DESC:FIELD OF THE INVENTION
The present invention relates last stage of axial compressor with high flow coefficient where the volume of air that has to be handled is high with flow co-efficient up to 0.111 for process applications.
The present invention is more particularly relates to shrouded 2D impeller centrifugal stage of axial compressor with flow coefficient ranging from 0.111 to 0.063 for process applications
BACKGROUND OF THE INVENTION
Axial compressors for process applications handles air flow rates ranging from 1,80,000 to 4,00,000 NM3 / hr and flow co-efficient of last stage ranges from 0.111 to 0.063 with moderate discharge pressure range of 6 to 7 bar. Axial compressors require either radial discharge or radial inlet for accommodating the compressor drive. This has been traditionally achieved by using an inter connecting ducting between last axial stage and radial discharge volute. Adapting inter connecting ducting slightly increases overall length of compressor and it is not possible to obtain higher pressure ratio in the last stage compared to other stages.
Prior art centrifugal stages have either 3D impeller or 2D impellers depending on the flow that has to be handled and they are generally used in multistage centrifugal compressors. The 3D impeller stages handle high flows rates with flow co-efficient ranging from 0.15 to 0.04 and has tendency of slightly increasing the overall axial length. 2D impeller stages handle relatively lesser flow rates with flow co-efficient less than 0.04 and are not suitable for handling larger flow rates that are encountered in the axial flow compressors for process applications. Hub & shroud diameters of both traditional prior at 2D & 3D impellers will not match with hub & tip diameters of the axial compressor which are typically higher. The prior art centrifugal stages receive guided flow at inlet by inlet ducting or by suction channel ducting depending on their location.
Related Prior Art:
[1] D1: US 6,340,291;
[2] D2: US 6,715,991;
[3] D3: US 7,563,074;
[4] D4: US 8,308,420;
A cylindrical blade for a rotor of radial type of impeller that is 2D in nature and typical coordinates with varying thickness from inlet to exit is reported in D1. Thickness of impeller blade is continuously increasing from inlet to exit and reaches to maximum at approximately 50% of meridional length and comes to minimum at exit. The impeller is meant for very low pressure raise in the order of 5000 Pa. Prior art D2 is 2D impeller of radial type and typical coordinates with varying radius are reported. The suction & pressure surface of the blade referred a convex & concave surface respectively. The blade coordinates are fixed and are given as ratio of the impeller exit radius.
The present invention differs from the prior arts on several counts and the efficiency achieved is higher than the existing 2D impeller stages of same flow coefficient. The salient feature of the present invention vis-a-vis the prior art D1, D2, D3 and D4 are described below.
1) Prior art D1, is a two dimensional centrifugal compressor impeller caters to air pollution control measuring instruments that are used for determining the bacterial content of the air. The present invention is related to last stage of a multi stage axial compressor for process applications to handle air.
2) Prior art D1, volume handled by the impeller is very low which is maximum of 9 m3 /hr. with head raise of 5000 Pa. whereas the present invention relatively handles large volumes of flow 85000 m3 / hr. and pressure raise will be 1.15 time to inlet pressure.
3) Prior art D1, caters to a plurality of impeller blades passing in part spiral manner from an outside circumferential edge to the bore defining impeller blade channels between adjacent impeller blades whereas the present invention has constant blasé thickness leaving the leading edge. The difference between both the geometries has been clearly vindicated in Fig.1 & Fig.2.
4) Prior art D2, caters to the two dimensional centrifugal compressor impeller where the geometry is represented in Cartesian co-ordinates in terms of impeller exit radius which is 200 mm as shown in Fig.3, whereas, the present invention geometry is in the form of Bezier polynomials and it is parameterized.
5) The prior art D2 geometry is of fixed type whereas the present invention can be scaled up and scaled down by 40% in diameter to produce other sizes of impellers with same high aerodynamic efficiency.
6) Prior art D3, caters to centrifugal compressor impeller of three dimensional twisted blade profiles as shown in Fig. 4 whereas the present invention has two dimensional blades of constant thickness. Comparing the geometries of D3 and present invention from Fig.2 and Fig.3, it is clearly vindicated that D3 and present inventions are entirely different.
7) The prior art D4, though it is of the same interest of improving the aerodynamic performance of centrifugal impeller, the methodology followed and the final geometry achieved in both the cases are entirely different. In the prior art D4, is of three dimensional blade impeller and the geometry is not defined neither in terms of coordinates nor in the parameterized form. Also the blade angle distribution, wrap angle distribution at various sections of blade from hub to shroud are not mentioned. In case of present invention, the geometry is clearly defined in the form of blade angle distribution, wrap angle distribution, hub & shroud contour slopes, and impeller passage area distribution. This is clearly vindicated from Fig. 5 & Fig. 6 where both the cases are distinguished in terms of impeller blade angle distribution from inlet to exit that completely defines the blade shape.
The present process industrial trend demands compact axial compressors with good stall margins without sacrificing the desired discharge pressure & performance. None of the prior arts discusses about the 2D impeller stages that are suitable for axial compressor. It is required to develop a 2D impeller centrifugal compressor stage for handling high volume flow rates that are encountered in axial compressors for process applications.
Every multi stage axial compressors for process applications will handle high flow rates with moderate discharge pressure range of 6 to 7 bar form ambient inlet pressure. The discharge of the air must be radial to accommodate the compressor drive. In case of axial compressors without centrifugal stage as last stage necessitates use of volute and inter connecting ducting which increases overall axial length of compressor. Air gets compressed gradually from inlet to exit of the axial compressor and the volume of air to be handled reduces at the last stages. Hence, it is possible to use a tailor made centrifugal stage as a last stage of axial compressor. Adapting a 2D impeller centrifugal stage as last for axial compressor will help in achieving radial discharge without increasing overall axial length with improved stall margins and increased discharge pressure.
OBJECTS OF THE INVENTION
The main objective of the invention is to develop a centrifugal stage that can be used as last stage of an axial compressor to have radial discharge minimizing the overall axial length of compressor & improving the rotor dynamic stability of compressor.
Another objective of invention is to improve the surge & chock margins of axial compressor desired for process application.
Another objective of invention is to have higher pressure ratio in the last stage of axial compressor for achieving desired delivery pressure.
SUMMARY OF THE INVENTION
Axial compressors for process applications handles high volume flow rates ranging from 1,80,000 to 4,00,000 NM3/ hr with moderate discharge pressure range of 6 to 7.0 bar. Axial compressors 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 enables us to achieve better surge & chock margins and slightly higher discharge pressure for the same operating speed. A centrifugal stage is being developed by considering the requirements of axial compressor for process applications.
BRIED DESCRIPTION OF DRAWINGS
The proposed invention will be better understood by the following description with reference to the accompanying drawings:
Figure.1: Impeller of prior art D1 with variable impeller blade thickness
Figure.2: Impeller of present invention with constant impeller blade thickness leaving the leading edge distinguishing from the prior art.
Figure.3: Impeller blade of prior art D2 where the 2D impeller is defined in terms Cartesian ordinates and is fixed type not scalable and other information not furnished.
Figure.4: Impeller of prior art Ref. [3] which is 3D in nature distinguishing from present invention.
Figure.5: Impeller blade angle distribution of prior art Ref. [4]
Figure.6: Impeller blade angle distribution of present invention distinguishing from prior art.
Figure.7: Impeller blade wrap angle “?“ distribution of present invention from impeller inlet to exit.
Figure.8: Impeller passage area distribution of present invention.
Figure.9: Impeller cross-section of present invention
Fig. 10: Impeller blade curvature distribution of present invention
Fig. 11: Impeller blade slope distribution of present invention
DETAILED DESCRIPTION
The present invention is described in detail below.
Every multi stage axial compressors for process applications will handle high flow rates with moderate discharge pressure range of 6 to 7 bar form ambient inlet pressure. The discharge of the air must be radial to accommodate the compressor drive. In case of axial compressors without centrifugal stage as last stage necessitates use of volute and inter connecting ducting which increases overall axial length of compressor. Air gets compressed gradually from inlet to exit of the axial compressor and the volume of air to be handled reduces at the last stages. Hence, it is possible to use a tailor made centrifugal stage as a last stage of axial compressor. Adapting a 2D impeller centrifugal stage as last for axial compressor will help in achieving radial discharge without increasing overall axial length with improved stall margins and increased discharge pressure.
In case of present invention, the impeller of 2Dcentrifugal stage receives inlet flow directly from an axial stage. This makes the present invention very unique in expression to meet various requirements like hub &shroud diameters to match with axial stage and are constrained by induced stresses & material strength with operating speed same as axial stages as it is mounted on the same shaft. Also, the present invention needs to accommodate balancing drum and various sealing arrangements unlike general centrifugal stages. The range of the present invention flow co-efficient is 0.111 to 0.063.
Axial compressor where the present invention will be adapted as last stage, the fluid enters in axial direction and leaves in the radial direction. Axial 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 usually in pairs, one rotating and one stationary attached to a stationary tubular casing. A pair of rotating and stationary airfoils is called a stage. The rotating airfoils, also known as blades or rotors, accelerate the gas. The stationary airfoils, also known as stators or vanes, convert the increased rotational kinetic energy into static pressure through diffusion and redirect the flow direction of the gas, preparing it for the rotor blades of the next stage. The cross-sectional area between rotor drum and casing is reduced in the flow direction to maintain an optimum Mach number using variable geometry as the gas is compressed.
The main objects of invention are parameterised impeller geometry and their definitions that control the entire impeller geometry are described in the following.
a) Impeller blade angle “ß” distribution at impeller hub & shroud surfaces along the impeller inlet to impeller exit is defined in Fig. 6. Blade angle distribution at hub & shroud are slightly different in the case of present invention unlike traditional 2D impellers. This blade angle distribution completely controls the flow physics within the impeller like diffusion, pressure recovery coefficient, pressure loss coefficient, relative velocity distribution, impeller passage area distribution, flow pattern at impeller exit and along the downstream of the compressor stage and thereby efficiency. The impeller blade angle variation along the percentage length of meridional flow path “B” of the present invention is shown in Fig.6 and it is governed by Equation.1 & Equation.2 at hub & shroud surfaces respectively.
ß = (-) 39.9999 + 0.0928432 * B - 0.000816585 * B 2 + 4.29248*10 -6 * B 3 - 1.80253*10 -8 * B 4 + 3.91547*10 -11 B 5 ------ Equation.1

ß = (-) 35.6243 - 0.271364 B + 0.0137778 B 2 - 0.00028241 B 3 + 2.62199*10 6 * B 4 - 9.09088*10 -9 * B 5 ------ Equation.2
b) Impeller wrap angle “?” distribution at impeller hub and shroud surfaces from impeller inlet to impeller exit influences blade loading / relative velocity distribution, the pressure recovery, flow behaviour and thereby efficiency. The present invention has the wrap angle distribution which increases uniformly from impeller inlet to impeller exit for achieving higher efficiency as in Fig.7. Numerically the wrap angle variation is (-) 74 deg. to (-) 90 deg. for hub and (-) 77 deg. to (-) 90 deg. for shroud with respect to meridional plane. The impeller wrap angle variation of the present invention along percentage length of meridional flow path “A” is shown in Fig. 7 and it is governed by Equation.3 & Equation.4 at hub & shroud surfaces respectively.
? = 17.0668 - 0.228455 A + 0.000907139 * A2 - 4.9019 * 10 -6 * A 3 + 2.05806 * 10 -8 * A 4 - 4.48947 * 10 – 11 * A 5 ---- Equation.3

? = 12.9227 - 0.150115 A + 1.40733 * 10 -6 * A 2 + 6.5491 * 10 -6 A 3 - 6.92684 * 10 -8 * A 4 + 2.455 * 10 -10 * A 5 ---- Equation.4
c) Impeller blades circumferential pitch in terms of degrees depends of the total number of blade in the impeller. The number of blades in the impeller shows major impact on the peak blade loading and total frictional losses. It is observed that low flow stages requires lesser number of impeller blades whereas high flow stages demands more number of impeller blades to achieve proper blade loading which results in efficient flow behaviour in the entire compressor stage. For the present invention which is a low flow impeller, the circumferential pitch is 17.14 deg. for achieving better blade loading and lesser friction losses. This has ultimately resulted higher efficiency of compressor.
d) Impeller inlet hub diameter ?D1 h” shown in Fig.9,influences the efficiency of compressor stage. Higher inlet hub diameter demands higher inlet shroud diameter which increases the inlet relative velocity at shroud. Increase in inlet shroud velocity reduces the impeller diffusion and there by pressure recovery and impeller efficiency. Inlet hub diameter also influences the multi stage axial compressor shaft diameter and compressor rotor dynamic behaviour. This diameter need to be compatible with the upstream axial stage geometry. Considering above all requirements, for the present invention which is of 1080 mm impeller diameter with flow range of 84% to 130% of design flow, the inlet hub diameter is 49% to 51% of impeller exit diameter.
e) Impeller inlet shroud diameter ?D1 s” shown in Fig.9, depend on the upsteam radial stage geometry and rotor dynamic requirements. This is basically controlled by operating range of compressor stage as reducing the inlet radius drastically will result in chocking at higher flow. Inlet shroud radius of the present invention that has enabled to achieve higher compressor efficiency is 68 % to 70 % of impeller exit diameter for the flow range of 84% to 130% of design flow.
f) Impeller blade width at exit ?B2” shown in Fig.9, plays major role in achieving the overall efficiency of corresponding compressor stage. Impeller exit width influences diffusion, flow associated problems like re-circulation, separation and low momentum zones. Impeller exit blade width is directly related to impeller exit blade angle ?ß2b” and flow coefficient. Increasing the impeller width demands higher pinching in the diffuser width in order to avoid flow associated problems like recirculation, separation and low momentum zones in the diffuser and further downstream. Diffuser pinching in the width more than 25% can cause flow disturbance at the diffuser inlet. For the present invention with 1080 mm impeller exit diameter where the flow co-efficient is high with the impeller exit blade angle (-) 64 deg. the impeller exit blade width is 0.051% of impeller diameter.
g) Impeller passage area distribution is mainly responsible for diffusion with in the impeller and influences the compressor stage efficiency. This passage area distribution varies from high flow to low flow stages. Continuous increase in the impeller passage area from impeller inlet to exit along the meridional flow path ensures conversion of kinetic energy in to pressure and also eliminates recirculation zones with in the impeller. The impeller passage area distribution for achieving the higher efficiency for present invention which is of high flow 2D impeller is shown in Fig. 8. The impeller passage area in mm2 “PA” distribution along the percentage length of meridional flow path “A” is governed by Equation.5.
PA = 0.170451 - 0.00516579 A + 0.000191715 A 2 - 3.07077 * 10 -6 * A 3 +
2.12972 * 10 -8 * A 4 - 4.57832 * 10 -11 * A 5 --------- Equation. 5
h) Impeller blade curvatures at hub and shroud contours influences the blade loading, relative velocity distribution, pressure rise and there by efficiency. Variation in curvature that has resulted higher efficiency in the present invention at hub and shroud surfaces are shown in Fig.10. Impeller blade curvature “C” is constant at hub surface and impeller slope variation at shroud section along the percentage meridional length “A” is governed by Equation. 6.
C = 0.0278949 - 0.00276621 * A + 0.000111707 * A 2 - 2.16461 * 10 -6 * A 3 + 1.98963 * 10 -8 * A 4 - 6.92669 * 10 -11 * A 5 --------- Equation. 6
i) Impeller hub and shroud disc contours slopes influences the blade loading, relative velocity distribution, pressure rise and there by efficiency. Variation in slope with respect to meridional plane at hub & shroud surface is shown in Fig.11. The present invention has constant slope at hub surface and variation in the shroud slope along the percentage length of meridional flow path “A” is governed by Equation.7.

S = 63.9684 + 1.88932 A - 0.0715731 * A 2 + 0.00134305 * A 3 - 0.0000120455 * A 4 + 4.14829 * 10 -8 * A 5 --------- Equation. 7
In the present invention, impeller geometrical parameters along the meridional flow path from impeller inlet to impeller exit like Blade angle, Wrap angle, Passage area, Inlet hub radius, Inlet shroud radius, Impeller blade exit width, Impeller blade circumferential pitch are finalised based on systematic design approach with the rich experience of compressor design & extensive CFD analysis and compressor stage performance testing of prototype where the efficiency improvement achieved is very close to the theoretical efficiency improvement.
In case of prior art, no information is available regarding the blade angle distribution, wrap angle distribution, slope distribution, curvature distribution and passage area distribution from impeller inlet to impeller exit is not defined. In case of prior art where blade angle distribution is defined is entirely different from present invention. Also no information is available regarding the impeller inlet hub radius and shroud radius. These are the major factors influenced the relative velocity distribution, blade loading, diffusion, static pressure distribution, flow behaviour within the impeller and further downstream which ultimately results in higher efficiency of compressor stage. The rigorous CFD studies have revealed that scaling of impeller up to 40% upward and 40% downward can give the same performance while the ratios of other impeller geometrical parameter are maintained.
,CLAIMS:WE CLAIM:
1. A centrifugal stage with 2D impeller for axial compressor with impeller blade angle distribution at inlet and exit of impeller characterized in that the impeller blade angle distribution at inlet is (-) 60 deg. and impeller blade angle distribution at outlet is (-) 64 deg. with maximum blade angle of (-) 64 deg. at impeller exit with impeller blade angle distributions for hub & shroud surfaces.

2. The Centrifugal stage with 2D impeller for axial compressor as claimed in claim 1, wherein the impeller blade wrap angle distribution at inlet & exit of the impeller is (-) 74 deg. to (-) 90 deg. for hub and (-) 77 deg. to (-) 90 deg. for shroud with respect to meridional plane.

3. The Centrifugal stage with 2D impeller for axial compressor as claimed in claim 2, wherein the impeller passage area continuously increases from inlet to exit.

4. The Centrifugal stage with 2D impeller for axial compressor as claimed in claim 3, wherein the impeller diameter is of 1080 mm and impeller exit width is of 55 mm with impeller scalability is of 40 % in diameter both in upper and lower side of diameter.

5. The Centrifugal stage with 2D impeller for axial compressor as claimed in claim 4, with impeller outer diameter ranging from 300 mm to 700 mm with inlet hub diameter as 49% to 51% of impeller outer diameter and inlet shroud diameter as 68 % to 70 % of impeller outer diameter.

6. The Centrifugal stage with 2D impeller for axial compressor as claimed in claim 5, with outer impeller diameter ranging from 650 mm to 1500 mm with exit blade width as 0.051 % of impeller exit diameter.

7. The Centrifugal stage with 2D impeller for axial compressor according to claim 5, where in the impeller blades circumferential pitch is 17.14 deg.