - 1.0 * t^2 - 0.58 * t + 2.1 Equation 14
Multistage axial compressor provided with appropriate rotor blade thickness distribution which has impact on lift force generated and performance of the machine. As the flow passes over the aerofoil blade, lift force is generated, which contributes to increase in tangential velocity of the fluid handled. Lift coefficient is directly proportional to lift force and inversely proportional to density and square of velocity. The lift force generated by upstream stages and downstream stages is of the same order of magnitude, but the lift coefficient required by the upstream stages 1 to 8 is higher due to lower density of air in these stages. The density in stages 9 to 16 is 2 to 3 times higher than the upstream stages 1 to 8, hence the lift coefficient is higher. The lift coefficient of the thinner sections is lower than those of thicker sections at incidence angle of +/- 8 degrees as found from experimental evidences. Lift co-efficient (Cl) is defined in Equation-15.

Rotor blade thickness distribution along the axial chord length is governed by
25 Equation-16 and Equation-17 to control the diffusion and acceleration across
the blade surface in order to minimize the overall pressure loss. Further the

upstream stages have longer heights than downstream stages which requires higher cross-sectional area to meet the strength and vibration requirement. Blade thickness distribution of rotors of stages 1 to 8 is shown in Fig.8.
X: 100.0 * t Equation 16
Y: - 0.34 * t^6 + 1.1 * t^5 - 1.4 * t^4 + 0.88 * t^3 - 0.31 * t^2 + 0.07 * t + 2 * 10 ^ (-3) --Equation 17
Similarly, rotor blade thickness distribution along the axial chord length for rotors of stages 9 to 16 is governed by Equation-18 and Equation-19 to control the diffusion and acceleration across the blade surface in order to minimize the overall pressure loss. Blade thickness distribution of rotors of stages 9 to 16 is shown in Fig.9.
X: 100.0 * t Equation 18
Y: - 0.2 * t^6 + 0.67 * t^5-0.85 * t^4 + 0.52 * t^3-0.19 * t^2 + 0.042 * t+1.3 * 10 ^ (-3) --- Equation 19
Similarly, stator blade thickness distribution along the axial chord length for rotors of stages 1 to 8 is governed by Equation-20 and Equation-21 to control the diffusion and acceleration across the blade surface in order to minimize the overall pressure loss. Blade thickness distribution of stators of stages 1 to 8 is shown in Fig.10.
X: 100.0 * t Equation 20
Y: - 0.24 * t^6 + 0.83 * t^5 - 1.1 * t^4+0.66 * t^3 - 0.23 * t^2+0.054 * t+1.5 * 10 ^(-3) -- Equation 21
Similarly, stator blade thickness distribution along the axial chord length for rotors of stages 9 to 15 is governed by Equation-22 and Equation-23 to control the diffusion and acceleration across the blade surface in order to minimize the overall pressure loss. Blade thickness distribution of stators of stages 9 to 15 is shown in Fig.11.
X: 100.0 * t Equation 22

Y: - 0.13 * t^6+0.44 * t^5 - 0.58 * t^4+0.37 * t^3 - 0.14 * t^2+0.033 * t+1*10 ^ (-3) ----- Equation 23
Multistage axial compressor provided with axial entry with bell mouth 12 for air entry without pre-swirl which helps in achieving higher efficiency and a radial discharge for accommodating the drive prime mover either steam turbine or electric motor. Compressor provided with inner and outer casing configuration for better structural rigidity, low noise emission, low vibrations and for maintaining constant blade clearances at all operating conditions. Horizontal split inner and outer casing along with bearing housings eases assembly and better access for maintenance.
Multistage axial compressor provided with rotor of constant hub diameter and inner casing with variable diameter which reduces continuously from compressor inlet to exit providing flow path to continuously compressed gas as the volume flow continuously reduces from inlet to exit. Constant hub diameter and variable tip diameter eases manufacturing.
Multistage axial compressor provided with first stage as plurality of rotatable inlet guide vanes 10, spaced circumferentially, fixed inside sleeve on inner casing 3. These guide vanes will be rotated by guide vane rotation mechanism 14 consists of pinion gears that are driven by circular rack gear. Inlet guide vanes deflect the flow and adds swirl of 20 deg. in order to make the flow subsonic for downstream stages and also controls the flow rather than just acting as deflator plate in prior art. Air enters the axial compressor with a velocity “C” as shown in Fig.12. The air velocity normally has an incidence of 3 degrees at the leading edge of inlet guide vane. Prior inventions claim that at normal operation the inlet guide vane is fully open to allow maximum volume of air to pass through the machine. At this operating condition the flow velocity at trailing edge of inlet guide vane is “C1”. The rotor blade situated downstream of inlet guide vane is rotating at velocity “U”, hence it receives the flow at velocity “W1” relative to its motion. In axial compressor machine the flow velocity at tip section of first rotor

blade is the maximum and limiting factor for high speed operation. As speed of blade increases the relative velocity “W1” increases and reaches speed of sound though inlet flow velocity “C1” is not near speed sound. In present invention the inlet guide vane is set with a stagger angle such that it deflects the flow at an angle of 20 deg. and the velocity near the trailing edge is “C1”, as shown in Fig.13. The rotor blade situated downstream of inlet guide vane receives the flow at velocity “W1” which is smaller than velocity “W2” of prior art as shown in Fig.14. Thus, Inlet guide vanes deflect the flow and adds swirl of 20 deg. in order to make the flow subsonic for downstream stages. Thus in the present invention higher speeds and pressure rise is possible with the same rotor tip diameter compared to prior art. This results in reduction in tip diameter and there by size of the machine.
Multistage axial compressor provided with radial stage 11 as last stage. This configuration helps in reducing the axial length of compressor as the inlet and diffuser radial stage acts as inter-connecting ducting for changing the flow direction from axial to radial direction. In general, radial stages will have higher stall margin up to 30% of rated flow and improving the overall stability of the compressor.
Multistage axial compressor provided with intermediate blow off system 7 located after half of the total axial number of stages. This feature helps in preventing stalling during starting, stopping and emergency shutdown of compressor. Last four stages of multistage axial compressor provided with optimized identical rotor and stator blades for ease of manufacturing without sacrificing the performance and stall margins.

WE CLAIM:
1. Multistage axial compressor having compressor with axial entry and radial discharge comprising of inner casing 3, outer casing 2, multiple rows with plurality of stationary blades 5 having predefined spacing and stagger angle arranged in circumferential grove in horizontal spit inner casing without axial overlapping, multiple rows with plurality of moving blades 6 having predefined spacing and stagger angle arranged in circumferential grove in rotor without axial overlapping of blades, detachable bell mouth 12, radial discharge volute 4, rotor assembly 1 mounted on journal bearings, seal drum 13, inner casing in two to three parts, modular horizontal split outer casing to accommodate variable sizes of inner casing in which stator blades have linear blade angle variation and rotor blades with non-linear blade angle variation.
2. Multistage axial compressor as claimed in claim 1, wherein the rotor hub is provided with constant diameter and inner casing with variable diameter.
3. Multistage axial compressor as claimed in claim 1, comprising of first stage as plurality of rotatable inlet guide vanes 10, spaced circumferentially, fixed inside sleeve on inner casing 3, rotated by inlet guide rotation mechanism 14 with pinion and driven by circular rack gear.
4. Multistage axial compressor as claimed in claim 1, comprising intermediate blow off system 7 after half of the total axial number of stages.