FIELD OF INVENTION
The present invention relates to a two dimensional impeller stage for centrifugal compressor with optimum efficiency and operating range.
BACKGROUND OF THE INVENTION
Centrifugal compressors are used to compress any gas medium to a higher discharge pressure as per the process requirement. Compressor is made of multiple impeller stages.
A standard impeller stage consists of an impeller (a rotating element) and a diaphragm (a static element). The diaphragms form the static gas flow path channels inside the compressor. The intermediate diaphragm between two impellers performs the double task of forming the diffuser and return channel. Diffuser converts kinetic energy of gas delivered by the preceding impeller to pressure energy and the return channel in the diaphragm helps to guide the gas to the inlet of next impeller. The diffusers are of free vortex type.
Impeller stages are characterised based upon non dimensional parameters viz. Flow Coefficient (Fl), Work Coefficient (TAU), Mach Number (Mu), Polytrophic Efficiency (ETAp) and Pressure Ratio (Pr). Different standard impeller stages with different blade

geometries are available for different Flow coefficients. These stages have fixed operating range and efficiency.
It is observed that these standard impeller stages are short falling, in catering to the current market requirements like:
1. As a part of energy saving scheme, most of the fertilizer production plants are looking for low cost energy per Tonne of Urea production.
2. In refineries, Fluid Catalytic Cracker Units operate with wide range of molecular weight variation and flow variation.
3. In refineries, Delayed Cooker Units have limitation on discharge temperature requirement from process point of view (polymerisation will happen above certain temperature).
Patents related to the centrifugal compressor stage are given below, [1] CN201925238U 2011-08-10, High-efficiency closed small-flow module stage [2] US6715991B2 2004-04-06, Rotor blade for centrifugal compressor with a medium flow coefficient.
[3] US 7563074B2 2009-07-21, Impeller for centrifugal compressor [4] US 6340291B1 2002-01-22, High pressure impeller with high efficiency for small volume flows for radial blowers of different size.

The difference between the present invention and prior arts are as below:
1. Prior art Ref. [1], refers to the small flow impeller module with fixed outlet angle
as 18.30 and tapered vaneless diffuser, whereas the present invention provides the
optimum range for blade outlet angle for a given flow coefficient and the vaneless diffuser is straight. The differences in the stage are indicated in Fig-1(a) and Fig-2(a) w.r.t diffuser clearly.
2. Prior art Ref. [2], refers to the rotor blade for medium flow coefficient, wherein the impeller is constructed by 2D vane having number of vanes as 17 and outer radius of 200mm, whereas the present invention gives the number of blades as a function of flow coefficient.
3. Prior art Ref. [2], refers to the rotor blade for medium flow coefficient, where in the impeller blade is constructed by number of discrete points which are function of outer radius, whereas the blade of the present invention is constructed with two or more circular arcs such that the flow analysis criteria mentioned in Eq-9 to Eq-11 are met.
4. Prior art Ref. [2], refers constant the blade contour for different flow coefficients, whereas the blade contour of present invention varies with flow coefficient, as inlet and outlet blade angles vary with flow coefficient.

5. The prior art Ref. [2] geometry is of fixed type whereas the present invention can be used for diameters from 250mm to 1000mm to produce other sizes of impellers with same high aerodynamic efficiency.
6. Prior art Ref. [3], caters to centrifugal compressor impeller of three dimensional twisted blade profiles as shown in Fig.3 whereas the present invention has two dimensional blades of constant thickness. Comparing the geometries of prior art [3] and present invention from Fig.3 and Fig.6, it is clearly vindicated that prior art [3] and present inventions are entirely different.
7. Prior art Ref. [4], 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 multi stage compressor for industrial applications to handle various gas mixtures including air.
8. Prior art Ref. [4], 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 blade thickness leaving the leading edge. The difference between both the geometries has been clearly vindicated in Fig.4 & Fig.6.

OBJECTS OF THE INVENTION
Therefore, it is an object of the invention to propose a two dimensional impeller stage for centrifugal compressor with optimum efficiency and operating range, which is capable of meeting various requirements of the market with higher efficiencies.
Another object of the invention is to propose a two dimensional impeller stage for centrifugal compressor with optimum efficiency and operating range, which is capable of improving operating range for impeller stages.
BRIEF DESCRIPTION OF THE ACCOMPANYING DRAWINGS
The advantages of the present invention will become apparent from the following detailed description of embodiments with reference to the accompanying drawings, in which:
Figure 1: Prior-art Impeller Stage view prior art Ref. [1] which is having tapered vaneless diffuser distinguishing from present invention which has straight vaneless diffuser.
Figure 2: Prior-art Impeller blade view prior art Ref. [2] which is having standard blade contour made of coordinates function of impeller outer diameter and other information is not provided.
Figure 3: Impeller of prior art Ref. [3] which is 3D in nature distinguishing from present invention.

Figure 4: Impeller of prior art Ref [4] with variable impeller blade thickness, whereas thickness of present invention is constant.
Figure 5: Impeller stage meridional view of present invention.
Figure 6: Impeller blade curvature of present invention.
Figure 7: Impeller blade angle (meridional) distribution of present invention.
Figure 8: Impeller lean angle (meridional) distribution of present invention.
Figure 9: Impeller blade exit angle (tangential) of present invention.
Figure 10: Number of Impeller blades of present invention.
DETAILED DESCRIPTION OF A PREFERRED EMBODIMENT OF THE INVENTION
Since the existing impeller stage is not providing the efficiency and operating range as required by market demand, the dimensions of the impeller stage on the basis of research work and experiment, such as diameters, widths, no. of vanes, radii of vanes, blade angles have been optimised to achieve higher efficiency. Improvement of the impeller stage means improving its efficiency, thereby reducing power required to achieve the desired polytrophic head.
The parameters that influence the efficiency of impeller are identified and worked to bring an improvement. The parameters are as follows:

a. Impeller Incident angle: The value is fixed to reduce the incidence losses and
also to improve the operating range.
b. Blade Exit angle: To reduce the exit energy loss from the impeller.
c. No. of impeller vanes: To achieve the required discharge pressure in the flow
path and also to give necessary mechanical strength to the impeller.
d. Secondary Mass flow ratio: This defines the boundary layer thickness. During
operation, velocity of the gas near the vanes reduces due to friction losses. This
parameter decides how much boundary layer is allowed so as to reduce the
losses.
e. Impeller exit deviation angle: This parameter is the difference between blade
angle and gas flow angle at the exit due to the rotation of the impeller. This
needs to be optimised so as to reduce the exit losses.
f. Diffuser stall margin: This is the difference between the gas flow angle at
diffuser outlet and the angle at which there will not be any further increase in
the polytrophic head. This parameter decides the allowable range of operation
more than the design flow rate.

g. Return Channel Inlet incidence angle: Return channel receives the gas from diffuser and this angle is to be optimised so that the gas flows smoothly over the return channel vanes without much pressure loss.
Eq.1 to Eq.7 are finalised to achieve the loss minimisation, thus improving the efficiency of impeller stage and its operating range.
Eq.8 to Eq.11 validate the individual parameters of Eq.1 to Eq.7 during the continuous iteration process of designing the impeller blade profiles.
Conventionally two dimensional impeller stages are designed, developed and standardised for a certain number of Flow Coefficients. These standardised impeller stages are used for construction of centrifugal compressor meeting the process requirement.
This standardisation of two dimensional impeller stages does not provide optimum performance of compressors for the current market process requirements.
In the present invention, for a given requirement, the Mean Line design of two dimensional impeller stage meeting following criteria is established:


Multi Stream Tube Flow Analysis is carried out on the impeller stage and Shroud, Hub & Blade Angle distribution of the impeller meeting following criteria are established.


To achieve above flow analysis requirements, the impeller is constructed as mentioned below.
1. The impeller camber line is constructed with two or more circular arcs.
2. The impeller shroud and hub Contours are constructed with line and two or more circular arcs.
3. The impeller leading edge from Hub to Shroud Surface is either horizontal or inclined line.
DETAILS OF EQUATIONS TO ACHIEVE PARAMETERS FOR OPTIMUM EFFICIENCY.
Eq. 5 gives the criteria for designing the impeller outlet width. With changing the impeller outlet width, the deviation of fluid flow w.r.t the blade angle varies. Thus the impeller outlet width shall be such that the Eq.5. Criteria met.

Eq. 6 gives the criteria for designing the diffuser outlet diameter and width. As the outlet diameter and width varies, the operating range of the impeller varies. In order to have the improved operating range the diffuser stall margin shall be as per Eq. 6.
Eq. 7 gives the criteria for the design of return channel blade inlet angle. As the return channel inlet diameter and width are a function of diffuser outlet diameter and width respectively, the return channel inlet angle shall be such that the Eq.7 is met.
Flow coefficient is non dimensional parameter and it is given by following formula

Operating Range is defined as difference between minimum (surge) flow required and maximum (choke) flow that can be handled by impeller.
This operating Range is a function of design flow of impeller and impeller blade angles. As the impeller is designed to the specific requirement instead using of the standard discrete geometry, the operating range is improved.

The parameters indicated in Eq.1 to Eq.11 - ensures the impeller to have higher operating range with optimum efficiency. The shroud and hub contours (Fig.2a) and the impeller blade profiles are varied to have higher efficiency compared to the existing impellers.
In Eq.2, β is the impeller outlet angle and F 1 is the impeller flow co-efficient. Generally impellers are designated based on the blade outlet angle. In the present invention, the optimum impeller blade outlet angle is given as function of flow coefficient. A range of ± 30 is provided for meeting the different discharge pressure requirements.
Relating to Eq. 3, number of impeller vanes are selected such that blade loading is minimum. Blade loading is function of pressure difference across pressure surface and suction surface. (This refers fig. 2b). Eq. 3 provides number of impeller vanes as a function of flow coefficient to have minimum blade loading.
Eq.4 provides the relationship between the secondary mass flow as a function of flow coefficient secondary mass flow is defined as the ratio of zero momentum mass flow at impeller discharge w.r.t the total mass at the impeller discharge.
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 exit blade angle (tangential) “β2” governs the pressure rise of gas in the impeller and flow along the downstream of impeller and there by efficiency. The exit blade angle as a function of flow coefficient (FI) of the present invention is governed by Equation1 and is shown in Fig.9. The tangential blade angle is equal to meridional angle plus 90°

b) Number of impeller blades “N” plays a major role in transforming the kinetic energy of impeller into pressure energy of the gas fluid. The higher the number of blades leads to higher wetting losses thus lower efficiency. The less number of blades leads to improper energy transfer to gas fluid and recirculation of gas with in the flow passage. The number of blades also governs the mechanical strength of impeller. The number of blades as a function of flow coefficient (FI) of the present invention is governed by Equation2 and is shown in Fig.10.

Number of impeller blades is selected as 13 for flow co-efficient of 0.009.
c) Impeller blade width at the inlet “B1” shown in Fig. 5, influences the operating range of compressor stage. Higher the width leads to higher choke margin, however the surge margin decreases considerably. Lower the width, surge margin improves compromising the choke margin. The impeller inlet width also controls the impeller inlet

incidence angle which is defined the gas flow angle minus impeller blade angle, from hub to shroud. The higher incidence angle the higher the incidence losses and lower the impeller efficiency. Considering the above, impeller blade width of the present invention that, has enabled to achieve higher efficiency, is 130% to 150% of impeller exit width (B2).
d) Diffuser width “B3” as shown in Fig.5, governs the stall margin of compressor stage. Higher diffuser width results in the lower wetting losses thereby increasing the efficiency of impeller stage. However higher width leads to internal recirculation of flow and thus reducing the overall operating range of compressor stage. Considering the effect of diffuser width on efficiency and operating range, diffuser width (B3) of the present invention that, has enabled to achieve higher efficiency with higher operating range, is 100% to 60% of impeller exit width.
e) Impeller blade camber line curvature shown in Fig.6, defines the shape of the blade. The camber line construction controls the manufacturing process to be adopted for manufacturing of impeller by CNC machining or conventional machining methods. The curvature of present invention is made up of two or more circular arcs that has enabled to achieve the desired efficiency with conventional machining methods.

f) Impeller blade angle “β” distribution along the impeller inlet to impeller exit is
defined in Fig. 6. This blade angle distribution completely controls the flow physics
within the impeller like diffusion, relative velocity distribution, impeller passage area
distribution, flow pattern at impeller exit and along the downstream of the compressor
stage and thereby efficiency and operating range. The impeller blade angle variation
along the percentage length of meridional flow path “M” of the present invention is
shown in Fig.7 and it is governed by Equation.3. β1 & β2 are impeller inlet and outlet
blade angles (meridional).
β= β1+ (3E-09*M5 - 8E-07*M4 + 5E-05*M3 - 0.0005*M2 + 0.0987*M + 0.0027)*(β2-β1)-Eq3. β is maintained at 200 for flow co-efficient of 0.009.
g) Impeller lean angle “γ” distribution along the impeller inlet to impeller exit is
defined in Fig.7. This lean angle distribution controls the profile of the blade along the
plane perpendicular to XY axis (Fig 2). The lean angle plays a major role in the
distribution of fluid within impeller flow passage from hub to shroud.By varying lean
angle the blade loading due to the gas flow can be adjusted and thus reducing the
stress on the impeller blades. The impeller lean angle variation along the percentage

length of meridional flow path “M” of the present invention is shown in Fig.8 and it is governed by Equation.4. β2 is impeller outlet blade angle (meridional).
γ= (90+β2-7)/100*M – (90+β2-7) ---Eq4.
Impeller lean angle ‘γ’ is maintained at 130 for flow co-efficient of 0.009 at impeller inlet.

WE CLAIM

1. A two dimensional impeller stage for centrifugal compressor with optimum
efficiency and operating range comprising with following parameters;
i) impeller inlet incidence angle maintained at less than 10 at hub and less than 60
at shroud, when impeller blade width maintained at 130% to 150% of impeller exit
width for achieving higher efficiency;
ii) impeller blade exit angle (tangential) (β) maintained at 200 for flow coefficient of
0.009, for better governing the pressure rise of gas in the impeller and flow along the
downstream of impeller in optimum way;
iii) impeller exit deviation angle maintained less than 30;
iv) impeller exit relative velocities at pressure and suction surfaces near hub and
shroud locations maintained within 1% variation;
v) impeller blade angle distribution contour is having single peak;
vi) impeller lean angle distribution contour is linear and maintained at -130 for flow
coefficient of 0.009 at impeller inlet;
vii) impeller vane, shroud and hub contours constructed with two or more circular
areas; wherein,
diffuser stall margin is maintained greater than or equal to 160 and return channel inlet incidence angle is maintained at less than 20 wherein vaneless diffuser is maintained straight and diffuser width (B3) is maintained as 100% to 60% of impeller

exit width for better governing the stall margin of compressor stage and for lowering wetting losses resulting increased efficiency wherein the number of impeller blades (N) selected as 13 for flow coefficient of 0.009 for optimal transformation of kinetic energy of impeller into pressure energy of the gas fluid to obtain improved efficiency.